JP3885697B2 - Control device for spark ignition engine - Google Patents

Description

【００６３】
吸入空気量制御手段４３は、アクチュエータ１８を制御することによりスロットル弁１７の開度（スロットル開度）を制御するものであり、運転状態に応じてマップ等から目標吸入空気量を求め、その目標吸入空気量に応じてスロットル開度を制御する。この場合、特殊運転モードでは、後述のように後続気筒（２番、３番気筒２Ｂ，２Ｃ）においては分岐吸気通路１６からの吸気導入が遮断された状態で先行気筒から導入されるガス中の過剰空気と新たに供給される燃料との比が理論空燃比（以下これを実質的な理論空燃比という）とされつつ燃焼が行われるので、先行、後続の２気筒分の要求トルクに応じた燃料の燃焼に必要な量の空気（２気筒分の燃料の量に対して理論空燃比となる量の空気）が先行気筒（１番、４番気筒２Ａ，２Ｄ）に供給されるように、スロットル開度が調節される。
【００６４】
燃焼制御手段４４は、燃料噴射制御手段４５と点火制御手段４６とからなっており、燃料噴射制御手段４５により、各気筒２Ａ〜２Ｄに設けられた燃料噴射弁９からの燃料噴射量及び噴射タイミングをエンジンの運転状態に応じて制御するとともに、点火制御手段４６により運転状態に応じた点火時期の制御及び点火停止等の制御を行う。そして、特に運転状態が特殊運転モードである場合と通常運転モードである場合とで燃焼の制御（燃料噴射の制御及び点火の制御）が変更される。
【００６５】
すなわち、特殊運転モードの場合、先行気筒（１番、４番気筒２Ａ，２Ｄ）に対しては、空燃比を理論空燃比よりも大きいリーン空燃比、好ましくは理論空燃比の略２倍もしくはそれ以上とするように燃料噴射量を制御するとともに、圧縮行程で燃料を噴射して混合気の成層化を行わせるように噴射タイミングを設定し、かつ、圧縮上死点付近で強制点火を行わせるように点火タイミングを設定する。一方、後続気筒（２番、３番気筒２Ｂ，２Ｃ）に対しては、先行気筒から導入されたリーン空燃比の既燃ガスに対して燃料を供給し、実質的な理論空燃比となるように燃料噴射量を制御するとともに、吸気行程で燃料を噴射するように噴射タイミングを設定する。そして運転状態に応じて圧縮着火または強制点火による燃焼を行う。
【００６６】
通常運転モードの場合には、各気筒２Ａ〜２Ｄの空燃比を理論空燃比もしくはそれ以下とするように燃料噴射量を制御し、例えば通常運転モードのうちの大部分の領域において理論空燃比とし、全開負荷及びその付近の運転領域で理論空燃比よりリッチとする。そして、この場合に、各気筒２Ａ〜２Ｄに対して吸気行程で燃料を噴射して混合気を均一化するように噴射タイミングを設定し、かつ、各気筒２Ａ〜２Ｄとも強制点火を行わせるようにする。
【００６７】
カム位相制御手段４９は、運転状態判別手段４１の結果に基き、カム位相可変機構３３ａ，３４ａの制御を行う。制御の詳細は後述するが、例えば特殊運転モードにおいて、比較的低負荷低回転の領域（図４の領域Ａ１）ではカム２７の位相を進ませる側にカム位相可変機構３４ａを制御し、カムシャフト３４の回転によって作動する既燃ガス排出弁３２ｂ、既燃ガス導入弁３１ｂおよび後続気筒排気弁３２の開閉時期が全体的に早期になるように設定する。一方、比較的高負荷高回転の領域（図４の領域Ａ２）ではカム２７の位相を遅らせる側にカム位相可変機構３４ａを制御し、カムシャフト３４の回転によって作動する既燃ガス排出弁３２ｂ、既燃ガス導入弁３１ｂおよび後続気筒排気弁３２の開閉時期が全体的に遅れるよう設定する。なお、カム位相可変機構３３ａ，３４ａは作動中の弁に対して作用するので、弁停止機構３５によって停止状態となっている弁は、カム位相可変機構３３ａ，３４ａの制御にかかわらず停止状態を維持する。
【００６８】
以上のような当実施形態の装置の作用を、図５〜図９を参照しつつ説明する。
特殊運転モードでは前述のように先行気筒排気弁３２ａ及び後続気筒吸気弁３１ａが停止状態、既燃ガス排出弁３２ｂ及び既燃ガス導入弁３１ｂが作動状態とされることにより、実質的な新気及びガスの流通経路は図６に示すようになり、先行気筒（１番，４番気筒）２Ａ，２Ｄから排出される既燃ガスがそのまま気筒間ガス通路２２を介して後続気筒（２番，３番気筒）２Ｂ，２Ｃに導入されるとともに、この後続気筒２Ｂ，２Ｃから排出されるガスのみが排気通路２０に導かれるような２気筒接続状態とされる。
【００６９】
この状態において、先行気筒２Ａ，２Ｄにそれぞれ吸気行程で吸気通路１５から新気が導入され（図６中の矢印ａ）、先行気筒２Ａ，２ＤではリニアＯ₂センサ２５により検出される空燃比が理論空燃比の略２倍ないしそれ以上の超リーン空燃比となるように燃料噴射量がフィードバック制御されつつ圧縮行程で燃料が噴射され、かつ、所定点火時期に点火が行われて、超リーン空燃比での成層燃焼が行われる（図５参照）。
【００７０】
その後、先行気筒２Ａ，２Ｄの吸気行程と後続気筒２Ｂ，２Ｃの排気行程が重なる期間に、先行気筒２Ａ，２Ｄから排出された既燃ガスがガス通路２２を通って後続気筒２Ｂ，２Ｃに導入される（図５中の白抜き矢印及び図６中の矢印ｂ）。そして、後続気筒２Ｂ，２Ｃでは、先行気筒２Ａ，２Ｄから導入されたリーン空燃比の既燃ガスに燃料が供給されて、実質的な理論空燃比となるように燃料噴射量が制御されつつ、吸気行程で燃料が噴射される。このとき、後続気筒２Ｂ，２Ｃでは点火プラグ７での強制点火が停止され、圧縮行程の上死点付近で燃焼室内の圧力、温度の上昇により圧縮着火が行われる。
【００７１】
こうして後続気筒２Ｂ，２Ｃでは、多量のＥＧＲガス相当の既燃ガス成分を含み、かつ、空燃比がリーンであるという条件下でも、同時圧縮着火により燃焼が急速に行われ、これにより熱効率が大幅に向上されることとなる。
【００７２】
このように、先行気筒２Ａ，２Ｄでは超リーンでの成層燃焼により熱効率が高められるとともにポンピングロスが低減され、一方、後続気筒２Ｂ，２Ｃでは、先行気筒２Ａ，２Ｄと同様にポンピングロス低減効果が得られるとともに、圧縮着火による燃焼を行う場合には、均一な混合気分布状態で圧縮着火が行われることにより熱効率が高められるとともに、これらの作用により、燃費が大幅に改善されることとなる。
【００７３】
また、先行気筒２Ａ，２Ｄでは理論空燃比の略２倍もしくはそれ以上のリーン空燃比とされることでＮＯｘ発生量が比較的少なく抑えられ、後続気筒２Ｂ，２Ｃでは、先行気筒２Ａ，２Ｄから既燃ガスが導入されることで多量のＥＧＲが行われているのと同等の状態となることからＮＯｘの発生が充分に抑制される。このような点からもエミッションの向上に有利となる。
【００７４】
但し、特殊運転モード中であっても、後続気筒２Ｂ，２Ｃの筒内温度が低く、圧縮着火し難い状態のときには後続気筒２Ｂ，２Ｃの着火を点火プラグ７による強制点火に切換える。また逆に、後続気筒２Ｂ，２Ｃの筒内温度が高過ぎてノッキング等の異常燃焼が起こるような状態のときには、特殊運転モードから通常運転モードに切換える。何れの場合も圧縮着火による特殊運転モードに比べ、燃費等の向上効果が抑制される。従って、これらの効果をより多く得るためには、特殊運転モード中の圧縮着火に適した運転領域を拡大することが望ましい。
【００７５】
特殊運転モード中の圧縮着火に適した運転領域を拡大するため、吸排気弁の開閉時期は次のように設定されている。
【００７６】
図８は、図５の吸排気行程部分を詳細に示したものであり、特殊運転モードにおける先行気筒２Ａ，２Ｄの先行気筒吸気弁３１および既燃ガス排出弁３２ｂの開閉時期と、後続気筒２Ｂ，２Ｃの既燃ガス導入弁３１ｂおよび後続気筒排気弁３２の開閉時期とを示す説明図である。図８（ａ）は特殊運転モードを行う運転状態のうち、比較的低負荷低回転の領域（図４の領域Ａ１）の場合であり、図８（ｂ）は同様に比較的高負荷高回転の領域（図４の領域Ａ２）の場合である。これらの図で、横軸はクランク角を示し、Ｔは上死点（ＴＤＣ）、Ｂは下死点（ＢＤＣ）である。ＴとＢとの間隔は１８０°ＣＡである。また、上段は先行気筒２Ａ，２Ｄを示し、下段はそれに対応する後続気筒２Ｂ，２Ｃを示す。そして、帯状の各部分は、各弁の開弁期間を示す。上段から下段に向かう白抜き矢印は、先行気筒２Ａ，２Ｄの排気行程と後続気筒２Ｂ，２Ｃの吸気行程とが重なっており、先行気筒２Ａ，２Ｄでの既燃ガスが後続気筒２Ｂ，２Ｃに導かれる状態を示す。
【００７７】
図８（ａ）において、上段には先行気筒２Ａ，２Ｄの既燃ガス排出弁３２ｂが開弁する既燃ガス排出弁の開弁期間８０と、先行気筒吸気弁３１が開弁する先行気筒吸気弁の開弁期間８１（斜線で示す）とを示す。下段には後続気筒２Ｂ，２Ｃの後続気筒排気弁３２が開弁する後続気筒排気弁の開弁期間８２と、既燃ガス導入弁３１ｂが開弁する既燃ガス導入弁の開弁期間８３とを示す。先行気筒吸気弁の開弁期間８１は、ＴＤＣ前約１０°ＣＡからＢＤＣ後約５５°ＣＡ（トータル約２４５°ＣＡ）に設定されている。これは、従来のエンジンの一般的な設定値である。これに対し、既燃ガス導入弁の開弁期間８３は、ＴＤＣ前約４５°ＣＡからほぼＢＤＣ（トータル約２２５°ＣＡ）に設定されている。また、後続気筒排気弁の開弁期間８２は、ＢＤＣ前約８０°ＣＡからＴＤＣ前約２５°ＣＡ（トータル約２３５°ＣＡ）に設定されている。
【００７８】
即ち、後続気筒吸気行程下死点９６と既燃ガス導入弁の閉弁時期９７との間隔（約０°ＣＡ）が、先行気筒吸気行程下死点９２と先行気筒吸気弁の閉弁時期９３との間隔（約５５°ＣＡ）よりも短くなるように設定されている。そして既燃ガス導入弁の開弁期間８３は先行気筒吸気弁の開弁期間８１および後続気筒排気弁の開弁期間８２よりも短くなっている。このように後続気筒２Ｂ，２Ｃでは、既燃ガス導入弁の開弁期間８３を短く設定するとともに、既燃ガス導入弁の閉弁時期９７を早期かつＢＤＣに近い時期に設定しているので、ピストン３がＢＤＣを過ぎて上昇中にも既燃ガス導入弁３１ｂが開いている期間が無い、若しくは短くなっている。このため、有効圧縮比が増大して幾何学的圧縮比に近づいている。この有効圧縮比の増大に伴い、後続気筒２Ｂ，２Ｃの筒内温度は上昇し易くなり、圧縮着火性が高くなっている。なお、このとき後続気筒２Ｂ，２Ｃに導入されるのは多連スロットル弁１７に絞られた新気ではなく、先行気筒２Ａ，２Ｄから排出された既燃ガスであり、容易に燃焼室４内に流入するため、既燃ガス導入弁３１ｂを早期に閉じてもその充填は充分になされる。
【００７９】
また、既燃ガス排出弁の開弁期間８０はＢＤＣ前約８０°ＣＡからＴＤＣ前約２５°ＣＡ（トータル約２３５°ＣＡ）に設定されている。従って、既燃ガス排出弁の閉弁時期９０は既燃ガス導入弁の閉弁時期９７（先行気筒２Ａ，２ＤではＴＤＣに相当）よりも約２５°ＣＡ進み側に設定されている。これは、従来のエンジンの一般的な設定値（ＴＤＣ後５０°ＣＡ程度）よりも早期になっている。既燃ガス排出弁３２ｂを早期に閉じることにより、先行気筒２Ａ，２Ｄの内部ＥＧＲ量が増大し、後続気筒２Ｂ，２Ｃに導入される既燃ガス温度が上昇する。
【００８０】
以上のように、後続気筒２Ｂ，２Ｃの有効圧縮比を増大させるとともに、後続気筒２Ｂ，２Ｃに導入される既燃ガス温度を上昇させることにより、後続気筒２Ｂ，２Ｃの筒内温度を上昇させて圧縮着火性を向上させている。このようにすることにより、圧縮着火による燃焼を行うことができる運転領域をより低負荷の領域まで拡大している。
【００８１】
図８（ｂ）において、上段には先行気筒２Ａ，２Ｄの既燃ガス排出弁３２ｂが開弁する既燃ガス排出弁の開弁期間１００と、先行気筒吸気弁３１が開弁する先行気筒吸気弁の開弁期間１０１（斜線で示す）とを示す。下段には後続気筒２Ｂ，２Ｃの後続気筒排気弁３２が開弁する後続気筒排気弁の開弁期間１０２と、既燃ガス導入弁３１ｂが開弁する既燃ガス導入弁の開弁期間１０３とを示す。先行気筒吸気弁の開弁期間１０１は、図８（ａ）の先行気筒吸気弁の開弁期間８１と同様の設定になっている。一方、既燃ガス排出弁の開弁期間１００、後続気筒排気弁の開弁期間１０２および既燃ガス導入弁の開弁期間１０３（それぞれ白抜き帯線で示す）は、図８（ａ）の既燃ガス排出弁の開弁期間８０、後続気筒排気弁の開弁期間８２および既燃ガス導入弁の開弁期間８３に対し、各期間の長さが等しく、各弁の開閉時期のみが全体的に約４５°ＣＡだけ遅らせた設定となっている。これは前述のように、カム位相制御手段４９によってカムシャフト３４の位相を約４５°ＣＡ遅らせることによりなされている（図３参照）。
【００８２】
図８（ｂ）の既燃ガス導入弁の閉弁時期１１４は、図８（ａ）の既燃ガス導入弁の閉弁時期９７よりも約４５°ＣＡ遅れている。このため、後続気筒２Ｂ，２Ｃでは、ピストン３が後続気筒吸気行程下死点１１３を過ぎて上昇中であっても、約４５°ＣＡの間既燃ガス導入弁３１ｂが開いている。従って、後続気筒２Ｂ，２Ｃの有効圧縮比は図８（ａ）の場合よりも減少しており、後続気筒２Ｂ，２Ｃの筒内温度は低下する。
【００８３】
また、既燃ガス排出弁の閉弁時期１１２は、図８（ａ）の既燃ガス排出弁の閉弁時期９０よりも約４５°ＣＡ遅く、先行気筒排気行程上死点１１１後約２０°ＣＡに設定されている。このため、先行気筒２Ａ，２Ｄの内部ＥＧＲは図８（ａ）の場合よりも減少し、後続気筒２Ｂ，２Ｃに導入される既燃ガス温度は低下する。
【００８４】
以上のように、後続気筒２Ｂ，２Ｃの有効圧縮比を減少させるとともに、後続気筒２Ｂ，２Ｃに導入される既燃ガス温度を低下させることにより、後続気筒２Ｂ，２Ｃの筒内温度を低下させ、ノッキング等の異常燃焼の発生を防止している。このようにすることにより、圧縮着火による燃焼を行うことができる運転領域をより高負荷の領域まで拡大している。
【００８５】
このように特殊運転モードでは、後続気筒２Ｂ，２Ｃの筒内温度が比較的低くなる運転領域ではその筒内温度を上昇させる方向に、筒内温度が比較的高くなる運転領域では低下させる方向に吸排気弁の開閉時期が設定されている。それによって、後続気筒２Ｂ，２Ｃが圧縮着火による適正な燃焼を行い得る運転領域を拡大することができ、より一層燃費改善効果と排ガス浄化の促進を図ることができる。
【００８６】
一方、通常運転モードでは前述のように先行気筒排気弁３２ａ及び後続気筒吸気弁３１ａが作動状態、既燃ガス排出弁３２ｂ及び既燃ガス導入弁３１ｂが停止状態とされることにより、実質的な新気及びガスの流通経路は図７に示すようになり、実質的に各気筒２Ａ〜２Ｄの吸気ポート１１，１１ａ及び排気ポート１２ａ，１２が独立し、吸気通路１５から各気筒２Ａ〜２Ｄの吸気ポート１１，１１ａに新気が導入されるとともに各気筒２Ａ〜２Ｄの排気ポート１２ａ，１２から排気通路２０に既燃ガスが排出される。
【００８７】
図９は、通常運転モードにおける先行気筒２Ａ，２Ｄの先行気筒吸気弁３１および既燃ガス排出弁３２ｂの開閉時期と、後続気筒２Ｂ，２Ｃの既燃ガス導入弁３１ｂおよび後続気筒排気弁３２の開閉時期とを示す説明図である。横軸はクランク角を示し、Ｔは上死点（ＴＤＣ）、Ｂは下死点（ＢＤＣ）である。ＴとＢの間隔は１８０°ＣＡである。また、上段は先行気筒２Ａ，２Ｄを示し、下段は後続気筒２Ｂ，２Ｃを示す。通常運転モードでは各気筒独立運転となっているので、先行気筒２Ａ，２Ｄ、後続気筒２Ｂ，２Ｃともに新気を導入して強制点火による燃焼を行っている。図９の上段には先行気筒２Ａ，２Ｄの先行気筒排気弁３２ａが開弁する既燃ガス排出弁の開弁期間１２０と、先行気筒吸気弁３１が開弁する先行気筒吸気弁の開弁期間１２１とを示す。下段には後続気筒２Ｂ，２Ｃの後続気筒排気弁３２が開弁する後続気筒排気弁の開弁期間１２２と、後続気筒吸気弁３１ａが開弁する後続気筒吸気弁の開弁期間１２３とを示す。
【００８８】
先行気筒２Ａ，２Ｄでの排気および後続気筒２Ｂ，２Ｃでの吸気は、特殊運転モードとは異なる弁によってなされるので、その開閉は異なるカムによりなされる。従って、先行気筒排気弁の開弁期間１２０および後続気筒吸気弁の開弁期間１２３は、図８（ａ）の既燃ガス排出弁の開弁期間８０および既燃ガス導入弁の開弁期間８３とは独立して設定することができる。図９では、先行気筒吸気弁の開弁期間１２１および後続気筒吸気弁の開弁期間１２３（斜線で示す）は、ＴＤＣ前約１０°ＣＡからＢＤＣ後約５５°ＣＡ（トータル約２４５°ＣＡ）に設定されている。また、先行気筒排気弁の開弁期間１２０および後続気筒排気弁の開弁期間１２２は、ＢＤＣ前約３０°ＣＡからＴＤＣ後約２５°ＣＡ（トータル約２３５°ＣＡ）に設定されている。これらの開弁期間は、従来のエンジンの一般的な設定に相当する。
【００８９】
そして、先行気筒吸気弁の開弁期間１２１および後続気筒吸気弁の開弁期間１２３は、カム位相可変機構３３ａによって前後に変動可能であり、先行気筒排気弁の開弁期間１２０および後続気筒排気弁の開弁期間１２２は、カム位相可変機構３４ａによって前後に変動可能である。従って、カム位相可変機構３３ａ，３４ａを制御することにより、先行気筒排気弁３２ａと先行気筒吸気弁３１とが共に開弁している期間（先行気筒排気弁の開弁期間１２０と先行気筒吸気弁の開弁期間１２１とが重なっている期間：バルブオーバーラップ）を変動し得るようになっている。カム位相制御手段４９は、高負荷になるほどバルブオーバーラップが大きくなるようにカム位相可変機構３３ａ，３４ａを制御して、負荷に応じて最適な燃焼効率が得られるようにしている。後続気筒２Ｂ，２Ｃに対しても同様の制御がなされている。
【００９０】
このようにして通常運転モードでは、負荷に応じて最適な吸排気時期に制御されるとともに、理論空燃比もしくはそれよりリッチとなるように吸入空気量及び燃料噴射量が制御されることにより、出力性能が確保される。
【００９１】
なお図８（ａ），（ｂ）および図９に示すパターンは、各開弁期間や開弁時期を限定するものではなく、特許請求の範囲内で適宜変更しても良い。例えば、図８（ａ）の既燃ガス導入弁の閉弁時期９７は、後続気筒吸気行程下死点９６よりも若干遅れるように設定しても良い。また、図８（ａ），（ｂ）および図９の先行気筒吸気弁の開弁期間８１，１０１，１２１を負荷に応じて異なる設定（当実施形態では、それはカム位相可変機構３３ａによりなされる）としても良く、当実施形態のように同一の設定としても良い（その場合、カム位相可変機構３３ａはなくても良い）。その他の具体的な設定値も、エンジンの要求特性に応じて好適な値に設定して良い。
【００９２】
また、特殊運転モードでの走行領域Ａを、領域Ａ１，Ａ２といった２分割ではなく、３以上に分割し、それぞれの領域に適した弁開閉時期を設定するようにしてもよい。更に分割による段階的な設定ではなく、連続的に変化させるようにしても良い。
【００９３】
次に、本発明の第２の実施形態について図１０〜図１３に基いて説明する。なお、これらの図において第１の実施形態と同一部分には同一符号を付して示し、その重複説明を省略する。
【００９４】
図１０は、第２の実施形態におけるカム切換機構１５０の部分斜視図であり、２点鎖線で示す既燃ガス導入弁３１ｂ等のためのものである。既燃ガス導入弁３１ｂの上方にはカムシャフト１５１が配設されている。カムシャフト１５１には第１カム１５２，第２カム１５４および第３カム１５６という独立したリフト特性を有する３種類のカムが一体回転するように設けられている。これらのカムと既燃ガス導入弁３１ｂとの間には、ロッカシャフト１７０に支持されたロッカアームセット１６０が設けられている。ロッカアームセット１６０は、第１ロッカアーム１６２，第２ロッカアーム１６４および第３ロッカアーム１６６という３種類のロッカアームの集合体である。第１ロッカアーム１６２の先端にはバルブ当接部１６３と、その軸線方向位置を微調整するためのアジャストスクリュー１６１が設けられており、バルブ当接部１６３は適切な位置で既燃ガス導入弁３１ｂ等の弁軸上端に当接している。第２ロッカアーム１６４および第３ロッカアーム１６６には、図外のスプリングが、これらのロッカアームを第２カム１５４および第３カム１５６に押圧するように設けられている。従って、ロッカアームセット１６０の各ロッカアームが、図示のように独立して可動である場合には、各ロッカアームの上面は第１カム１５２，第２カム１５４および第３カム１５６の外周部に当接し、カム当接部の形状（各カムの回転半径）に応じてロッカシャフト１７０を支軸として上下に揺動する。
【００９５】
ロッカアームセット１６０の内部には、後述するように５本のプランジャが２列に設けられている（図１１参照。図１０ではそのプランジャ穴のひとつである第４プランジャ穴２０４が見えている）。これらのプランジャの動きにより、第１ロッカアーム１６２は第２ロッカアーム１６４または第３ロッカアーム１６６と一体となり、連動し得る。ロッカシャフト１７０の内部には、プランジャを油圧作動させるためのオイルを導く第１作動油給排用の通路１７２および第２作動油給排用の通路１７４が設けられている。
【００９６】
第１カム１５２は、弁停止用のカムであり、カムシャフト１５１と同心円の外周形状を有する。したがって第１ロッカアーム１６２は、その上面が第１カム１５２の外周面に常時当接している（第２ロッカアーム１６４および第３ロッカアーム１６６と切り離されている）とき、カムシャフト１５１が回転しても揺動しない。すなわち既燃ガス導入弁３１ｂ等は閉弁状態で停止している。
【００９７】
第２カム１５４は、低負荷（または低速）用のカムであり、第１カム１５２と同一の外周形状を有する部分と、それより突出した外周形状を有する部分とからなる。したがって第２ロッカアーム１６４は、その上面が第２カム１５４の外周面に常時当接している（後述の第３ロッカアーム１６６と切り離されている）とき、カムシャフト１５１の回転に伴い、所定のクランク角で所定量だけ下方に揺動する。そして、このとき第１ロッカアーム１６２と第２ロッカアーム１６４とが連動するようになっていれば、第１ロッカアーム１６２の動作は第２カム１５４による第２ロッカアーム１６４の揺動と同一のものとなる。すなわち既燃ガス導入弁３１ｂは所定時期に所定量だけ開弁する。
【００９８】
第３カム１５６は、高負荷（または高速）用のカムであり、第２カム１５４と同一の外周形状を有する部分と、それより突出した外周形状を有する部分とからなる。したがって第３ロッカアーム１６６は、その上面が第３カム１５６の外周面に常時当接しており、カムシャフト１５１の回転に伴い、所定のクランク角で所定量だけ下方に揺動する。そして、このとき第１ロッカアーム１６２と第３ロッカアーム１６６とが連動するようになっていれば、第１ロッカアーム１６２の動作は第３カム１５６による第３ロッカアーム１６６の揺動と同一のものとなる。すなわち既燃ガス導入弁３１ｂは所定時期に所定量だけ開弁する（開弁期間は第２ロッカアーム１６４のみが第１ロッカアーム１６２と連動する場合の開弁期間を包含する）。
【００９９】
図１１は、ロッカアームセット１６０の内部に設けられた５本のプランジャの作動を示す説明図である。図１１（ａ）は第１ロッカアーム１６２が第２ロッカアーム１６４および第３ロッカアーム１６６と切り離された状態、図１１（ｂ）は第１ロッカアーム１６２が第２ロッカアーム１６４のみと連動する状態、図１１（ｃ）は第１ロッカアーム１６２が第２ロッカアーム１６４および第３ロッカアーム１６６と連動する状態を示す。
【０１００】
第１ロッカアーム１６２の内部には第１プランジャ穴２０１および第４プランジャ穴２０４が設けられている。第１プランジャ穴２０１は第２ロッカアーム１６４側に開口した円形断面の凹穴である。第１プランジャ穴２０１の底部には第１作動油給排用の通路１７２から第１作動油導入路１７３が導かれている。第１プランジャ穴２０１内には円柱状の第１プランジャ１８１が嵌挿されている。第１プランジャ１８１は、その外周面で第１作動油導入路１７３に導かれた作動油をシールしつつ、第１プランジャ穴２０１内を滑らかに摺動する。第１プランジャ１８１の全長は、第１プランジャ穴２０１の深さよりも短い。
【０１０１】
第４プランジャ穴２０４は第２ロッカアーム１６４側と第３ロッカアーム１６６側とを貫通する貫通穴である。第４プランジャ穴２０４内には円柱状の第４プランジャ１８４が嵌挿されている。第４プランジャ１８４の全長は、第４プランジャ穴２０４の深さ（第１ロッカアーム１６２の板厚）と等しい。第４プランジャ１８４は、第４プランジャ穴２０４内を滑らかに摺動する。
【０１０２】
第２ロッカアーム１６４の内部には第２プランジャ穴２０２および第５プランジャ穴２０５が設けられている。第２プランジャ穴２０２は第１ロッカアーム１６２側に開口した円形断面の凹穴であり、第１プランジャ穴２０１と等しい直径となっている。第２プランジャ穴２０２の底部にはエア抜き穴２０６が設けられ、リークしたオイルを逃がしつつ内部の気圧を大気圧に保つ。第２プランジャ穴２０２内には有底円筒状で外径が第１プランジャ１８１と等しい第２プランジャ１８２が嵌挿されている。第２プランジャ１８２は、第２プランジャ穴２０２内を滑らかに摺動する。第２プランジャ１８２の全長は、第２プランジャ穴２０２の深さと等しい。第２プランジャ１８２の第１プランジャ１８１と当接する端部は、球状に成形されている。第２プランジャ１８２の内側凹部には第２プランジャスプリング１８７が設けられ、第２プランジャ１８２を常時第１プランジャ１８１側に付勢している。
【０１０３】
第５プランジャ穴２０５は第１ロッカアーム１６２側に開口した円形断面の凹穴であり、第４プランジャ穴２０４と等しい直径となっている。第５プランジャ穴２０５の底部にはエア抜き穴２０７が設けられ、リークしたオイルを逃がしつつ内部の気圧を大気圧に保つ。第５プランジャ穴２０５内には有底円筒状で外径が第４プランジャ１８４と等しい第５プランジャ１８５が嵌挿されている。第５プランジャ１８５は、第５プランジャ穴２０５内を滑らかに摺動する。第５プランジャ１８５の全長は、第５プランジャ穴２０５の深さよりも短い。第５プランジャ１８５の第４プランジャ１８４と当接する端部は、球状に成形されている。第５プランジャ１８５の内側凹部には第５プランジャスプリング１８９が設けられ、第５プランジャ１８５を常時第４プランジャ１８４側に付勢している。
【０１０４】
第３ロッカアーム１６６の内部には第３プランジャ穴２０３が設けられている。第３プランジャ穴２０３は第１ロッカアーム１６２側に開口した円形断面の凹穴であり、第４プランジャ穴２０４と等しい直径となっている。第３プランジャ穴２０３の底部には第２作動油給排用の通路１７４から第２作動油導入路１７５が導かれている。第３プランジャ穴２０３内には円柱状で外径が第４プランジャ１８４と等しい第３プランジャ１８３が嵌挿されている。第３プランジャ１８３は、その外周面で第２作動油導入路１７５に導かれた作動油をシールしつつ、第３プランジャ穴２０３内を滑らかに摺動する。第３プランジャ１８３の全長は、第３プランジャ穴２０３の深さと等しい。第３プランジャ１８３の第４プランジャ１８４と当接する端部は、球状に成形されている。
【０１０５】
図１１（ａ）は、第１ロッカアーム１６２が第２ロッカアーム１６４および第３ロッカアーム１６６と切り離された状態を示し、第１作動油給排用の通路１７２に作動油圧が供給（以下油圧ＯＮという）され、第２作動油給排用の通路１７４に作動油圧が供給されていない（以下油圧ＯＦＦという）。第１作動油給排用の通路１７２から導かれた第１作動油導入路１７３が油圧ＯＮとなり、第１プランジャ１８１を右側（図の矢印方向）に押圧する。その押圧力は第２プランジャスプリング１８７の付勢力よりも大きく、第１プランジャ１８１は第２プランジャ１８２と一体となって右側に移動している。第２プランジャ１８２の全長が第２プランジャ穴２０２の深さと等しいので、第１プランジャ１８１と第２プランジャ１８２との接点は第１ロッカアーム１６２と第２ロッカアーム１６４との合わせ面内にある。
【０１０６】
一方、第２作動油給排用の通路１７４から導かれた第２作動油導入路１７５が油圧ＯＦＦとなっているので、第３プランジャ１８３、第４プランジャ１８４および第５プランジャ１８５は、第５プランジャスプリング１８９の付勢力によって一体となって左側（図の矢印方向）に移動している。第３プランジャ１８３の全長が第３プランジャ穴２０３の深さと等しいので、第３プランジャ１８３と第４プランジャ１８４との接点は第１ロッカアーム１６２と第３ロッカアーム１６６との合わせ面内にある。更に第４プランジャ１８４の全長が第４プランジャ穴２０４の深さと等しいので、第４プランジャ１８４と第５プランジャ１８５との接点は第１ロッカアーム１６２と第２ロッカアーム１６４との合わせ面内にある。
【０１０７】
このように、各プランジャの接点が各ロッカアームの合わせ面内にあるため、第１ロッカアーム１６２は第２ロッカアーム１６４および第３ロッカアーム１６６から切り離された状態となっている。このため第１ロッカアーム１６２は、その上面に当接する第１カム１５２による作動、すなわちロッカシャフト１７０まわりの揺動停止を行い、既燃ガス導入弁３１ｂを閉弁状態で停止させる。
【０１０８】
図１１（ｂ）は、第１ロッカアーム１６２が第２ロッカアーム１６４のみと連動する状態を示し、第１作動油給排用の通路１７２、第２作動油給排用の通路１７４ともに油圧ＯＦＦとなっている。第１作動油給排用の通路１７２から導かれた第１作動油導入路１７３が油圧ＯＦＦとなっているので、第１プランジャ１８１および第２プランジャ１８２は、第２プランジャスプリング１８７の付勢力によって左側（図の矢印方向）に移動している。第１プランジャ１８１の全長が第１プランジャ穴２０１の深さより短いので、第２プランジャ１８２の一部は第１プランジャ穴２０１に入り込んでいる。
【０１０９】
一方、第２作動油給排用の通路１７４から導かれた第２作動油導入路１７５が油圧ＯＦＦとなっているので、図１１（ａ）と同様、第３プランジャ１８３と第４プランジャ１８４との接点は第１ロッカアーム１６２と第３ロッカアーム１６６との合わせ面内にあり、第４プランジャ１８４と第５プランジャ１８５との接点は第１ロッカアーム１６２と第２ロッカアーム１６４との合わせ面内にある。
【０１１０】
このように、第２プランジャ１８２の一部が第１プランジャ穴２０１に入り込むことにより、第１ロッカアーム１６２は第２ロッカアーム１６４と連動する。また、第３プランジャ１８３と第４プランジャ１８４との接点が第１ロッカアーム１６２と第３ロッカアーム１６６との合わせ面内にあるので、第１ロッカアーム１６２と第３ロッカアーム１６６とは切り離された状態となっている。従って第１ロッカアーム１６２は、第２ロッカアーム１６４の上面に当接する第２カム１５４による作動を行う。すなわちカムシャフト１５１の回転に伴い、所定のクランク角で所定量だけ下方に揺動し、既燃ガス導入弁３１ｂを開閉させる。
【０１１１】
図１１（ｃ）は、第１ロッカアーム１６２が第２ロッカアーム１６４および第３ロッカアーム１６６と連動する状態を示し、第１作動油給排用の通路１７２が油圧ＯＦＦ、第２作動油給排用の通路１７４が油圧ＯＮとなっている。第１作動油給排用の通路１７２から導かれた第１作動油導入路１７３が油圧ＯＦＦとなっているので、図１１（ｂ）と同様、第２プランジャ１８２の一部は第１プランジャ穴２０１に入り込んでいる。
【０１１２】
一方、第２作動油給排用の通路１７４から導かれた第２作動油導入路１７５が油圧ＯＮとなっているので、第３プランジャ１８３を右側（図の矢印方向）に押圧する。その押圧力は第５プランジャスプリング１８９の付勢力よりも大きく、第３プランジャ１８３は第４プランジャ１８４および第５プランジャ１８５と一体となって右側に移動している。第５プランジャ１８５の全長が第５プランジャ穴２０５の深さよりも短いので、第４プランジャ１８４の一部が第５プランジャ穴２０５に入り込み、さらに第３プランジャ１８３の一部が第４プランジャ穴２０４に入り込んでいる。
【０１１３】
このように、第２プランジャ１８２の一部が第１プランジャ穴２０１に入り込み、第３プランジャ１８３の一部が第４プランジャ穴２０４に入り込むことにより、第１ロッカアーム１６２は第２ロッカアーム１６４および第３ロッカアーム１６６と連動する。従って第１ロッカアーム１６２は、第３ロッカアーム１６６の上面に当接する、最も回転半径の大きな第３カム１５６による作動を行う。すなわちカムシャフト１５１の回転に伴い、所定のクランク角で所定量だけ下方に揺動し、既燃ガス導入弁３１ｂを開閉させる。その開弁期間は図１１（ｂ）の場合よりも長くなっている。
【０１１４】
以上のカム切換機構１５０は、既燃ガス導入弁３１ｂおよび既燃ガス排出弁３２ｂのためのものであるが、後続気筒吸気弁３１ａおよび先行気筒排気弁３２ａにも同様のカム切換機構１５０ａ（図１０に括弧書きで示す）が設けられている。但し、カム切換機構１５０ａでは第２カム１５４と第３カム１５６は同一形状である。それらのカムに当接するロッカアームセット１６０ａは、図１０に示すように第１ロッカアーム１６２ａ、第２ロッカアーム１６４ａおよび第３ロッカアーム１６６ａからなる。カム切換機構１５０ａは、第１ロッカアーム１６２ａが第２ロッカアーム１６４ａおよび第３ロッカアーム１６６ａから切り離され、後続気筒吸気弁３１ａおよび先行気筒排気弁３２ａを閉弁状態で停止させる状態と、第１ロッカアーム１６２ａが第２ロッカアーム１６４ａおよび第３ロッカアーム１６６ａと連動し、第２カム１５４および第３カム１５６の回転によって後続気筒吸気弁３１ａおよび先行気筒排気弁３２ａを開閉させる状態とに切換える。
【０１１５】
図１２は、ロッカアームセット１６０ａの内部に設けられた３本のプランジャの作動を示す説明図である。図１２（ａ）は第１ロッカアーム１６２が第２ロッカアーム１６４および第３ロッカアーム１６６と切り離された状態、図１２（ｂ）は第１ロッカアーム１６２が第２ロッカアーム１６４および第３ロッカアーム１６６と連動する状態を示す。
【０１１６】
ロッカアームセット１６０ａ内のプランジャ構造は、ロッカアームセット１６０内のプランジャ構造のうち、第３プランジャ１８３、第４プランジャ１８４および第５プランジャ１８５まわりの構造を設けたようなものとなっており、その詳細構造の説明はロッカアームセット１６０における記述と重複するので省略する。但し、第３プランジャ１８３の左端には、第１作動油給排用の通路１７２から第１作動油導入路１７３ａが導かれている点がロッカアームセット１６０とは異なる。また、第２作動油給排用の通路１７４は、構造上省略できる場合にはなくても良い。
【０１１７】
図１２（ａ）は、第１ロッカアーム１６２が第２ロッカアーム１６４および第３ロッカアーム１６６と切り離された状態を示し、第１作動油給排用の通路１７２が油圧ＯＦＦとなっている。第１作動油給排用の通路１７２から導かれた第１作動油導入路１７３ａが油圧ＯＦＦとなっているので、第３プランジャ１８３、第４プランジャ１８４および第５プランジャ１８５は、第５プランジャスプリング１８９の付勢力によって一体となって左側（図の矢印方向）に移動している。従って、第３プランジャ１８３と第４プランジャ１８４との接点は第１ロッカアーム１６２ａと第３ロッカアーム１６６ａとの合わせ面内にあり、第４プランジャ１８４と第５プランジャ１８５との接点は第１ロッカアーム１６２ａと第２ロッカアーム１６４ａとの合わせ面内にある。
【０１１８】
このように、各プランジャの接点が各ロッカアームの合わせ面内にあるため、第１ロッカアーム１６２ａは第２ロッカアーム１６４ａおよび第３ロッカアーム１６６ａから切り離された状態となっている。このため第１ロッカアーム１６２ａは、その上面に当接する第１カム１５２による作動、すなわちロッカシャフト１７０まわりの揺動停止を行い、既燃ガス導入弁後続気筒吸気弁３１ａおよび先行気筒排気弁３２ａを閉弁状態で停止させる。
【０１１９】
図１２（ｂ）は、第１ロッカアーム１６２が第２ロッカアーム１６４および第３ロッカアーム１６６と連動する状態を示し、第１作動油給排用の通路１７２が油圧ＯＮとなっている。第１作動油給排用の通路１７２から導かれた第１作動油導入路１７３ａが油圧ＯＮとなっているので、第３プランジャ１８３を右側（図の矢印方向）に押圧する。その押圧力は第５プランジャスプリング１８９の付勢力よりも大きく、第３プランジャ１８３は第４プランジャ１８４および第５プランジャ１８５と一体となって右側に移動している。従って第４プランジャ１８４の一部が第５プランジャ穴２０５に入り込み、さらに第３プランジャ１８３の一部が第４プランジャ穴２０４に入り込んでいる。
【０１２０】
このように、第４プランジャ１８４の一部が第５プランジャ穴２０５に入り込み、さらに第３プランジャ１８３の一部が第４プランジャ穴２０４に入り込むことにより、第１ロッカアーム１６２は第２ロッカアーム１６４および第３ロッカアーム１６６と連動する。従って第１ロッカアーム１６２は、第２ロッカアーム１６４ａおよび第３ロッカアーム１６６ａの上面に当接する第２カム１５４および第３カム１５６（同形状）による作動を行う。すなわちカムシャフト１５１の回転に伴い、所定のクランク角で所定量だけ下方に揺動し、既燃ガス導入弁３１ｂを開閉させる。
【０１２１】
図１３は当実施形態の駆動、制御系統の構成を示している。第１コントロール弁１７６および第２コントロール弁１７７は、第１作動油給排用の通路１７２および第２作動油給排用の通路１７４に導く第１作動油および第２作動油を制御（油圧ＯＮ／ＯＦＦ）するためのコントロール弁である。カム切換機構１５０は既燃ガス導入弁３１ｂ、既燃ガス排出弁３２ｂに設けられ、カム切換機構１５０ａは後続気筒吸気弁３１ａ、先行気筒排気弁３２ａに設けられている。
【０１２２】
ＥＣＵ４０ａは、第１の実施形態のＥＣＵ４０に対し、弁停止機構制御手段４２およびカム位相制御手段４９に替えてカム切換制御手段１９０を備える点が異なっている。
【０１２３】
カム切換制御手段１９０は、特殊運転モードと通常運転モードとに応じ、あるいは運転状態に応じ、第１コントロール弁１７６および第２コントロール弁１７７を制御することにより、カム切換機構１５０およびカム切換機構１５０ａを次のように制御する。
【０１２４】
特殊運転モードのうち、低負荷低速領域：
・第１作動油圧−ＯＦＦ，第２作動油圧−ＯＦＦ
・先行気筒排気弁３２ａ、後続気筒吸気弁３１ａを停止状態（図１２（ａ））
・既燃ガス排出弁３２ｂ、既燃ガス導入弁３１ｂを
第２カム１５４（低速カム）による作動状態（図１１（ｂ））
特殊運転モードのうち、高負荷高速領域：
・第１作動油圧−ＯＦＦ，第２作動油圧−ＯＮ
・先行気筒排気弁３２ａ、後続気筒吸気弁３１ａを停止状態（図１２（ａ））
・既燃ガス排出弁３２ｂ、既燃ガス導入弁３１ｂを
第３カム１５６（高速カム）による作動状態（図１１（ｃ））
通常運転モード：
・第１作動油圧−ＯＮ，第２作動油圧−ＯＦＦ
・先行気筒排気弁３２ａ、後続気筒吸気弁３１ａを
第２カム１５４および第３カム１５６による作動状態（図１２（ｂ））
・既燃ガス排出弁３２ｂ、既燃ガス導入弁３１ｂを停止状態（図１１（ａ））。
【０１２５】
次に当実施形態の装置の作用を説明するが、第１の実施形態と重複する部分の記述は省略する。図１４は、当実施形態の吸排気行程部分を詳細に示した説明図である。図１４（ａ）は特殊運転モードを行う運転状態のうち、比較的低負荷低回転の領域（図４の領域Ａ１）の場合であり、図１４（ｂ）は同様に比較的高負荷高回転の領域（図４の領域Ａ２）の場合である。その他表記方法については図８に準ずる。
【０１２６】
図１４（ａ）の上段には先行気筒２Ａ，２Ｄの既燃ガス排出弁３２ｂが開弁する既燃ガス排出弁の開弁期間２３０（斜線で示す）と、先行気筒吸気弁３１が開弁する先行気筒吸気弁の開弁期間２３１とを示す。下段には後続気筒２Ｂ，２Ｃの後続気筒排気弁３２が開弁する後続気筒排気弁の開弁期間２３２と、既燃ガス導入弁の開弁期間２３３（斜線で示す）とを示す。また、既燃ガス排出弁の開弁期間２３０および既燃ガス導入弁の開弁期間２３３の直下に、後述する既燃ガス排出弁の開弁期間２５０および既燃ガス導入弁の開弁期間２５３（いずれも特殊運転モードで比較的高回転高負荷の場合のもの）を参考として２点鎖線で示す。このような弁開閉時期の切換えは、カム切換機構１５０，１５０ａおよびカム切換制御手段１９０による。先行気筒吸気弁の開弁期間２３１は、ＴＤＣ前約１０°ＣＡからＢＤＣ後約５５°ＣＡ（トータル約２４５°ＣＡ）に設定されている。これは、従来のエンジンの一般的な設定値である。これに対し、既燃ガス導入弁の開弁期間２３３は、ほぼＴＤＣからＢＤＣ（トータル約１８０°ＣＡ）に設定されている。また、後続気筒排気弁の開弁期間２３２は、ＢＤＣ前約４５°ＣＡからＴＤＣ後約１０°ＣＡ（トータル約２３５°ＣＡ）に設定されている。
【０１２７】
即ち、後続気筒吸気行程下死点２４７と既燃ガス導入弁の閉弁時期２４８との間隔（約０°ＣＡ）が、先行気筒吸気行程下死点２４２と先行気筒吸気弁の閉弁時期２４３との間隔（約５５°ＣＡ）よりも短くなるように設定されている。そして既燃ガス導入弁の開弁期間２３３は先行気筒吸気弁の開弁期間２３１および後続気筒排気弁の開弁期間２３２よりも短くなっている。このように後続気筒２Ｂ，２Ｃでは、既燃ガス導入弁の開弁期間２３３を短く設定するとともに、既燃ガス導入弁の閉弁時期２４８を早期かつＢＤＣに近い時期に設定しているため、ピストン３がＢＤＣを過ぎて上昇中にも既燃ガス導入弁３１ｂが開いている期間が無い、若しくは短くなっている。このため、有効圧縮比が増大して幾何学的圧縮比に近づいている。この有効圧縮比の増大に伴い、後続気筒２Ｂ，２Ｃの筒内温度は上昇し易くなり、圧縮着火性が高くなっている。
【０１２８】
また、既燃ガス排出弁の開弁期間２３０はＢＤＣ前約４５°ＣＡからＴＤＣ前約２０°ＣＡ（トータル約２０５°ＣＡ）に設定されている。従って、既燃ガス排出弁の閉弁時期２４１は既燃ガス導入弁の閉弁時期２４８（先行気筒２Ａ，２ＤではＴＤＣに相当）よりも約２０°ＣＡ進み側に設定されている。これは、従来のエンジンの一般的な設定値（ＴＤＣ後５０°ＣＡ程度）よりも早期になっている。既燃ガス排出弁３２ｂを早期に閉じることにより、先行気筒２Ａ，２Ｄの内部ＥＧＲが増大し、後続気筒２Ｂ，２Ｃに導入される既燃ガス温度が上昇する。
【０１２９】
後続気筒排気弁の開弁期間２３２が、ＢＤＣ前約４５°ＣＡからＴＤＣ後約１０°ＣＡ（トータル約２３５°ＣＡ）に設定されているので、後続気筒排気弁３２は後続気筒排気行程上死点２４４では開弁している。一方、既燃ガス導入弁の開弁時期２４６は、ほぼ後続気筒吸気行程上死点２４５となるように設定されている。すなわち、後続気筒排気弁３２を後続気筒排気行程上死点２４４までは開弁しつつ、既燃ガス導入弁の開弁時期２４６が後続気筒吸気行程上死点２４５となるように設定されている。このため、後続気筒排気行程上死点２４４付近でのバルブオーバーラップが短縮されるので、後続気筒２Ｂ，２Ｃに導入された既燃ガスが、そのまま後続気筒排気弁３２を通して排気通路２０に排出される、いわゆる吹き抜けが防止されるとともに、後続気筒２Ｂ，２Ｃの有効圧縮率を増大させることができる。
【０１３０】
以上のように、後続気筒２Ｂ，２Ｃの有効圧縮比を増大させるとともに、後続気筒２Ｂ，２Ｃに導入される既燃ガス温度を上昇させることにより、後続気筒２Ｂ，２Ｃの筒内温度を上昇させて圧縮着火性を向上させている。このようにすることにより、圧縮着火による燃焼を行うことができる運転領域をより低負荷の領域まで拡大している。
【０１３１】
図１４（ｂ）は、比較的高負荷高回転の領域（図４の領域Ａ２）の場合の説明図である。上段には先行気筒２Ａ，２Ｄの既燃ガス排出弁３２ｂが開弁する既燃ガス排出弁の開弁期間２５０（網目で示す）と、先行気筒吸気弁の開弁期間２３１（図１４（ａ）と共通）とを示す。下段には後続気筒排気弁の開弁期間２３２（図１４（ａ）と共通）と、既燃ガス導入弁３１ｂが開弁する既燃ガス導入弁の開弁期間２５３（網目で示す）とを示す。また、既燃ガス排出弁の開弁期間２５０および既燃ガス導入弁の開弁期間２５３の直下に、図１４（ａ）の既燃ガス排出弁の開弁期間２３０および既燃ガス導入弁の開弁期間２３３を参考として２点鎖線で示す。
【０１３２】
既燃ガス導入弁の開弁期間２５３は、ＴＤＣ前約１０°ＣＡからＢＤＣ後約５５°ＣＡに設定されている。すなわち、低負荷低回転領域における既燃ガス導入弁の開弁期間２３３に対し、１０°ＣＡ早期に開弁し、５５°遅く閉弁する。従って、後続気筒２Ｂ，２Ｃの有効圧縮比は図１４（ａ）の場合よりも減少しており、後続気筒２Ｂ，２Ｃの筒内温度は低下する。
【０１３３】
また、既燃ガス排出弁の開弁期間２５０は、ＢＤＣ前約４５°ＣＡからＴＤＣ後約１０°ＣＡに設定されている。すなわち図１４（ａ）の既燃ガス排出弁の開弁期間２３０よりも約３０°ＣＡ長くなっている。このため、先行気筒２Ａ，２Ｄの内部ＥＧＲは図１４（ａ）の場合よりも減少し、後続気筒２Ｂ，２Ｃに導入される既燃ガス温度は低下する。
【０１３４】
以上のように、後続気筒２Ｂ，２Ｃの有効圧縮比を減少させるとともに、後続気筒２Ｂ，２Ｃに導入される既燃ガス温度を低下させることにより、後続気筒２Ｂ，２Ｃの筒内温度を低下させ、ノッキング等の異常燃焼の発生を防止している。このようにすることにより、圧縮着火による燃焼を行うことができる運転領域をより高負荷の領域まで拡大している。
【０１３５】
なお図１４（ａ），（ｂ）に示すパターンは、各開弁期間や開弁時期を限定するものではなく、特許請求の範囲内で適宜変更しても良い。例えば、図１４（ａ）の既燃ガス導入弁の閉弁時期２４８は、後続気筒吸気行程下死点２４７よりも若干遅れるように設定しても良い。その他の具体的な設定値も、エンジンの要求特性に応じて好適な値に設定して良い。
【０１３６】
当実施形態では、吸排気バルブの設置箇所やそれぞれの機能に応じて、３通りのカム切換えをするもの、２通りのカム切換えをするもの及びカム切換えをしないものが混在するような構造としたが、その組み合わせはこれに限定するものではなく、例えば全てのバルブに対し３通りのカムとロッカアームセット１６０を対応させて、３通りのカム切換えをするようにしても良い。そして、第２の実施形態では不変としていた先行気筒吸気弁の開弁期間２３１や後続気筒排気弁の開弁期間２３２を、カム切換えによって変動させるようにしても良い。
【０１３７】
本発明は、上記第１、第２の実施形態に限定されることなく、特許請求の範囲に記載した発明の範囲内で、種々の変形が可能である。例えば各弁の開閉時期を要求に応じて変動させる手段は、第１の実施形態のようにカムシャフトの位相を変動させる機構や第２の実施形態のようにカムを切換える機構に限定するものではなく、特許請求の範囲に示すような設定が得られるものであれば良い。
【０１３８】
本発明は４気筒エンジンに限定するものではなく、先行気筒と後続気筒とからなる１対の気筒対を、３対以上好適に組み合わせた６気筒以上のエンジンに適用しても良い。
【０１３９】
【発明の効果】
以上のように本発明の制御装置は、少なくとも低負荷低回転域で、排気行程と吸気行程が重なる一対の気筒間において排気行程側の気筒である先行気筒から排出される既燃ガスがそのまま吸気行程側の気筒である後続気筒に気筒間ガス通路を介して導入され、この後続気筒から排出される排ガスが排気通路に導かれるような２気筒接続状態にガス流通経路を構成するとともに、上記２気筒接続状態にあるときに、上記先行気筒では理論空燃比よりも所定量大きいリーン空燃比とした状態で強制点火により燃焼を行わせ、上記後続気筒では、先行気筒から導入されたリーン空燃比の既燃ガスに燃料を供給するとともに圧縮着火により燃焼を行わせるように各気筒の燃焼を制御する燃焼制御手段とを備え、少なくとも上記２気筒接続状態にあるときに、上記先行気筒に新気を導入する先行気筒吸気弁と上記気筒間ガス通路から上記後続気筒に既燃ガスを導入する既燃ガス導入弁とが設けられ、上記２気筒接続状態となる運転領域のうち、少なくとも低負荷側の所定領域において、上記後続気筒の吸気行程下死点と上記既燃ガス導入弁の閉弁時期との間隔が、上記先行気筒の吸気行程下死点と上記先行気筒吸気弁の閉弁時期との間隔よりも短くなるように設定されていることを特徴とするので、三元触媒だけで充分に排気浄化性能を確保しつつ、リーン燃焼やポンピングロス低減等による燃費改善効果を得るとともに、その寄与度の高い運転領域を拡大することにより、更に大きな燃費改善効果を得ることができる。
【図面の簡単な説明】
【図１】本発明の第１実施形態による装置を備えたエンジン全体の概略平面図である。
【図２】エンジン本体等の概略断面図である。
【図３】第１実施形態の制御系統のブロック図である。
【図４】運転領域を示す説明図である。
【図５】各気筒の排気行程、吸気行程、燃料噴射時期および点火時期等を示す図である。
【図６】低負荷低回転時の実質的な新気およびガスの流通経路を示す説明図である。
【図７】高負荷、高低回転側の運転領域にある時の実質的な新気およびガスの流通経路を示す説明図である。
【図８】第１実施形態の特殊運転モードにおける吸排気弁の開閉時期を示す説明図であり、（ａ）は、そのうち比較的低負荷低回転の場合、（ｂ）は、同じく比較的高負荷高回転の場合を示す。
【図９】通常運転モードにおける吸排気弁の開閉時期を示す説明図である。
【図１０】第２実施形態のカム切換機構を示す部分斜視図である。
【図１１】第２実施形態における３種のカム切換機構のためのプランジャ作動説明図である。
【図１２】第２実施形態における２種のカム切換機構のためのプランジャ作動説明図である。
【図１３】第２実施形態の制御系統のブロック図である。
【図１４】第２実施形態の特殊運転モードにおける吸排気弁の開閉時期を示す説明図であり、（ａ）は、そのうち比較的低負荷低回転の場合、（ｂ）は、同じく比較的高負荷高回転の場合を示す。
【符号の説明】
１ エンジン本体
２Ａ，２Ｄ １番，４番気筒（先行気筒）
２Ｂ，２Ｃ ２番，３番気筒（後続気筒）
９ 燃料噴射弁
１１ 吸気ポート
１１，１１ａ，１１ｂ 吸気ポート
１２，１２ａ，１２ｂ 排気ポート
１５ 吸気通路
２０ 排気通路
２２ 気筒間ガス通路
３１ 先行気筒吸気弁
３１ａ 後続気筒吸気弁
３１ｂ 既燃ガス導入弁
３２ 後続気筒排気弁
３２ａ 先行気筒排気弁
３２ｂ 既燃ガス排出弁
３３ａ，３４ａ カム位相可変機構
４０，４０ａ ＥＣＵ
４２ 弁停止機構制御手段
４４ 燃焼制御手段
４９ カム位相制御手段
１５０ カム切換機構
１６０ ロッカアームセット
１９０ カム切換制御手段[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a control device for a spark ignition engine, and more particularly to a device for controlling the combustion state of each cylinder and the opening / closing timing of intake and exhaust valves in order to improve fuel efficiency and emissions in a multi-cylinder engine.
[0002]
[Prior art]
Conventionally, in a spark ignition type engine, a technique for improving fuel efficiency has been studied by performing combustion in a state in which the air-fuel ratio of the air-fuel mixture in each cylinder is set to a lean air-fuel ratio larger than the stoichiometric air-fuel ratio. Provided with a fuel injection valve that directly injects fuel into the room, and in the low rotation and low load range, etc., stratified combustion is performed by injecting fuel from the fuel injection valve in the compression stroke, thereby realizing super lean combustion Is known (for example, see Patent Document 1).
[0003]
In such an engine, an ordinary three-way catalyst (a catalyst having a high purification performance in the vicinity of the theoretical air-fuel ratio with respect to HC, CO, and NOx) alone as an exhaust gas purification catalyst is sufficient for NOx during lean operation. Since the purification performance cannot be obtained, a lean NOx catalyst is provided that adsorbs NOx in an oxygen-excess atmosphere and separates and reduces NOx in an oxygen-concentrated atmosphere. When such a lean NOx catalyst is used, if the NOx adsorption amount of the lean NOx catalyst increases during the lean operation, for example, as shown in the above cited reference 1, additional fuel is added during the expansion stroke in addition to the main combustion. By injecting, the air-fuel ratio of the exhaust gas is enriched and CO is generated, thereby promoting the separation and reduction of NOx.
[0004]
An engine that performs lean operation as described above requires the lean NOx catalyst in order to ensure NOx purification performance during lean operation. Further, a three-way catalyst is also required for exhaust purification in a region operated at a stoichiometric air-fuel ratio such as a high load region, the lean NOx catalyst is provided in addition to the three-way catalyst, and the lean A NOx catalyst requires a relatively large capacity in order to secure a certain amount of NOx adsorption, and is expensive compared to a three-way catalyst, which is disadvantageous in terms of cost.
[0005]
Moreover, in order to maintain the purification performance of the lean NOx catalyst, the NOx is temporarily removed by the removal of NOx, the supply of additional fuel for reduction, etc. at predetermined intervals such that the NOx adsorption amount increases as described above. It is necessary to enrich the air-fuel ratio, which reduces the fuel efficiency improvement effect due to lean combustion.
[0006]
In response to such a problem, the applicant of the present application is a multi-cylinder engine that performs a cycle consisting of intake, compression, expansion, and exhaust strokes. The burned gas discharged from the preceding cylinder, which is the exhaust stroke side cylinder, is introduced as it is into the succeeding cylinder, which is the cylinder on the intake stroke side, and the gas discharged from this succeeding cylinder is exhausted with a three-way catalyst. When the two cylinders are in a connected state, combustion is performed in a state where the preceding cylinder has a lean air-fuel ratio larger than the theoretical air-fuel ratio by a predetermined amount, and the succeeding cylinder is introduced from the preceding cylinder. It has been considered to control the combustion state so that combustion is performed in a state where the fuel is supplied to the burned gas having a lean air-fuel ratio to achieve the stoichiometric air-fuel ratio (Japanese Patent Application No. 2002-024548). .
[0007]
According to this, at least in the low-load low-rotation region, combustion at a lean air-fuel ratio is performed in the preceding cylinder, and the heat efficiency is increased and the pumping loss is reduced. In the succeeding cylinder, fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder and the stoichiometric air-fuel ratio is set, so that fuel efficiency is obtained by reducing pumping loss. Moreover, since only the stoichiometric burned gas discharged from the subsequent cylinders is guided to the exhaust passage provided with the three-way catalyst, sufficient exhaust purification performance is ensured with only the three-way catalyst, and no lean NOx catalyst is required. Become.
[0008]
Further, if the combustion in the subsequent cylinder is caused by compression ignition in which the combustion chamber is heated to high pressure and high pressure in the latter half of the compression stroke so that the fuel is self-ignited, the entire combustion chamber burns at once, so it does not contribute to work. Slow combustion is avoided, which is more advantageous for improving fuel economy. In order to compress and ignite the subsequent cylinders, it is necessary to set the combustion chamber to a high temperature and high pressure. By setting the two cylinders in a connected state as described above, high-temperature burned gas is introduced from the preceding cylinders into the subsequent cylinders. Therefore, compression ignition is possible.
[0009]
[Patent Document 1]
Japanese Patent Laid-Open No. 10-274085
[0010]
[Problems to be solved by the invention]
As described above, in order to obtain a significant fuel efficiency improvement effect while ensuring sufficient exhaust purification performance with only a three-way catalyst, combustion control is performed in a connected state of two cylinders, particularly combustion by compression ignition. Is effective. In order to perform compression ignition in the succeeding cylinder, the combustion chamber must be at a high temperature and a high pressure. Conversely, if the temperature is too high, abnormal combustion such as knocking may occur. Therefore, it is necessary that the operation region in which the compression ignition can be properly performed is within a certain range such that the combustion chamber has a suitable temperature and pressure. In order to obtain a greater fuel economy improvement effect, it has been desired to expand the range.
[0011]
The present invention has been made in consideration of the above-mentioned problems, and while achieving sufficient exhaust purification performance with only a three-way catalyst, it achieves fuel efficiency improvement effects such as lean combustion and reduction of pumping loss, and its contribution It is an object of the present invention to provide a control device for a spark ignition engine capable of obtaining a greater fuel efficiency improvement effect by expanding a high operating range.
[0012]
[Means for Solving the Problems]
According to a first aspect of the present invention, there is provided a multi-cylinder spark ignition engine in which each cylinder performs a cycle consisting of intake, compression, expansion, and exhaust strokes with a predetermined phase difference. Thus, between the pair of cylinders in which the exhaust stroke and the intake stroke overlap, the burned gas discharged from the preceding cylinder which is the cylinder on the exhaust stroke side is directly introduced to the subsequent cylinder which is the cylinder on the intake stroke side via the inter-cylinder gas passage. The exhaust gas discharged from the succeeding cylinder is configured to have a gas flow path in a two-cylinder connection state in which the exhaust gas is guided to the exhaust passage. Combustion is performed by forced ignition in a state where the lean air-fuel ratio is large by a predetermined amount. In the succeeding cylinder, the burned gas having the lean air-fuel ratio introduced from the preceding cylinder is converted to burned gas.An amount that is substantially the stoichiometric air-fuel ratioCombustion control means for controlling combustion of each cylinder so as to supply fuel and cause combustion by compression ignitionA preceding cylinder intake valve for introducing fresh air into the preceding cylinder, a preceding cylinder exhaust valve for discharging exhaust gas from the preceding cylinder to the exhaust passage, and introducing burned gas into the subsequent cylinder from the inter-cylinder gas passage A burnt gas introduction valve that performs the following cylinder intake valve that introduces fresh air into the subsequent cylinder,At least when the two cylinders are connectedIs because the preceding cylinder intake valve opens.New air introduced in the preceding cylinderAs the burned gas introduction valve opens,Burned gas is introduced from the inter-cylinder gas passage into the subsequent cylinder.On the other hand, when the two cylinders are connected, the preceding cylinder exhaust valve and the succeeding cylinder intake valve are closed,Among the operating regions in which the two cylinders are connected, at least in a predetermined region on the low load side, the interval between the intake stroke bottom dead center of the succeeding cylinder and the closing timing of the burned gas introduction valve is the same as that of the preceding cylinder. The spark ignition engine control apparatus is set to be shorter than an interval between an intake stroke bottom dead center and a closing timing of the preceding cylinder intake valve.
[0013]
According to this configuration, at least in the low-load low-rotation region, combustion is performed by forced ignition at a lean air-fuel ratio in which excessive air is present in the preceding cylinder. This lean combustion increases thermal efficiency and reduces pumping loss. Significant fuel economy effect can be obtained. In the succeeding cylinder, additional fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder, and combustion is performed by compression ignition. Since the gas introduced from the preceding cylinder through the inter-cylinder gas passage is hot, vaporization of the additional fuel is promoted and the entire combustion chamber is combusted all at once by compression ignition, so slow combustion that does not contribute to work Can be avoided and a high fuel efficiency improvement effect can be obtained. In addition, the NOx generation amount is suppressed to be relatively small by performing combustion at a lean air-fuel ratio in the preceding cylinder, and a large amount of EGR (exhaust gas recirculation) is introduced in the succeeding cylinder by introducing burned gas from the preceding cylinder. Since the state is equivalent to that performed, the generation of NOx is sufficiently suppressed, and exhaust gas purification is promoted.
[0014]
In addition, since only the exhaust gas generated by combustion at the stoichiometric air-fuel ratio discharged from the succeeding cylinder is guided to the exhaust passage, sufficient exhaust purification performance can be ensured only by providing a three-way catalyst in the exhaust passage. While performing lean combustion with high thermal efficiency in the preceding cylinder, a relatively expensive lean NOx catalyst, which has been conventionally required when performing lean combustion, is not required, and thus the cost can be reduced.
[0015]
Further, at least in a predetermined region on the low load side in the operation region in which the two cylinders are connected, the interval between the intake stroke bottom dead center of the subsequent cylinder and the closing timing of the burned gas introduction valve is set to be equal to that of the preceding cylinder. It is set to be shorter than the interval between the intake stroke bottom dead center and the closing timing of the preceding cylinder intake valve. That is, the closing timing of the burned gas introduction valve (which is a kind of intake valve) of the subsequent cylinder is closed earlier (closer to the bottom dead center) than when fresh air is sucked (preceding cylinder). ing. Therefore, the effective compression ratio in the subsequent cylinder can be increased.
[0016]
In the conventional engine, the closing timing of the intake valve is generally delayed from the bottom dead center of the intake stroke (crank angle is about 50 °). This is in order to compensate for the delay in inhalation due to the inhalation resistance, and the new air charging efficiency is improved by closing after the bottom dead center. However, on the other hand, since the intake valve is open while the piston passes the bottom dead center and rises, the effective compression ratio is smaller than the geometric compression ratio. According to this configuration, the closing timing of the burned gas introduction valve can be made closer to the bottom dead center, so that the effective compression ratio can be made closer to the geometric compression ratio, that is, increased. At this time, it is not the fresh air throttled by the throttle valve that is introduced into the subsequent cylinder, but the burned gas discharged from the preceding cylinder. Therefore, even if the burned gas introduction valve is closed early, its filling is sufficient. Made.
[0017]
When the effective compression ratio of the succeeding cylinder is increased, the in-cylinder temperature is increased, so that compression ignition is easily performed. Therefore, even in the low load region where the in-cylinder temperature does not easily rise and the compression ignitability is low, the compression ignitability can be improved. In other words, since the operating range in which combustion by compression ignition can be performed can be expanded to a lower load range, fuel efficiency can be further improved and exhaust gas purification can be promoted.
[0018]
Note that each stroke, top dead center (TDC), bottom dead center (BDC), and the like refer to the operation and position of the piston, but the piston position is also a function of the crank angle. In the present specification, the strokes and piston positions are also represented by the crank angle. For example, when the piston moves from the top dead center to the bottom dead center and the crankshaft rotates 180 °, the interval between the top dead center and the bottom dead center is a crank angle of 180 ° (hereinafter referred to as 180 ° CA). . The unit of the timing and period when the intake valve is opened or the intake valve is opened is the crank angle.
[0019]
According to a second aspect of the present invention, in the control device for the spark ignition engine according to the first aspect, there is provided a subsequent cylinder exhaust valve that discharges exhaust gas of the subsequent cylinder when at least the two cylinders are connected. In the operating region in which the cylinder is connected, at least in the predetermined region on the low load side, the succeeding cylinder exhaust valve is opened until the exhaust stroke top dead center of the succeeding cylinder, and the burned gas introduction valve is opened. The valve timing is set so as to be the top dead center of the intake stroke of the subsequent cylinder.
[0020]
In this way, it is possible to shorten the period (so-called valve overlap) in which the subsequent cylinder exhaust valve and the burned gas introduction valve are both open near the top dead center of the exhaust stroke of the subsequent cylinder (zero). Including the case). The conventional general setting is that the opening timing of the intake valve is earlier than the intake stroke top dead center (about 5 to 10 ° CA), but in this configuration, it is matched to the intake stroke top dead center. It is. By shortening the valve overlap in this way, the burned gas introduced into the succeeding cylinder is directly discharged into the exhaust passage through the succeeding cylinder exhaust valve, so-called blow-through is prevented, and the effective compression rate of the succeeding cylinder is reduced. Can be increased. As the effective compression ratio increases, the compression ignitability improves, so that the fuel consumption can be further improved and exhaust gas purification can be promoted as described above.
[0021]
According to a third aspect of the present invention, in the control device for the spark ignition type engine according to the first or second aspect, the burned gas introduction valve is controlled in a predetermined region on the high load side in the operation region where the two cylinders are connected. The valve closing timing is set to be delayed from the timing when the valve closing timing is in a predetermined region on the low load side.
[0022]
In this case, the load is relatively high, the in-cylinder temperature is high, and the compression ignitability is high. However, the in-cylinder temperature is excessively high and the abnormal combustion such as knocking may occur. By delaying the closing timing of the burned gas introduction valve, the effective compression ratio of the succeeding cylinder can be reduced. If the effective compression ratio of the succeeding cylinder is reduced, the in-cylinder temperature is lowered, so that abnormal combustion such as knocking can be prevented. Therefore, the operating range in which combustion by compression ignition can be performed without accompanying abnormal combustion can be expanded to a higher load side, fuel efficiency can be further improved, and exhaust gas purification can be promoted.
[0023]
According to a fourth aspect of the present invention, in the control device for the spark ignition engine according to the first or second aspect, the burnt gas introduction is performed in a predetermined region on the high load high rotation side in the operation region where the two cylinders are connected. The valve closing timing is set to a delay side with respect to the timing when the valve is in a predetermined region on the low load low rotation side.
[0024]
In this way, in addition to the invention of claim 3, it is possible to perform control in consideration of the rotational speed. Since the in-cylinder temperature also changes depending on the number of rotations, more detailed and appropriate compression ignition performance can be controlled by considering the number of rotations as a control condition.
[0025]
According to a fifth aspect of the present invention, in the spark ignition engine control device according to any one of the first to fourth aspects, when the at least two cylinders are connected, the burned gas of the preceding cylinder is transferred to the cylinder. In the operation region where the burned gas discharge valve for discharging to the intermediate gas passage is provided and the two cylinders are connected, the closing timing of the burned gas discharging valve is higher than the closing timing of the burned gas introduction valve. And set the open period of the burned gas discharge valve and the open period of the burned gas introduction valve.Each individuallyWhile maintaining a constant predetermined value, the opening timing of the burned gas discharge valve and the opening timing of the burned gas introduction valve fluctuate back and forth while keeping the timing difference constant according to at least the engine load. It is set as follows.
[0026]
In this way, the closing timing of the burned gas discharge valve can be changed back and forth according to the engine load, so the closing timing of the burned gas introduction valve is near the bottom dead center when the load is relatively low. And the effective compression ratio can be increased.
[0027]
At this time, the closing timing of the burned gas discharge valve is ahead of the top dead center because it is ahead of the closing timing of the burned gas introduction valve. This is earlier than a general exhaust valve closing timing of a conventional engine (about 50 ° CA after top dead center). By closing the burned gas discharge valve early, the remaining amount of burned gas in the preceding cylinder increases (hereinafter referred to as increase in internal EGR). When the internal EGR of the preceding cylinder increases, the in-cylinder temperature rises, so the temperature of the burned gas discharged from the preceding cylinder and introduced into the succeeding cylinder rises.
[0028]
That is, the effective compression ratio of the subsequent cylinder can be increased, and the temperature of the burned gas introduced into the subsequent cylinder can be increased. As a result, the in-cylinder temperature of the subsequent cylinder can be raised to improve the compression ignitability. In this way, the operating range in which combustion by compression ignition can be performed can be expanded to a lower load range, so that fuel efficiency can be further improved and exhaust gas purification can be promoted.
[0029]
On the other hand, when the load is relatively high, the valve opening period is generally delayed so that the closing timing of the burned gas introduction valve is delayed from the bottom dead center of the intake stroke, and the closing timing of the burned gas discharge valve is set. Can be delayed. The effective compression ratio of the succeeding cylinder can be reduced by closing the burned gas introduction valve after the bottom dead center. The burned gas discharge valveOpenBy delaying the valve timing, the internal EGR of the preceding cylinder can be reduced, and the temperature of the burned gas introduced into the succeeding cylinder can be lowered.
[0030]
That is, the effective compression ratio of the succeeding cylinder can be reduced, and the burnt gas temperature introduced into the succeeding cylinder can be lowered. As a result, it is possible to prevent the in-cylinder temperature of the subsequent cylinder from becoming excessively high and causing abnormal combustion. Therefore, the operating range in which combustion by compression ignition can be performed without accompanying abnormal combustion can be expanded to a higher load side, fuel efficiency can be further improved, and exhaust gas purification can be promoted.
[0031]
When changing the opening and closing timing of each valve, the opening period of the burned gas discharge valve and the opening period of the burned gas introduction valveEach individuallyWhile maintaining a constant predetermined value, the timing difference between these valve opening timings is kept constant. In this way, in an engine having a structure in which the opening / closing timing of each valve is uniquely set by the shape of each cam related to the opening / closing of each valve, it is not necessary to switch the cam, etc. Can always use the same cam. The valve opening timing can be changed by changing the phase (relative angle) between the camshaft that rotates integrally with the cam related to the opening and closing of the valve and the crankshaft. This structure can be made simpler than that in which each cam is independently controlled, and a small, light, and low-cost engine can be obtained.
[0032]
According to a sixth aspect of the present invention, there is provided a multi-cylinder spark ignition engine in which each cylinder performs a cycle comprising intake, compression, expansion, and exhaust strokes with a predetermined phase difference. Thus, between the pair of cylinders in which the exhaust stroke and the intake stroke overlap, the burned gas discharged from the preceding cylinder which is the cylinder on the exhaust stroke side is directly introduced to the subsequent cylinder which is the cylinder on the intake stroke side via the inter-cylinder gas passage. The exhaust gas discharged from the succeeding cylinder is configured to have a gas flow path in a two-cylinder connection state in which the exhaust gas is guided to the exhaust passage. Combustion is performed by forced ignition in a state where the lean air-fuel ratio is large by a predetermined amount. In the succeeding cylinder, the burned gas having the lean air-fuel ratio introduced from the preceding cylinder is converted to burned gas.An amount that is substantially the stoichiometric air-fuel ratioCombustion control means for controlling combustion of each cylinder so as to supply fuel and cause combustion by compression ignitionA preceding cylinder intake valve for introducing fresh air into the preceding cylinder, a preceding cylinder exhaust valve for discharging exhaust gas from the preceding cylinder to the exhaust passage, and introducing burned gas into the subsequent cylinder from the inter-cylinder gas passage A burnt gas introduction valve that performs the following cylinder intake valve that introduces fresh air into the subsequent cylinder,At least when the two cylinders are connectedIs because the preceding cylinder intake valve opens.New air introduced in the preceding cylinderAs the burned gas introduction valve opens,Burned gas is introduced from the inter-cylinder gas passage into the subsequent cylinder.On the other hand, when the two cylinders are connected, the preceding cylinder exhaust valve and the succeeding cylinder intake valve are closed,The burned gas introduction valve opening period in at least a predetermined region on the low load side of the operation region in which the two cylinders are connected.Is onShorter than the opening period of the leading cylinder intake valveAnd the interval between the intake stroke bottom dead center of the succeeding cylinder and the closing timing of the burned gas introduction valve is greater than the interval between the intake stroke bottom dead center of the preceding cylinder and the valve closing timing of the preceding cylinder intake valve. Also becomes shorterThe spark ignition type engine control device is characterized in that it is set as described above.
[0033]
According to this configuration, similarly to the effect of the first aspect, in the preceding cylinder, the thermal efficiency is increased by lean combustion, the pumping loss is reduced, and a significant fuel efficiency effect is obtained, and the succeeding cylinder performs combustion by compression ignition. As a result, a higher fuel efficiency improvement effect can be obtained. In addition, the exhaust purification performance is sufficiently secured while being low-cost only by providing the three-way catalyst.
[0034]
Furthermore, when the vehicle is traveling at a relatively low load, the burned gas introduction valve opening period of the subsequent cylinderFirstShorter than the opening period of the row cylinder intake valveAnd the closing timing of the burned gas introduction valve is closer to the intake bottom dead center than the closing timing of the preceding cylinder intake valveTherefore, the timing for closing the burned gas introduction valve can be relatively advanced, and the effective compression ratio of the succeeding cylinder can be increased. When the effective compression ratio of the succeeding cylinder is increased, the in-cylinder temperature is increased, so that compression ignition is easily performed. Therefore, even in the low load region where the in-cylinder temperature is unlikely to rise and the compression ignitability is low, the compression ignitability can be improved, and as described above, fuel efficiency can be further improved and exhaust gas purification can be promoted. Can do.
[0035]
According to a seventh aspect of the present invention, in the control device for the spark ignition engine according to the sixth aspect, a succeeding cylinder exhaust valve is provided for exhausting exhaust gas from the succeeding cylinder when at least the two cylinders are connected. Of the operating region in which the cylinder is connected, at least in a predetermined region on the low load side, the open period of the burned gas introduction valve is set to be shorter than the open period of the subsequent cylinder exhaust valve. It is characterized by that.
[0037]
thisAs a result, the closing timing of the burned gas introduction valve is relatively advanced in any case. Therefore, the effective compression ratio of the subsequent cylinder can be increased, and as described above, the fuel consumption can be further improved and exhaust gas purification can be promoted.
[0038]
Claim8The invention of claim6In the control apparatus for a spark ignition engine described above, a burned gas discharge valve for discharging burned gas of the preceding cylinder to the inter-cylinder gas passage when at least the two cylinders are connected is provided. The burned gas introduction valve opening timing is set to be delayed from the burned gas discharge valve opening timing in at least a predetermined region on the low load side in the operation range to be connected. It is characterized by.
[0039]
In this way, the closing timing of the burned gas introduction valve can be set near the bottom dead center of the intake stroke, so that the effective compression ratio of the succeeding cylinder can be increased, and fuel efficiency is further improved as described above. And exhaust gas purification can be promoted.
[0040]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings. FIG. 1 shows a schematic configuration of an engine according to a first embodiment of the present invention, and FIG. 2 schematically shows a structure of one cylinder of an engine body 1 and intake / exhaust valves provided for the cylinder. . In these drawings, the engine body 1 has a plurality of cylinders, and in the illustrated embodiment, has four cylinders 2A to 2D. A piston 3 is fitted into each of the cylinders 2 </ b> A to 2 </ b> D, and a combustion chamber 4 is formed above the piston 3.
[0041]
A spark plug 7 is provided at the top of the combustion chamber 4 of each cylinder 2 </ b> A to 2 </ b> D, and the tip of the plug faces the combustion chamber 4. An ignition circuit 8 capable of controlling the ignition timing by electronic control is connected to the spark plug 7.
[0042]
A fuel injection valve 9 that directly injects fuel into the combustion chamber 4 is provided at a side portion of the combustion chamber 4. The fuel injection valve 9 includes a needle valve and a solenoid (not shown). When a pulse signal is input, the fuel injection valve 9 is driven and opened for a time corresponding to the pulse width at the pulse input timing. It is comprised so that the quantity of fuel according to may be injected. The fuel injection valve 9 is supplied with fuel by a fuel pump (not shown) through a fuel supply passage and the like, and is supplied with fuel higher than the pressure in the combustion chamber during the compression stroke. A system is configured.
[0043]
Further, intake ports 11, 11a, 11b and exhaust ports 12, 12a, 12b are opened to the combustion chambers 4 of the respective cylinders 2A to 2D, and an intake passage 15 and an exhaust passage 20 are connected to these ports. Each port has a preceding cylinder intake valve 31, a subsequent cylinder intake valve 31a, a burned gas introduction valve 31b, a subsequent cylinder exhaust valve 32, a preceding cylinder exhaust valve 32a, and a burned gas discharge valve 32b (these valves are conventional engines). Equivalent to the intake and exhaust valves). Furthermore, camshafts 33 and 34 and cams 26 and 27 that rotate integrally therewith are provided above them.
[0044]
Each cylinder performs a cycle consisting of intake, compression, expansion, and exhaust strokes with a predetermined phase difference. In the case of a four-cylinder engine, the first cylinder 2A, second cylinder from one end in the cylinder row direction When the cylinder 2B, the third cylinder 2C, and the fourth cylinder 2D are called, as shown in FIG. 5, the cycle is 180 degrees at a crank angle in the order of the first cylinder 2A, the third cylinder 2C, the fourth cylinder 2D, and the second cylinder 2B. It is performed with a phase difference of °. In FIG. 5, EX is an exhaust stroke, IN is an intake stroke, F is fuel injection, S is forced ignition, and the star mark in the figure is compression ignition (forced ignition depending on conditions). Represents that.
[0045]
Between a pair of cylinders in which the exhaust stroke and the intake stroke overlap, a cylinder on the intake stroke side (referred to herein as a preceding cylinder) from the cylinder on the exhaust stroke side when the exhaust stroke and the intake stroke overlap (this specification) The inter-cylinder gas passage 22 is provided so that the burned gas can be directly introduced to the subsequent cylinder). In the four-cylinder engine of this embodiment, as shown in FIG. 5, the exhaust stroke (EX) of the first cylinder 2A and the intake stroke (IN) of the second cylinder 2B overlap, and the exhaust stroke (EX) of the fourth cylinder 2D. ) And the intake stroke (IN) of the third cylinder 2C overlap, so that the first cylinder 2A and the second cylinder 2B, and the fourth cylinder 2D and the third cylinder 2C form a pair, respectively, and the first cylinder 2A and the fourth cylinder The cylinder 2D is the preceding cylinder, the second cylinder 2B, and the third cylinder 2C are the subsequent cylinders.
[0046]
The intake / exhaust port of each cylinder and the intake passage, exhaust passage, and inter-cylinder gas passage connected to the cylinder are specifically configured as follows.
[0047]
The first cylinder 2A and the fourth cylinder 2D, which are the preceding cylinders, respectively include an intake port 11 for introducing fresh air, and a first exhaust port 12a for sending burned gas (exhaust gas) to the exhaust passage. A second exhaust port 12b for leading the burned gas to the subsequent cylinder is provided. The second cylinder 2B and the third cylinder 2C, which are the subsequent cylinders, respectively, have a first intake port 11a for introducing fresh air and a second intake port for introducing burned gas from the preceding cylinder. 11b and an exhaust port 12 for sending burned gas to the exhaust passage.
[0048]
In the example shown in FIG. 1, the intake ports 11 in the first and fourth cylinders 2A and 2D and the first intake ports 11a in the second and third cylinders 2B and 2C are two per cylinder, the left half of the combustion chamber. The first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders 2A and 2D and the second intake port 11b and the exhaust port in the second and third cylinders 2B and 2C are provided in parallel on the part side. 12 are provided in parallel on the right half side of the combustion chamber.
[0049]
The intake port 11 in the first and fourth cylinders 2A and 2D and the first intake port 11a in the second and third cylinders 2B and 2C are connected to the downstream ends of the branch intake passages 16 for each cylinder in the intake passage 15. Yes. In the vicinity of the downstream end of each branch intake passage 16, a multiple throttle valve 17 that is linked to each other via a common shaft is provided. This multiple throttle valve 17 is driven by an actuator 18 in accordance with a control signal, The intake air amount is adjusted. Note that an air flow sensor 19 that detects an intake air flow rate is provided in a common intake passage upstream of the collecting portion in the intake passage 15.
[0050]
An upstream end of a branch exhaust passage 21 for each cylinder in the exhaust passage 20 is connected to the first exhaust port 12a in the first and fourth cylinders 2A and 2D and the exhaust port 12 in the second and third cylinders 2B and 2C. Yes. Further, an inter-cylinder gas passage 22 is provided between the first cylinder 2A and the second cylinder 2B and between the third cylinder 2C and the fourth cylinder 2D, and the first, fourth cylinder 2A, The upstream end of the inter-cylinder gas passage 22 is connected to the 2D second exhaust port 12b, and the downstream end of the inter-cylinder gas passage 22 is connected to the second intake port 11b of the second and third cylinders 2B and 2C as the subsequent cylinders. Is connected.
[0051]
The inter-cylinder gas passage 22 has a linear O whose output changes linearly according to the oxygen concentration.₂A sensor 25 is provided, and the fuel injection amount for the preceding cylinders 2A and 2D having a predetermined lean air-fuel ratio is feedback-controlled according to the output thereof.
[0052]
In the exhaust passage 20, downstream of the branch exhaust passage 21, an air-fuel ratio is detected by detecting the oxygen concentration in the exhaust gas.₂A sensor 23 is provided. O₂The sensor 23 outputs λO whose output changes suddenly near the theoretical air-fuel ratio.₂This is a sensor.₂Based on the output of the sensor 23, the fuel injection amount for the succeeding cylinders 2B and 2C (including the cylinders 2A and 2D when each cylinder is in an independent state) is feedback-controlled. Furthermore O₂A three-way catalyst 24 for purifying exhaust gas is provided in the exhaust passage 20 downstream of the sensor 23. As is generally known, the three-way catalyst 24 is highly purified against HC, CO, and NOx when the air-fuel ratio of the exhaust gas is close to the stoichiometric air-fuel ratio (that is, the excess air ratio λ is λ = 1). It is a catalyst showing performance.
[0053]
The valves for opening and closing the intake / exhaust ports of each cylinder and the valve operating mechanism for these valves are as follows.
[0054]
The intake port 11, the first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders 2A and 2D are respectively provided with a preceding cylinder intake valve 31, a preceding cylinder exhaust valve 32a and a burned gas discharge valve 32b. Further, the first intake port 11a, the second intake port 11b, and the exhaust port 12 in the second and third cylinders 2B and 2C are respectively provided with a subsequent cylinder intake valve 31a, a burned gas introduction valve 31b, and a subsequent cylinder exhaust valve 32. It has been. Each of these valves is opened by being pushed down by the cams 26 and 27 when each cylinder is in the intake stroke or the exhaust stroke, but the opening / closing timing is not necessarily limited to the top dead center or the bottom dead center. It is set at a time shifted by degrees CA (crank angle) to several tens of degrees CA.
[0055]
Furthermore, in this embodiment, the opening / closing timing of each valve is changed according to the conditions by the cam phase variable mechanisms 33a and 34a, or the valve stop mechanism 35 is closed.
[0056]
The cam phase variable mechanisms 33a and 34a are conventionally known mechanisms that vary the rotational phase of the camshafts 33 and 34 with respect to the rotational phase of the crankshaft. As shown in FIG. 1, the camshaft 33 is provided with a cam phase varying mechanism 33a, and the camshaft 34 is provided with a cam phase varying mechanism 34a, which are controlled independently (see FIG. 3). Accordingly, the opening / closing timings of the preceding cylinder intake valve 31 and the succeeding cylinder intake valve 31a that are opened / closed by the rotation of the camshaft 33 are entirely changed back and forth by the cam phase variable mechanism 33a. Similarly, the opening / closing timings of the burned gas introduction valve 31b, the succeeding cylinder exhaust valve 32, the preceding cylinder exhaust valve 32a, and the burned gas discharge valve 32b that are opened and closed by the rotation of the camshaft 34 are entirely controlled by the cam phase variable mechanism 34a. Fluctuate back and forth.
[0057]
The valve stop mechanism 35 switches each valve between an operating state and a stopped state, and is provided in the succeeding cylinder intake valve 31a, the burned gas introduction valve 31b, the preceding cylinder exhaust valve 32a, and the burned gas discharge valve 32b. . Since the mechanism is conventionally known and detailed illustration is omitted, for example, a hydraulic chamber capable of supplying and discharging hydraulic oil is provided in a tappet interposed between a cam and a valve shaft. When hydraulic oil is supplied to the valve, the cam operation is transmitted to the valve and the valve is opened and closed.When the hydraulic oil is discharged from the hydraulic chamber, the cam operation is not transmitted to the valve and the valve is stopped. It is something that has come to be. A first control valve 37 is provided in the hydraulic oil supply / discharge passage 36 to the valve stop mechanism 35 of the succeeding cylinder intake valve 31a and the preceding cylinder exhaust valve 32a, and the burned gas introduction valve 31b and the burned gas discharge are provided. A second control valve 39 is provided in the hydraulic oil supply / discharge passage 38 with respect to the valve 32b (see FIG. 3).
[0058]
FIG. 3 shows the configuration of the drive and control system in this embodiment. In this figure, an ECU (control unit) 40 for engine control composed of a microcomputer or the like is provided with an air flow sensor 19, O₂Sensor 23 and linear O₂A signal from the sensor 25 is input, and a signal from an engine speed sensor 47 for detecting the engine speed and an accelerator position sensor 48 for detecting the accelerator position (depressing amount of the accelerator pedal) are input in order to determine the driving state. Is done. Control signals are sent from the ECU 40 to the ignition circuit 8, the fuel injection valves 9, the actuator 18 of the multiple throttle valve 17, the first and second control valves 37 and 39, and the cam phase variable mechanisms 33a and 34a. Is output.
[0059]
The ECU 40 constitutes a control means for performing combustion while keeping the gas flow path in a two-cylinder connected state (see FIG. 6) at least in a low-load low-rotation region, and includes an operating state determination means 41, a valve stop mechanism. Control means 42, intake air amount control means 43, combustion control means 44, and cam phase control means 49 are provided.
[0060]
The operating state discriminating means 41 checks the operating state of the engine (engine speed and engine load) by signals from the rotational speed sensor 47 and the accelerator opening sensor 48, and the operating state is a low load and low speed as shown in FIG. Side operation area A (engine load T1 or less and engine speed r1 or less) or high load side or high rotation side operation area B (engine load exceeds T1 or engine speed exceeds r1) It is determined whether it is in the area. Among the operation areas A, the operation area A1 is a relatively low load and low rotation area, and the operation area A2 is a relatively high load and high rotation area. Under a predetermined condition (for example, a state where the engine is completely warmed up), the operation region A is operated in the special operation mode in which the two cylinders are connected, and in the operation region B, the normal operation mode in which each cylinder is in an independent state. Drive in.
[0061]
The valve stop mechanism control means 42 controls each of the valve stop mechanisms 35 as follows by controlling the control valves 37 and 39 according to the special operation mode and the normal operation mode.
[0062]

[0063]
The intake air amount control means 43 controls the opening degree (throttle opening degree) of the throttle valve 17 by controlling the actuator 18, and obtains the target intake air amount from a map or the like according to the operating state, and the target The throttle opening is controlled according to the intake air amount. In this case, in the special operation mode, as will be described later, in the succeeding cylinders (second and third cylinders 2B and 2C), in the gas introduced from the preceding cylinder in a state where the intake introduction from the branch intake passage 16 is blocked. Combustion is performed while the ratio of excess air to the newly supplied fuel is the stoichiometric air-fuel ratio (hereinafter referred to as a substantial stoichiometric air-fuel ratio), so that it corresponds to the required torque for the preceding and succeeding two cylinders. The amount of air necessary for the combustion of fuel (the amount of air that is the stoichiometric air-fuel ratio with respect to the amount of fuel for two cylinders) is supplied to the preceding cylinders (first, fourth cylinders 2A, 2D), The throttle opening is adjusted.
[0064]
The combustion control unit 44 includes a fuel injection control unit 45 and an ignition control unit 46. The fuel injection control unit 45 causes the fuel injection amount and the injection timing from the fuel injection valves 9 provided in the respective cylinders 2A to 2D. Is controlled according to the operating state of the engine, and the ignition control means 46 controls the ignition timing and the ignition stop according to the operating state. In particular, the combustion control (fuel injection control and ignition control) is changed depending on whether the operation state is the special operation mode or the normal operation mode.
[0065]
That is, in the special operation mode, the air-fuel ratio for the preceding cylinders (the first and fourth cylinders 2A and 2D) is a lean air-fuel ratio larger than the stoichiometric air-fuel ratio, preferably about twice the stoichiometric air-fuel ratio or The fuel injection amount is controlled as described above, the injection timing is set so that fuel is injected in the compression stroke and the mixture is stratified, and forced ignition is performed near the compression top dead center. The ignition timing is set as follows. On the other hand, for the subsequent cylinders (second and third cylinders 2B and 2C), fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder so that the actual stoichiometric air-fuel ratio is obtained. The fuel injection amount is controlled at the same time, and the injection timing is set so that the fuel is injected in the intake stroke. Then, combustion by compression ignition or forced ignition is performed according to the operating state.
[0066]
In the normal operation mode, the fuel injection amount is controlled so that the air-fuel ratio of each of the cylinders 2A to 2D is equal to or less than the stoichiometric air-fuel ratio. Then, the stoichiometric air-fuel ratio is made richer in the fully open load and the operation region in the vicinity thereof. In this case, the injection timing is set so that the air-fuel mixture is made uniform by injecting fuel to each of the cylinders 2A to 2D and the cylinders 2A to 2D are forcedly ignited. To.
[0067]
The cam phase control means 49 controls the cam phase variable mechanisms 33a and 34a based on the result of the operation state determination means 41. Although details of the control will be described later, for example, in the special operation mode, the cam phase variable mechanism 34a is controlled so as to advance the phase of the cam 27 in the region of relatively low load and low rotation (region A1 in FIG. 4). The open / close timing of the burned gas discharge valve 32b, the burned gas introduction valve 31b, and the subsequent cylinder exhaust valve 32 that are operated by the rotation of 34 is set so as to be advanced as a whole. On the other hand, in the region of relatively high load and high rotation (region A2 in FIG. 4), the cam phase varying mechanism 34a is controlled so as to delay the phase of the cam 27, and the burned gas discharge valve 32b is operated by the rotation of the camshaft 34. The opening / closing timing of the burned gas introduction valve 31b and the subsequent cylinder exhaust valve 32 is set to be delayed as a whole. Since the cam phase variable mechanisms 33a and 34a act on the operating valve, the valve stopped by the valve stop mechanism 35 is stopped regardless of the control of the cam phase variable mechanisms 33a and 34a. maintain.
[0068]
The operation of the apparatus of the present embodiment as described above will be described with reference to FIGS.
In the special operation mode, as described above, the preceding cylinder exhaust valve 32a and the succeeding cylinder intake valve 31a are stopped, and the burned gas discharge valve 32b and the burned gas introduction valve 31b are activated, so that substantially fresh air is generated. The gas flow path is as shown in FIG. 6, and the burned gas discharged from the preceding cylinders (first and fourth cylinders) 2A and 2D is directly passed through the inter-cylinder gas passage 22 to the subsequent cylinders (second and second cylinders). (Cylinder No. 3) is introduced into 2B and 2C, and a two-cylinder connection state is established in which only the gas discharged from the succeeding cylinders 2B and 2C is guided to the exhaust passage 20.
[0069]
In this state, fresh air is introduced into the preceding cylinders 2A and 2D from the intake passage 15 during the intake stroke (arrow a in FIG. 6), and linear O is detected in the preceding cylinders 2A and 2D.₂The fuel is injected in the compression stroke while the fuel injection amount is feedback controlled so that the air-fuel ratio detected by the sensor 25 becomes a super-lean air-fuel ratio approximately twice or more than the stoichiometric air-fuel ratio, and at a predetermined ignition timing. Ignition is performed, and stratified combustion is performed at a super lean air-fuel ratio (see FIG. 5).
[0070]
Thereafter, burned gas discharged from the preceding cylinders 2A and 2D is introduced into the succeeding cylinders 2B and 2C through the gas passage 22 during a period in which the intake strokes of the preceding cylinders 2A and 2D overlap with the exhaust strokes of the succeeding cylinders 2B and 2C. (The white arrow in FIG. 5 and the arrow b in FIG. 6). In the succeeding cylinders 2B and 2C, fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinders 2A and 2D, and the fuel injection amount is controlled so as to obtain a substantial stoichiometric air-fuel ratio. Fuel is injected during the intake stroke. At this time, in the succeeding cylinders 2B and 2C, forced ignition by the spark plug 7 is stopped, and compression ignition is performed near the top dead center of the compression stroke due to an increase in pressure and temperature in the combustion chamber.
[0071]
In this way, the subsequent cylinders 2B and 2C rapidly burn by simultaneous compression ignition even under the condition that they contain a large amount of burned gas components equivalent to EGR gas and the air-fuel ratio is lean, which greatly increases the thermal efficiency. Will be improved.
[0072]
As described above, the preceding cylinders 2A and 2D increase the thermal efficiency and reduce the pumping loss by stratified combustion in the ultra-lean state, while the succeeding cylinders 2B and 2C have the effect of reducing the pumping loss similarly to the preceding cylinders 2A and 2D. In addition, when combustion is performed by compression ignition, thermal efficiency is improved by performing compression ignition in a uniform mixture distribution state, and fuel efficiency is greatly improved by these actions.
[0073]
The preceding cylinders 2A and 2D have a lean air-fuel ratio that is approximately twice or more than the stoichiometric air-fuel ratio, so that the amount of NOx generated is relatively small. In the succeeding cylinders 2B and 2C, the preceding cylinders 2A and 2D Since the burned gas is introduced, the state is equivalent to that in which a large amount of EGR is performed, so that the generation of NOx is sufficiently suppressed. This is also advantageous for improving emissions.
[0074]
However, even in the special operation mode, when the in-cylinder temperature of the subsequent cylinders 2B and 2C is low and it is difficult to perform compression ignition, the ignition of the subsequent cylinders 2B and 2C is switched to forced ignition by the spark plug 7. Conversely, when the in-cylinder temperature of the succeeding cylinders 2B and 2C is too high and abnormal combustion such as knocking occurs, the special operation mode is switched to the normal operation mode. In any case, the improvement effect of the fuel consumption or the like is suppressed as compared with the special operation mode by compression ignition. Therefore, in order to obtain more of these effects, it is desirable to expand the operation region suitable for compression ignition during the special operation mode.
[0075]
In order to expand the operation range suitable for compression ignition in the special operation mode, the intake / exhaust valve opening / closing timing is set as follows.
[0076]
FIG. 8 shows in detail the intake / exhaust stroke portion of FIG. 5, and the opening / closing timings of the preceding cylinder intake valve 31 and the burned gas discharge valve 32b of the preceding cylinders 2A and 2D in the special operation mode, and the succeeding cylinder 2B. , 2C burned gas introduction valve 31b and the opening and closing timing of the subsequent cylinder exhaust valve 32. FIG. 8A shows the case of a relatively low-load low-rotation region (region A1 in FIG. 4) in the operation state in which the special operation mode is performed, and FIG. This is the case of the region (region A2 in FIG. 4). In these drawings, the horizontal axis indicates the crank angle, T is top dead center (TDC), and B is bottom dead center (BDC). The interval between T and B is 180 ° CA. The upper row shows the preceding cylinders 2A and 2D, and the lower row shows the subsequent cylinders 2B and 2C corresponding thereto. And each strip | belt-shaped part shows the valve opening period of each valve. The white arrows from the upper stage to the lower stage indicate that the exhaust strokes of the preceding cylinders 2A and 2D overlap the intake strokes of the subsequent cylinders 2B and 2C, and the burned gas in the preceding cylinders 2A and 2D is transferred to the subsequent cylinders 2B and 2C. Indicates the led state.
[0077]
In FIG. 8A, in the upper stage, the burned gas discharge valve opening period 80 in which the burned gas discharge valve 32b of the preceding cylinders 2A and 2D is opened, and the preceding cylinder intake in which the preceding cylinder intake valve 31 is opened. A valve opening period 81 (shown by hatching) is shown. The lower stage includes a valve opening period 82 of the subsequent cylinder exhaust valve in which the subsequent cylinder exhaust valve 32 of the subsequent cylinders 2B and 2C is opened, and a valve opening period 83 of the burned gas introduction valve in which the burned gas introduction valve 31b is opened. Indicates. The opening period 81 of the leading cylinder intake valve is set from about 10 ° CA before TDC to about 55 ° CA after BDC (total of about 245 ° CA). This is a general setting value for a conventional engine. On the other hand, the open period 83 of the burned gas introduction valve is set from approximately 45 ° CA before TDC to approximately BDC (total approximately 225 ° CA). Further, the valve opening period 82 of the subsequent cylinder exhaust valve is set from about 80 ° CA before BDC to about 25 ° CA before TDC (total of about 235 ° CA).
[0078]
In other words, the interval (approximately 0 ° CA) between the trailing cylinder intake stroke bottom dead center 96 and the burned gas introduction valve closing timing 97 is equal to the preceding cylinder intake stroke bottom dead center 92 and the preceding cylinder intake valve closing timing 93. Is set to be shorter than the interval (about 55 ° CA). The burned gas introduction valve opening period 83 is shorter than the preceding cylinder intake valve opening period 81 and the subsequent cylinder exhaust valve opening period 82. Thus, in the succeeding cylinders 2B and 2C, the burned gas introduction valve opening period 83 is set short, and the burned gas introduction valve closing timing 97 is set early and close to BDC. There is no or short period during which the burned gas introduction valve 31b is open even when the piston 3 has risen past the BDC. For this reason, the effective compression ratio increases and approaches the geometric compression ratio. As the effective compression ratio increases, the in-cylinder temperatures of the succeeding cylinders 2B and 2C are likely to rise, and the compression ignitability is increased. At this time, what is introduced into the succeeding cylinders 2B and 2C is not the fresh air throttled by the multiple throttle valve 17, but the burned gas discharged from the preceding cylinders 2A and 2D. Therefore, even if the burned gas introduction valve 31b is closed early, the filling is sufficiently performed.
[0079]
The open period 80 of the burned gas discharge valve is set from about 80 ° CA before BDC to about 25 ° CA before TDC (total of about 235 ° CA). Therefore, the closing timing 90 of the burned gas discharge valve is set to be approximately 25 ° CA advance from the closing timing 97 of the burned gas introduction valve (corresponding to TDC in the preceding cylinders 2A and 2D). This is earlier than a general set value (about 50 ° CA after TDC) of a conventional engine. By closing the burned gas discharge valve 32b early, the amount of internal EGR in the preceding cylinders 2A and 2D increases, and the temperature of the burned gas introduced into the succeeding cylinders 2B and 2C rises.
[0080]
As described above, the effective compression ratio of the succeeding cylinders 2B and 2C is increased, and the in-cylinder temperature of the succeeding cylinders 2B and 2C is increased by increasing the burnt gas temperature introduced into the succeeding cylinders 2B and 2C. To improve compression ignitability. By doing in this way, the operation area | region which can perform the combustion by compression ignition is expanded to the area | region of a low load.
[0081]
In FIG. 8B, in the upper stage, the burned gas discharge valve opening period 100 in which the burned gas discharge valve 32b of the preceding cylinders 2A, 2D is opened, and the preceding cylinder intake in which the preceding cylinder intake valve 31 is opened. A valve opening period 101 (indicated by diagonal lines) is shown. In the lower stage, the opening period 102 of the succeeding cylinder exhaust valve in which the succeeding cylinder exhaust valve 32 of the succeeding cylinders 2B and 2C is opened, and the opening period 103 of the burned gas introduction valve in which the burned gas introduction valve 31b is opened, Indicates. The opening period 101 of the preceding cylinder intake valve is set similarly to the opening period 81 of the preceding cylinder intake valve in FIG. On the other hand, the valve opening period 100 of the burned gas discharge valve, the valve opening period 102 of the subsequent cylinder exhaust valve, and the valve opening period 103 of the burned gas introduction valve (respectively indicated by white band lines) are shown in FIG. The length of each period is equal to the opening period 80 of the burned gas discharge valve, the opening period 82 of the subsequent cylinder exhaust valve, and the opening period 83 of the burned gas introduction valve, and only the opening / closing timing of each valve is entirely Therefore, the setting is delayed by about 45 ° CA. As described above, the cam phase control means 49 delays the phase of the cam shaft 34 by about 45 ° CA (see FIG. 3).
[0082]
The closing timing 114 of the burned gas introduction valve in FIG. 8B is delayed by about 45 ° CA from the closing timing 97 of the burned gas introduction valve in FIG. For this reason, in the succeeding cylinders 2B and 2C, the burned gas introduction valve 31b is open for about 45 ° CA even when the piston 3 is rising past the bottom cylinder dead center 113 of the succeeding cylinder intake stroke. Therefore, the effective compression ratio of the succeeding cylinders 2B and 2C is smaller than that in the case of FIG. 8A, and the in-cylinder temperature of the succeeding cylinders 2B and 2C is lowered.
[0083]
Moreover, the closing timing 112 of the burned gas discharge valve is about 45 ° CA later than the closing timing 90 of the burned gas discharge valve in FIG. 8A and is about 20 ° after the top dead center 111 of the preceding cylinder exhaust stroke. CA is set. For this reason, the internal EGR of the preceding cylinders 2A and 2D is smaller than that in the case of FIG. 8A, and the burnt gas temperature introduced into the succeeding cylinders 2B and 2C is lowered.
[0084]
As described above, the effective compression ratio of the succeeding cylinders 2B and 2C is reduced, and the in-cylinder temperature of the succeeding cylinders 2B and 2C is lowered by lowering the burnt gas temperature introduced into the succeeding cylinders 2B and 2C. The occurrence of abnormal combustion such as knocking is prevented. By doing in this way, the operation area | region which can perform the combustion by compression ignition is expanded to the area | region of a higher load.
[0085]
Thus, in the special operation mode, the in-cylinder temperature of the succeeding cylinders 2B and 2C is increased in the operation region where the in-cylinder temperature is relatively low, and is decreased in the operation region where the in-cylinder temperature is relatively high. The opening / closing timing of the intake / exhaust valve is set. As a result, the operating range in which the succeeding cylinders 2B and 2C can perform proper combustion by compression ignition can be expanded, and the fuel efficiency improvement effect and the exhaust gas purification can be further promoted.
[0086]
On the other hand, in the normal operation mode, as described above, the preceding cylinder exhaust valve 32a and the succeeding cylinder intake valve 31a are in an operating state, and the burned gas discharge valve 32b and the burned gas introduction valve 31b are stopped. The flow path of fresh air and gas is as shown in FIG. 7, and the intake ports 11 and 11a and the exhaust ports 12a and 12 of each cylinder 2A to 2D are substantially independent, and each cylinder 2A to 2D is connected from the intake passage 15 to each other. Fresh air is introduced into the intake ports 11 and 11a, and burned gas is discharged from the exhaust ports 12a and 12 of the cylinders 2A to 2D to the exhaust passage 20.
[0087]
FIG. 9 shows the opening / closing timing of the preceding cylinder intake valve 31 and the burned gas discharge valve 32b of the preceding cylinders 2A and 2D in the normal operation mode, and the burned gas introduction valve 31b and the succeeding cylinder exhaust valve 32 of the succeeding cylinders 2B and 2C. It is explanatory drawing which shows opening and closing time. The horizontal axis indicates the crank angle, T is top dead center (TDC), and B is bottom dead center (BDC). The interval between T and B is 180 ° CA. The upper row shows the preceding cylinders 2A and 2D, and the lower row shows the succeeding cylinders 2B and 2C. Since each cylinder is independently operated in the normal operation mode, fresh air is introduced into the preceding cylinders 2A and 2D and the succeeding cylinders 2B and 2C to perform combustion by forced ignition. In the upper part of FIG. 9, the burned gas exhaust valve opening period 120 in which the preceding cylinder exhaust valves 32 a of the preceding cylinders 2 </ b> A and 2 </ b> D are opened, and the preceding cylinder intake valve opening period in which the preceding cylinder intake valve 31 is opened are shown. 121. The lower stage shows a valve opening period 122 of the subsequent cylinder exhaust valve in which the subsequent cylinder exhaust valve 32 of the subsequent cylinders 2B and 2C is opened, and a valve opening period 123 of the subsequent cylinder intake valve in which the subsequent cylinder intake valve 31a is opened. .
[0088]
The exhaust in the preceding cylinders 2A and 2D and the intake in the subsequent cylinders 2B and 2C are performed by valves different from those in the special operation mode, and therefore the opening and closing are performed by different cams. Therefore, the opening period 120 of the preceding cylinder exhaust valve and the opening period 123 of the succeeding cylinder intake valve are the opening period 80 of the burned gas discharge valve and the opening period 83 of the burned gas introduction valve in FIG. Can be set independently. In FIG. 9, the valve opening period 121 of the preceding cylinder intake valve and the valve opening period 123 (shown by hatching) of the subsequent cylinder intake valve are from about 10 ° CA before TDC to about 55 ° CA after BDC (total of about 245 ° CA). Is set to Further, the opening period 120 of the preceding cylinder exhaust valve and the opening period 122 of the subsequent cylinder exhaust valve are set from about 30 ° CA before BDC to about 25 ° CA after TDC (total of about 235 ° CA). These valve opening periods correspond to general settings of a conventional engine.
[0089]
The opening period 121 of the preceding cylinder intake valve and the opening period 123 of the succeeding cylinder intake valve can be changed back and forth by the cam phase variable mechanism 33a, and the opening period 120 of the preceding cylinder exhaust valve and the succeeding cylinder exhaust valve are variable. The valve opening period 122 can be changed back and forth by the cam phase varying mechanism 34a. Therefore, by controlling the cam phase variable mechanisms 33a and 34a, the period during which both the preceding cylinder exhaust valve 32a and the preceding cylinder intake valve 31 are open (the opening period 120 of the preceding cylinder exhaust valve and the preceding cylinder intake valve). The valve overlap period (valve overlap: valve overlap) can be varied. The cam phase control means 49 controls the cam phase variable mechanisms 33a and 34a so that the valve overlap becomes larger as the load becomes higher, so that optimum combustion efficiency is obtained according to the load. The same control is performed for the succeeding cylinders 2B and 2C.
[0090]
In this way, in the normal operation mode, the optimum intake / exhaust timing is controlled according to the load, and the intake air amount and the fuel injection amount are controlled so as to become the stoichiometric air-fuel ratio or richer, so that the output Performance is ensured.
[0091]
Note that the patterns shown in FIGS. 8A, 8B, and 9 do not limit each valve opening period or valve opening timing, and may be changed as appropriate within the scope of the claims. For example, the closing timing 97 of the burned gas introduction valve in FIG. 8A may be set so as to be slightly behind the bottom dead center 96 of the subsequent cylinder intake stroke. Further, the valve opening periods 81, 101, and 121 of the preceding cylinder intake valve in FIGS. 8A, 8B, and 9 are set differently depending on the load (in this embodiment, this is done by the cam phase variable mechanism 33a. Or the same setting as in the present embodiment (in this case, the cam phase variable mechanism 33a may not be provided). Other specific set values may be set to suitable values according to the required characteristics of the engine.
[0092]
Further, the travel region A in the special operation mode may be divided into three or more instead of the two regions A1 and A2, and the valve opening / closing timing suitable for each region may be set. Further, it may be changed continuously instead of stepwise by division.
[0093]
Next, a second embodiment of the present invention will be described with reference to FIGS. In these drawings, the same parts as those in the first embodiment are denoted by the same reference numerals, and redundant description thereof is omitted.
[0094]
FIG. 10 is a partial perspective view of the cam switching mechanism 150 in the second embodiment, and is for the burned gas introduction valve 31b and the like indicated by a two-dot chain line. A camshaft 151 is disposed above the burned gas introduction valve 31b. The camshaft 151 is provided with three types of cams having independent lift characteristics, ie, a first cam 152, a second cam 154, and a third cam 156, so as to rotate integrally. A rocker arm set 160 supported by a rocker shaft 170 is provided between these cams and the burned gas introduction valve 31b. The rocker arm set 160 is an aggregate of three types of rocker arms, a first rocker arm 162, a second rocker arm 164, and a third rocker arm 166. A valve abutting portion 163 and an adjusting screw 161 for finely adjusting the axial direction position are provided at the tip of the first rocker arm 162. The valve abutting portion 163 is at an appropriate position at the burned gas introduction valve 31b. It contacts the upper end of the valve shaft. The second rocker arm 164 and the third rocker arm 166 are provided with springs (not shown) so as to press the rocker arms against the second cam 154 and the third cam 156. Therefore, when each rocker arm of the rocker arm set 160 is movable independently as shown in the figure, the upper surface of each rocker arm abuts on the outer periphery of the first cam 152, the second cam 154, and the third cam 156, The rocker shaft 170 swings up and down according to the shape of the cam contact portion (the rotation radius of each cam).
[0095]
Inside the rocker arm set 160, five plungers are provided in two rows as will be described later (see FIG. 11. In FIG. 10, the fourth plunger hole 204, which is one of the plunger holes, is visible). By the movement of these plungers, the first rocker arm 162 can be integrated with and interlocked with the second rocker arm 164 or the third rocker arm 166. Inside the rocker shaft 170, a first hydraulic oil supply / discharge passage 172 and a second hydraulic oil supply / discharge passage 174 for introducing oil for hydraulically operating the plunger are provided.
[0096]
The first cam 152 is a valve stop cam and has an outer peripheral shape concentric with the camshaft 151. Therefore, when the upper surface of the first rocker arm 162 is always in contact with the outer peripheral surface of the first cam 152 (separated from the second rocker arm 164 and the third rocker arm 166), the first rocker arm 162 does not swing even if the camshaft 151 rotates. Does not move. That is, the burned gas introduction valve 31b and the like are stopped in a closed state.
[0097]
The second cam 154 is a low-load (or low-speed) cam, and includes a portion having the same outer peripheral shape as the first cam 152 and a portion having an outer peripheral shape protruding therefrom. Therefore, when the upper surface of the second rocker arm 164 is always in contact with the outer peripheral surface of the second cam 154 (separated from the third rocker arm 166 described later), a predetermined crank angle is generated along with the rotation of the cam shaft 151. Swings downward by a predetermined amount. At this time, if the first rocker arm 162 and the second rocker arm 164 are interlocked, the operation of the first rocker arm 162 is the same as the swing of the second rocker arm 164 by the second cam 154. That is, the burnt gas introduction valve 31b is opened by a predetermined amount at a predetermined time.
[0098]
The third cam 156 is a high-load (or high-speed) cam, and includes a portion having the same outer peripheral shape as the second cam 154 and a portion having an outer peripheral shape protruding therefrom. Therefore, the upper surface of the third rocker arm 166 is always in contact with the outer peripheral surface of the third cam 156 and swings downward by a predetermined amount at a predetermined crank angle as the camshaft 151 rotates. At this time, if the first rocker arm 162 and the third rocker arm 166 are interlocked, the operation of the first rocker arm 162 is the same as the swing of the third rocker arm 166 by the third cam 156. That is, the burned gas introduction valve 31b is opened by a predetermined amount at a predetermined time (the valve opening period includes a valve opening period when only the second rocker arm 164 is interlocked with the first rocker arm 162).
[0099]
FIG. 11 is an explanatory view showing the operation of five plungers provided inside the rocker arm set 160. 11A shows a state in which the first rocker arm 162 is separated from the second rocker arm 164 and the third rocker arm 166, FIG. 11B shows a state in which the first rocker arm 162 is interlocked only with the second rocker arm 164, FIG. c) shows a state in which the first rocker arm 162 is interlocked with the second rocker arm 164 and the third rocker arm 166.
[0100]
A first plunger hole 201 and a fourth plunger hole 204 are provided inside the first rocker arm 162. The first plunger hole 201 is a concave hole having a circular cross section that opens to the second rocker arm 164 side. A first hydraulic oil introduction passage 173 is led from a first hydraulic oil supply / discharge passage 172 to the bottom of the first plunger hole 201. A cylindrical first plunger 181 is fitted in the first plunger hole 201. The first plunger 181 slides smoothly in the first plunger hole 201 while sealing the hydraulic oil guided to the first hydraulic oil introduction path 173 on the outer peripheral surface thereof. The total length of the first plunger 181 is shorter than the depth of the first plunger hole 201.
[0101]
The fourth plunger hole 204 is a through hole that passes through the second rocker arm 164 side and the third rocker arm 166 side. A columnar fourth plunger 184 is fitted in the fourth plunger hole 204. The total length of the fourth plunger 184 is equal to the depth of the fourth plunger hole 204 (plate thickness of the first rocker arm 162). The fourth plunger 184 slides smoothly in the fourth plunger hole 204.
[0102]
A second plunger hole 202 and a fifth plunger hole 205 are provided inside the second rocker arm 164. The second plunger hole 202 is a concave hole having a circular cross section that opens to the first rocker arm 162 side, and has the same diameter as the first plunger hole 201. An air vent hole 206 is provided at the bottom of the second plunger hole 202, and the internal air pressure is kept at atmospheric pressure while releasing leaked oil. A second plunger 182 having a bottomed cylindrical shape and an outer diameter equal to that of the first plunger 181 is fitted in the second plunger hole 202. The second plunger 182 slides smoothly in the second plunger hole 202. The total length of the second plunger 182 is equal to the depth of the second plunger hole 202. The end of the second plunger 182 that contacts the first plunger 181 is formed in a spherical shape. A second plunger spring 187 is provided in the inner concave portion of the second plunger 182, and always biases the second plunger 182 toward the first plunger 181 side.
[0103]
The fifth plunger hole 205 is a concave hole having a circular cross section opened to the first rocker arm 162 side, and has the same diameter as the fourth plunger hole 204. An air vent hole 207 is provided at the bottom of the fifth plunger hole 205 to keep the internal atmospheric pressure at atmospheric pressure while allowing leaked oil to escape. A fifth plunger 185 having a bottomed cylindrical shape and an outer diameter equal to that of the fourth plunger 184 is fitted in the fifth plunger hole 205. The fifth plunger 185 slides smoothly in the fifth plunger hole 205. The total length of the fifth plunger 185 is shorter than the depth of the fifth plunger hole 205. The end of the fifth plunger 185 that contacts the fourth plunger 184 is formed in a spherical shape. A fifth plunger spring 189 is provided in the inner concave portion of the fifth plunger 185, and always biases the fifth plunger 185 toward the fourth plunger 184 side.
[0104]
A third plunger hole 203 is provided inside the third rocker arm 166. The third plunger hole 203 is a concave hole having a circular cross section that opens to the first rocker arm 162 side, and has the same diameter as the fourth plunger hole 204. A second hydraulic oil introduction passage 175 is led from the second hydraulic oil supply / discharge passage 174 to the bottom of the third plunger hole 203. A third plunger 183 having a cylindrical shape and an outer diameter equal to that of the fourth plunger 184 is fitted in the third plunger hole 203. The third plunger 183 smoothly slides in the third plunger hole 203 while sealing the hydraulic oil guided to the second hydraulic oil introduction path 175 on the outer peripheral surface thereof. The total length of the third plunger 183 is equal to the depth of the third plunger hole 203. The end of the third plunger 183 that contacts the fourth plunger 184 is formed in a spherical shape.
[0105]
FIG. 11A shows a state in which the first rocker arm 162 is separated from the second rocker arm 164 and the third rocker arm 166, and the hydraulic pressure is supplied to the first hydraulic oil supply / discharge passage 172 (hereinafter referred to as hydraulic pressure ON). The hydraulic pressure is not supplied to the second hydraulic oil supply / discharge passage 174 (hereinafter referred to as hydraulic pressure OFF). The first hydraulic oil introduction passage 173 guided from the first hydraulic oil supply / discharge passage 172 is turned on, and the first plunger 181 is pressed to the right (in the direction of the arrow in the figure). The pressing force is larger than the urging force of the second plunger spring 187, and the first plunger 181 moves to the right together with the second plunger 182. Since the entire length of the second plunger 182 is equal to the depth of the second plunger hole 202, the contact point between the first plunger 181 and the second plunger 182 is in the mating surface between the first rocker arm 162 and the second rocker arm 164.
[0106]
On the other hand, since the second hydraulic oil introduction path 175 guided from the second hydraulic oil supply / discharge passage 174 is turned off, the third plunger 183, the fourth plunger 184, and the fifth plunger 185 The plunger spring 189 is moved to the left (in the direction of the arrow in the figure) together by the urging force of the plunger spring 189. Since the entire length of the third plunger 183 is equal to the depth of the third plunger hole 203, the contact between the third plunger 183 and the fourth plunger 184 is in the mating surface between the first rocker arm 162 and the third rocker arm 166. Further, since the total length of the fourth plunger 184 is equal to the depth of the fourth plunger hole 204, the contact point between the fourth plunger 184 and the fifth plunger 185 is in the mating surface between the first rocker arm 162 and the second rocker arm 164.
[0107]
Thus, since the contact of each plunger exists in the mating surface of each rocker arm, the first rocker arm 162 is separated from the second rocker arm 164 and the third rocker arm 166. Therefore, the first rocker arm 162 is actuated by the first cam 152 in contact with the upper surface thereof, that is, the rocking stop around the rocker shaft 170 is stopped, and the burned gas introduction valve 31b is stopped in the closed state.
[0108]
FIG. 11B shows a state in which the first rocker arm 162 is interlocked only with the second rocker arm 164, and both the first hydraulic oil supply / discharge passage 172 and the second hydraulic oil supply / discharge passage 174 are turned off. ing. Since the first hydraulic oil introduction passage 173 guided from the first hydraulic oil supply / discharge passage 172 is turned off, the first plunger 181 and the second plunger 182 are driven by the urging force of the second plunger spring 187. It has moved to the left (arrow direction in the figure). Since the entire length of the first plunger 181 is shorter than the depth of the first plunger hole 201, a part of the second plunger 182 enters the first plunger hole 201.
[0109]
On the other hand, since the second hydraulic oil introduction passage 175 guided from the second hydraulic oil supply / discharge passage 174 is turned off, the third plunger 183, the fourth plunger 184, and the like, as in FIG. Is in the mating surface of the first rocker arm 162 and the third rocker arm 166, and the contact of the fourth plunger 184 and the fifth plunger 185 is in the mating surface of the first rocker arm 162 and the second rocker arm 164.
[0110]
Thus, when a part of the second plunger 182 enters the first plunger hole 201, the first rocker arm 162 is interlocked with the second rocker arm 164. Further, since the contact point between the third plunger 183 and the fourth plunger 184 is within the mating surface of the first rocker arm 162 and the third rocker arm 166, the first rocker arm 162 and the third rocker arm 166 are separated. ing. Accordingly, the first rocker arm 162 is operated by the second cam 154 that contacts the upper surface of the second rocker arm 164. That is, as the camshaft 151 rotates, the camshaft 151 swings downward by a predetermined amount at a predetermined crank angle to open and close the burned gas introduction valve 31b.
[0111]
FIG. 11C shows a state in which the first rocker arm 162 is interlocked with the second rocker arm 164 and the third rocker arm 166, and the first hydraulic oil supply / discharge passage 172 is turned off and the second hydraulic oil supply / discharge passage is shown. The passage 174 is hydraulically ON. Since the first hydraulic oil introduction passage 173 guided from the first hydraulic oil supply / discharge passage 172 is turned off, a part of the second plunger 182 is formed in the first plunger hole as in FIG. 201.
[0112]
On the other hand, since the second hydraulic oil introduction path 175 guided from the second hydraulic oil supply / discharge passage 174 is ON, the third plunger 183 is pressed to the right (in the direction of the arrow in the figure). The pressing force is larger than the urging force of the fifth plunger spring 189, and the third plunger 183 moves to the right together with the fourth plunger 184 and the fifth plunger 185. Since the total length of the fifth plunger 185 is shorter than the depth of the fifth plunger hole 205, a part of the fourth plunger 184 enters the fifth plunger hole 205, and a part of the third plunger 183 enters the fourth plunger hole 204. It has entered.
[0113]
Thus, a part of the second plunger 182 enters the first plunger hole 201 and a part of the third plunger 183 enters the fourth plunger hole 204, whereby the first rocker arm 162 and the third rocker arm 164 Interlocks with the rocker arm 166. Therefore, the first rocker arm 162 is actuated by the third cam 156 having the largest rotation radius that is in contact with the upper surface of the third rocker arm 166. That is, as the camshaft 151 rotates, the camshaft 151 swings downward by a predetermined amount at a predetermined crank angle to open and close the burned gas introduction valve 31b. The valve opening period is longer than that in the case of FIG.
[0114]
The cam switching mechanism 150 described above is for the burned gas introduction valve 31b and the burned gas discharge valve 32b, but the cam switching mechanism 150a (see FIG. 5) is also used for the succeeding cylinder intake valve 31a and the preceding cylinder exhaust valve 32a. 10 is shown in parentheses). However, in the cam switching mechanism 150a, the second cam 154 and the third cam 156 have the same shape. As shown in FIG. 10, the rocker arm set 160a that abuts against these cams includes a first rocker arm 162a, a second rocker arm 164a, and a third rocker arm 166a. In the cam switching mechanism 150a, the first rocker arm 162a is disconnected from the second rocker arm 164a and the third rocker arm 166a and the succeeding cylinder intake valve 31a and the preceding cylinder exhaust valve 32a are stopped in the closed state, and the first rocker arm 162a is stopped. In conjunction with the second rocker arm 164a and the third rocker arm 166a, the rotation of the second cam 154 and the third cam 156 switches to a state in which the subsequent cylinder intake valve 31a and the preceding cylinder exhaust valve 32a are opened and closed.
[0115]
FIG. 12 is an explanatory view showing the operation of the three plungers provided inside the rocker arm set 160a. 12A shows a state in which the first rocker arm 162 is separated from the second rocker arm 164 and the third rocker arm 166, and FIG. 12B shows a state in which the first rocker arm 162 is interlocked with the second rocker arm 164 and the third rocker arm 166. Indicates.
[0116]
The plunger structure in the rocker arm set 160a is such that the structure around the third plunger 183, the fourth plunger 184, and the fifth plunger 185 is provided among the plunger structures in the rocker arm set 160, and its detailed structure. Since the description of is duplicated with the description of the rocker arm set 160, a description thereof will be omitted. However, it differs from the rocker arm set 160 in that the first hydraulic oil introduction path 173a is guided from the first hydraulic oil supply / discharge passage 172 to the left end of the third plunger 183. Further, the second hydraulic oil supply / discharge passage 174 may be omitted if the structure can be omitted.
[0117]
FIG. 12A shows a state in which the first rocker arm 162 is separated from the second rocker arm 164 and the third rocker arm 166, and the first hydraulic fluid supply / discharge passage 172 is turned off. Since the first hydraulic oil introduction passage 173a guided from the first hydraulic oil supply / discharge passage 172 is turned off, the third plunger 183, the fourth plunger 184, and the fifth plunger 185 include the fifth plunger spring. They are moved together to the left (in the direction of the arrow in the figure) by the biasing force of 189. Therefore, the contact between the third plunger 183 and the fourth plunger 184 is in the mating surface between the first rocker arm 162a and the third rocker arm 166a, and the contact between the fourth plunger 184 and the fifth plunger 185 is between the first rocker arm 162a and the first rocker arm 162a. It is in the mating surface with the second rocker arm 164a.
[0118]
Thus, since the contact of each plunger exists in the mating surface of each rocker arm, the first rocker arm 162a is separated from the second rocker arm 164a and the third rocker arm 166a. For this reason, the first rocker arm 162a is actuated by the first cam 152 in contact with the upper surface thereof, that is, swinging around the rocker shaft 170 is stopped, and the burned gas introduction valve subsequent cylinder intake valve 31a and the preceding cylinder exhaust valve 32a are closed. Stop in the valve state.
[0119]
FIG. 12B shows a state in which the first rocker arm 162 is interlocked with the second rocker arm 164 and the third rocker arm 166, and the first hydraulic oil supply / discharge passage 172 is hydraulically ON. Since the first hydraulic oil introduction passage 173a guided from the first hydraulic oil supply / discharge passage 172 is in the hydraulic pressure ON state, the third plunger 183 is pressed to the right (in the direction of the arrow in the figure). The pressing force is larger than the urging force of the fifth plunger spring 189, and the third plunger 183 moves to the right together with the fourth plunger 184 and the fifth plunger 185. Therefore, a part of the fourth plunger 184 enters the fifth plunger hole 205, and a part of the third plunger 183 enters the fourth plunger hole 204.
[0120]
In this way, a part of the fourth plunger 184 enters the fifth plunger hole 205 and a part of the third plunger 183 enters the fourth plunger hole 204, whereby the first rocker arm 162 and the second rocker arm 164 Interlocks with 3 rocker arms 166. Accordingly, the first rocker arm 162 is operated by the second cam 154 and the third cam 156 (same shape) that are in contact with the upper surfaces of the second rocker arm 164a and the third rocker arm 166a. That is, as the camshaft 151 rotates, the camshaft 151 swings downward by a predetermined amount at a predetermined crank angle to open and close the burned gas introduction valve 31b.
[0121]
FIG. 13 shows the configuration of the drive and control system of this embodiment. The first control valve 176 and the second control valve 177 control the first hydraulic fluid and the second hydraulic fluid guided to the first hydraulic fluid supply / discharge passage 172 and the second hydraulic fluid supply / discharge passage 174 (hydraulic ON) / OFF) is a control valve. The cam switching mechanism 150 is provided in the burned gas introduction valve 31b and the burned gas discharge valve 32b, and the cam switching mechanism 150a is provided in the subsequent cylinder intake valve 31a and the preceding cylinder exhaust valve 32a.
[0122]
The ECU 40a is different from the ECU 40 of the first embodiment in that a cam switching control unit 190 is provided instead of the valve stop mechanism control unit 42 and the cam phase control unit 49.
[0123]
The cam switching control means 190 controls the first control valve 176 and the second control valve 177 in accordance with the special operation mode and the normal operation mode, or in accordance with the operation state, so that the cam switching mechanism 150 and the cam switching mechanism 150a are controlled. Is controlled as follows.
[0124]
Among special operation modes, low load and low speed areas:
・ First hydraulic pressure-OFF, second hydraulic pressure-OFF
The leading cylinder exhaust valve 32a and the trailing cylinder intake valve 31a are stopped (FIG. 12 (a)).
・ Burn gas discharge valve 32b, burnt gas introduction valve 31b
Operating state by the second cam 154 (low speed cam) (FIG. 11B)
Among special operation modes, high-load and high-speed areas:
・ First hydraulic pressure -OFF, second hydraulic pressure -ON
The leading cylinder exhaust valve 32a and the trailing cylinder intake valve 31a are stopped (FIG. 12 (a)).
・ Burn gas discharge valve 32b, burnt gas introduction valve 31b
Operation state by the third cam 156 (high-speed cam) (FIG. 11C)
Normal operation mode:
・ First hydraulic pressure -ON, second hydraulic pressure -OFF
・ The leading cylinder exhaust valve 32a and the trailing cylinder intake valve 31a
Operating state by the second cam 154 and the third cam 156 (FIG. 12B)
The burned gas discharge valve 32b and the burned gas introduction valve 31b are stopped (FIG. 11 (a)).
[0125]
Next, the operation of the apparatus of the present embodiment will be described, but description of portions overlapping with those of the first embodiment will be omitted. FIG. 14 is an explanatory view showing in detail the intake / exhaust stroke portion of the present embodiment. FIG. 14A shows the case of a relatively low-load low-rotation region (region A1 in FIG. 4) in the operation state in which the special operation mode is performed, and FIG. This is the case of the region (region A2 in FIG. 4). Other notation methods are based on FIG.
[0126]
In the upper part of FIG. 14 (a), the burned gas discharge valve opening period 230 (shown by hatching) in which the burned gas discharge valve 32b of the preceding cylinders 2A and 2D is opened, and the preceding cylinder intake valve 31 are opened. And a valve opening period 231 of the preceding cylinder intake valve. The lower stage shows a valve opening period 232 of the succeeding cylinder exhaust valve in which the succeeding cylinder exhaust valve 32 of the succeeding cylinders 2B and 2C is opened, and a valve opening period 233 of the burned gas introduction valve (indicated by hatching). Also, immediately after the burned gas discharge valve opening period 230 and the burned gas introduction valve opening period 233, the burned gas discharge valve opening period 250 and the burned gas introduction valve opening period 253, which will be described later, are provided. (All are in the special operation mode and relatively high rotation and high load) are shown by a two-dot chain line for reference. Such switching of the valve opening / closing timing is performed by the cam switching mechanisms 150 and 150a and the cam switching control means 190. The opening period 231 of the leading cylinder intake valve is set from about 10 ° CA before TDC to about 55 ° CA after BDC (total of about 245 ° CA). This is a general setting value for a conventional engine. On the other hand, the open period 233 of the burned gas introduction valve is set from about TDC to BDC (total of about 180 ° CA). Further, the valve opening period 232 of the subsequent cylinder exhaust valve is set from about 45 ° CA before BDC to about 10 ° CA after TDC (total about 235 ° CA).
[0127]
That is, the interval (about 0 ° CA) between the trailing cylinder intake stroke bottom dead center 247 and the burned gas introduction valve closing timing 248 is equal to the preceding cylinder intake stroke bottom dead center 242 and the preceding cylinder intake valve closing timing 243. Is set to be shorter than the interval (about 55 ° CA). The burned gas introduction valve opening period 233 is shorter than the preceding cylinder intake valve opening period 231 and the subsequent cylinder exhaust valve opening period 232. Thus, in the succeeding cylinders 2B and 2C, the open period 233 of the burnt gas introduction valve is set short, and the closing timing 248 of the burnt gas introduction valve is set early and close to BDC. There is no or short period during which the burned gas introduction valve 31b is open even when the piston 3 has risen past the BDC. For this reason, the effective compression ratio increases and approaches the geometric compression ratio. As the effective compression ratio increases, the in-cylinder temperatures of the succeeding cylinders 2B and 2C are likely to rise, and the compression ignitability is increased.
[0128]
The open period 230 of the burned gas discharge valve is set from about 45 ° CA before BDC to about 20 ° CA before TDC (total of about 205 ° CA). Therefore, the closing timing 241 of the burned gas discharge valve is set to be about 20 ° CA advance from the closing timing 248 of the burned gas introduction valve (corresponding to TDC in the preceding cylinders 2A and 2D). This is earlier than a general set value (about 50 ° CA after TDC) of a conventional engine. By closing the burned gas discharge valve 32b early, the internal EGR of the preceding cylinders 2A, 2D increases, and the burned gas temperature introduced into the succeeding cylinders 2B, 2C rises.
[0129]
Since the valve opening period 232 of the subsequent cylinder exhaust valve is set from about 45 ° CA before BDC to about 10 ° CA after TDC (total of about 235 ° CA), the subsequent cylinder exhaust valve 32 is dead due to the subsequent cylinder exhaust stroke. At point 244, the valve is open. On the other hand, the opening timing 246 of the burned gas introduction valve is set so as to be approximately the top dead center 245 of the subsequent cylinder intake stroke. That is, the opening timing 246 of the burned gas introduction valve is set to the succeeding cylinder intake stroke top dead center 245 while the succeeding cylinder exhaust valve 32 is opened to the succeeding cylinder exhaust stroke top dead center 244. . For this reason, since the valve overlap in the vicinity of the top dead center 244 of the subsequent cylinder exhaust stroke is shortened, the burned gas introduced into the subsequent cylinders 2B and 2C is directly discharged to the exhaust passage 20 through the subsequent cylinder exhaust valve 32. In other words, so-called blow-through is prevented, and the effective compression rate of the succeeding cylinders 2B and 2C can be increased.
[0130]
As described above, the effective compression ratio of the succeeding cylinders 2B and 2C is increased, and the in-cylinder temperature of the succeeding cylinders 2B and 2C is increased by increasing the burnt gas temperature introduced into the succeeding cylinders 2B and 2C. To improve compression ignitability. By doing in this way, the operation area | region which can perform the combustion by compression ignition is expanded to the area | region of a low load.
[0131]
FIG. 14B is an explanatory diagram in the case of a relatively high load and high speed region (region A2 in FIG. 4). In the upper stage, the burned gas discharge valve opening period 250 (shown by a mesh) in which the burned gas discharge valves 32b of the preceding cylinders 2A and 2D are opened, and the opening period 231 of the preceding cylinder intake valve (see FIG. ) And common). The lower stage includes a valve opening period 232 for the subsequent cylinder exhaust valve (common to FIG. 14A) and a valve opening period 253 for the burned gas introduction valve for opening the burnt gas introduction valve 31b (shown by a mesh). Show. Further, immediately after the burned gas discharge valve opening period 250 and the burned gas introduction valve opening period 253, the burned gas discharge valve opening period 230 and the burned gas introduction valve of FIG. A two-dot chain line indicates the valve opening period 233 as a reference.
[0132]
The open period 253 of the burned gas introduction valve is set from about 10 ° CA before TDC to about 55 ° CA after BDC. That is, the burned gas introduction valve is opened 10 ° CA early and closed 55 ° later with respect to the open period 233 of the burned gas introduction valve in the low load low rotation region. Accordingly, the effective compression ratio of the subsequent cylinders 2B and 2C is smaller than that in the case of FIG. 14A, and the in-cylinder temperature of the subsequent cylinders 2B and 2C is decreased.
[0133]
The open period 250 of the burned gas discharge valve is set from about 45 ° CA before BDC to about 10 ° CA after TDC. That is, it is about 30 ° CA longer than the open period 230 of the burned gas discharge valve in FIG. For this reason, the internal EGR of the preceding cylinders 2A and 2D is smaller than that in the case of FIG. 14A, and the burnt gas temperature introduced into the succeeding cylinders 2B and 2C is lowered.
[0134]
As described above, the effective compression ratio of the succeeding cylinders 2B and 2C is reduced, and the in-cylinder temperature of the succeeding cylinders 2B and 2C is lowered by lowering the burnt gas temperature introduced into the succeeding cylinders 2B and 2C. The occurrence of abnormal combustion such as knocking is prevented. By doing in this way, the operation area | region which can perform the combustion by compression ignition is expanded to the area | region of a higher load.
[0135]
Note that the patterns shown in FIGS. 14A and 14B do not limit each valve opening period and valve opening timing, and may be appropriately changed within the scope of the claims. For example, the closing timing 248 of the burned gas introduction valve in FIG. 14A may be set to be slightly delayed from the bottom dead center 247 of the subsequent cylinder intake stroke. Other specific set values may be set to suitable values according to the required characteristics of the engine.
[0136]
In the present embodiment, the structure is such that there are a mixture of three types of cam switching, two types of cam switching, and no cam switching according to the location of the intake / exhaust valve and the respective functions. However, the combination is not limited to this, and for example, three types of cams may be switched by making three types of cams and the rocker arm set 160 correspond to all valves. The valve opening period 231 of the preceding cylinder intake valve and the valve opening period 232 of the subsequent cylinder exhaust valve, which are unchanged in the second embodiment, may be changed by cam switching.
[0137]
The present invention is not limited to the first and second embodiments, and various modifications are possible within the scope of the invention described in the claims. For example, the means for changing the opening / closing timing of each valve as required is not limited to a mechanism for changing the phase of the camshaft as in the first embodiment or a mechanism for switching the cam as in the second embodiment. There is no limitation as long as the setting shown in the claims can be obtained.
[0138]
The present invention is not limited to a four-cylinder engine, and may be applied to a six-cylinder or more engine in which a pair of cylinders composed of a preceding cylinder and a succeeding cylinder is suitably combined in three or more pairs.
[0139]
【The invention's effect】
As described above, in the control device of the present invention, the burned gas discharged from the preceding cylinder, which is the cylinder on the exhaust stroke side, is directly taken in between the pair of cylinders where the exhaust stroke and the intake stroke overlap at least in the low load and low rotation range. The gas flow path is configured in a two-cylinder connection state in which the exhaust gas discharged from the subsequent cylinder is introduced into the subsequent cylinder, which is the stroke side cylinder, via the inter-cylinder gas path, and is guided to the exhaust path. When the cylinder is connected, combustion is performed by forced ignition with a lean air-fuel ratio that is a predetermined amount larger than the stoichiometric air-fuel ratio in the preceding cylinder, and in the succeeding cylinder, the lean air-fuel ratio introduced from the preceding cylinder is reduced. Combustion control means for controlling the combustion of each cylinder so as to supply fuel to the burned gas and perform combustion by compression ignition, and at least the two-cylinder connected state In addition, a preceding cylinder intake valve that introduces fresh air into the preceding cylinder and a burned gas introduction valve that introduces burned gas into the succeeding cylinder from the inter-cylinder gas passage are provided so that the two cylinders are connected. Among the regions, at least in a predetermined region on the low load side, the interval between the intake stroke bottom dead center of the subsequent cylinder and the closing timing of the burned gas introduction valve is set to the intake stroke bottom dead center of the preceding cylinder and the preceding stroke. It is characterized by being set to be shorter than the interval between the cylinder intake valve and the valve closing timing, so it is possible to ensure exhaust purification performance with only a three-way catalyst, while reducing lean combustion and pumping loss, etc. By obtaining a fuel efficiency improvement effect and expanding a driving range where the contribution is high, a greater fuel efficiency improvement effect can be obtained.
[Brief description of the drawings]
FIG. 1 is a schematic plan view of an entire engine including an apparatus according to a first embodiment of the present invention.
FIG. 2 is a schematic cross-sectional view of an engine body and the like.
FIG. 3 is a block diagram of a control system of the first embodiment.
FIG. 4 is an explanatory diagram showing an operation region.
FIG. 5 is a diagram showing an exhaust stroke, an intake stroke, a fuel injection timing, an ignition timing, and the like of each cylinder.
FIG. 6 is an explanatory diagram showing substantial fresh air and gas flow paths during low load and low rotation.
FIG. 7 is an explanatory diagram showing a substantial fresh air and gas flow path when in an operation region on a high load, high and low rotation side.
FIG. 8 is an explanatory diagram showing the opening / closing timing of the intake / exhaust valve in the special operation mode of the first embodiment. FIG. 8A shows a relatively low load and low rotation, and FIG. The case of high load rotation is shown.
FIG. 9 is an explanatory diagram showing opening and closing timings of intake and exhaust valves in a normal operation mode.
FIG. 10 is a partial perspective view showing a cam switching mechanism of a second embodiment.
FIG. 11 is a diagram illustrating plunger operation for three types of cam switching mechanisms in the second embodiment.
FIG. 12 is a plunger operation explanatory view for two types of cam switching mechanisms in the second embodiment.
FIG. 13 is a block diagram of a control system of a second embodiment.
FIG. 14 is an explanatory diagram showing the opening / closing timing of the intake / exhaust valve in the special operation mode of the second embodiment, where (a) shows a relatively low load and low rotation, and (b) shows a relatively high value. The case of high load rotation is shown.
[Explanation of symbols]
1 Engine body
2A, 2D 1st and 4th cylinders (preceding cylinder)
2B, 2C 2nd and 3rd cylinders (following cylinders)
9 Fuel injection valve
11 Intake port
11, 11a, 11b Intake port
12, 12a, 12b Exhaust port
15 Intake passage
20 Exhaust passage
22 Gas passage between cylinders
31 Leading cylinder intake valve
31a Subsequent cylinder intake valve
31b Burned gas introduction valve
32 Successive cylinder exhaust valve
32a Lead cylinder exhaust valve
32b Burned gas discharge valve
33a, 34a Cam phase variable mechanism
40, 40a ECU
42 Valve stop mechanism control means
44 Combustion control means
49 Cam phase control means
150 Cam switching mechanism
160 Rocker arm set
190 Cam switching control means

Claims (8)

In a multi-cylinder spark ignition engine in which each cylinder performs a cycle consisting of intake, compression, expansion, and exhaust strokes with a predetermined phase difference,
At least in the low load and low rotation range, between the pair of cylinders where the exhaust stroke and the intake stroke overlap, the burned gas discharged from the preceding cylinder which is the cylinder on the exhaust stroke side is directly between the cylinders in the subsequent cylinder which is the cylinder on the intake stroke side The gas flow path is configured in a two-cylinder connection state in which exhaust gas introduced through the gas passage and exhausted from the subsequent cylinder is guided to the exhaust passage,
When the two cylinders are connected, combustion is performed by forced ignition in a state where the preceding cylinder has a lean air-fuel ratio that is a predetermined amount larger than the theoretical air-fuel ratio, and in the subsequent cylinder, the lean air introduced from the preceding cylinder is performed. Combustion control means for controlling the combustion of each cylinder so as to supply a quantity of fuel substantially equal to the stoichiometric air-fuel ratio to the burned gas of the fuel ratio and to cause combustion by compression ignition ;
A preceding cylinder intake valve that introduces fresh air into the preceding cylinder, a preceding cylinder exhaust valve that discharges exhaust gas from the preceding cylinder to the exhaust passage, and an existing cylinder that introduces burned gas into the subsequent cylinder from the inter-cylinder gas passage. A fuel gas introduction valve and a subsequent cylinder intake valve for introducing fresh air into the subsequent cylinder,
At least when the two cylinders are connected, new air is introduced into the preceding cylinder by opening the preceding cylinder intake valve , and the burnt gas introduction valve is opened to open the space between the cylinders. While burned gas is introduced from the gas passage to the subsequent cylinder, the preceding cylinder exhaust valve and the subsequent cylinder intake valve are closed,
Among the operating regions in which the two cylinders are connected, at least in a predetermined region on the low load side, the interval between the intake stroke bottom dead center of the succeeding cylinder and the closing timing of the burnt gas introduction valve is the same as that of the preceding cylinder. A control device for a spark ignition type engine, characterized in that it is set to be shorter than an interval between an intake stroke bottom dead center and a closing timing of the preceding cylinder intake valve. A succeeding cylinder exhaust valve that exhausts exhaust gas from the succeeding cylinder when at least the two cylinders are connected;
The burned gas introduction valve is opened while the succeeding cylinder exhaust valve is opened to the top dead center of the exhaust stroke of the succeeding cylinder in at least a predetermined region on the low load side in the operation region in which the two cylinders are connected. 2. The spark ignition engine control apparatus according to claim 1, wherein the valve opening timing is set to be the top dead center of the intake stroke of the subsequent cylinder.   Of the operating range in which the two cylinders are connected, in the predetermined region on the high load side, the closing timing of the burned gas introduction valve is set to be behind the timing when the closed gas in the predetermined region on the low load side. 3. The spark ignition engine control device according to claim 1, wherein the control device is a spark ignition engine.   Of the operating region in which the two cylinders are connected, in a predetermined region on the high load high rotation side, the timing for closing the burned gas introduction valve is higher than that in the predetermined region on the low load low rotation side. 3. The spark ignition engine control device according to claim 1, wherein the control device is set to a delay side. A burned gas discharge valve for discharging the burned gas of the preceding cylinder into the inter-cylinder gas passage when at least the two cylinders are connected;
In the operation region in which the two-cylinder connection state is established, the closing timing of the burned gas discharge valve is set to a more advanced side than the closing timing of the burned gas introduction valve,
While the opening period of the burned gas discharge valve and the opening period of the burned gas introduction valve are each individually set to a predetermined value,
The opening timing of the burned gas discharge valve and the opening timing of the burned gas introduction valve are set so as to fluctuate back and forth while keeping the timing difference constant according to at least the engine load. The control device for a spark ignition type engine according to any one of claims 1 to 4. In a multi-cylinder spark ignition engine in which each cylinder performs a cycle consisting of intake, compression, expansion, and exhaust strokes with a predetermined phase difference,
At least in the low load and low rotation range, between the pair of cylinders where the exhaust stroke and the intake stroke overlap, the burned gas discharged from the preceding cylinder which is the cylinder on the exhaust stroke side is directly between the cylinders in the subsequent cylinder which is the cylinder on the intake stroke side The gas flow path is configured in a two-cylinder connection state in which exhaust gas introduced through the gas passage and exhausted from the subsequent cylinder is guided to the exhaust passage,
When the two cylinders are connected, combustion is performed by forced ignition in a state where the preceding cylinder has a lean air-fuel ratio that is a predetermined amount larger than the theoretical air-fuel ratio, and in the subsequent cylinder, the lean air introduced from the preceding cylinder is performed. Combustion control means for controlling the combustion of each cylinder so as to supply a quantity of fuel substantially equal to the stoichiometric air-fuel ratio to the burned gas of the fuel ratio and to cause combustion by compression ignition ;
A preceding cylinder intake valve that introduces fresh air into the preceding cylinder, a preceding cylinder exhaust valve that discharges exhaust gas from the preceding cylinder to the exhaust passage, and an existing cylinder that introduces burned gas into the subsequent cylinder from the inter-cylinder gas passage. A fuel gas introduction valve and a subsequent cylinder intake valve for introducing fresh air into the subsequent cylinder,
At least when the two cylinders are connected, new air is introduced into the preceding cylinder by opening the preceding cylinder intake valve , and the burnt gas introduction valve is opened to open the space between the cylinders. While burned gas is introduced from the gas passage to the subsequent cylinder, the preceding cylinder exhaust valve and the subsequent cylinder intake valve are closed,
Of the operation region where the above-mentioned two-cylinder connected condition in a predetermined area of at least the low-load side, shorter than the open period of the upper Symbol preceding cylinders intake valve open period of said burnt gas introduction valve, and the following cylinder The interval between the intake stroke bottom dead center and the burned gas introduction valve closing timing is shorter than the interval between the intake stroke bottom dead center of the preceding cylinder and the closing timing of the preceding cylinder intake valve. A control device for a spark ignition engine characterized by being set. A succeeding cylinder exhaust valve that exhausts exhaust gas from the succeeding cylinder when at least the two cylinders are connected;
Of the operating region in which the two cylinders are connected, at least in a predetermined region on the low load side, the open period of the burned gas introduction valve is set to be shorter than the open period of the subsequent cylinder exhaust valve. The control device for a spark ignition engine according to claim 6.   A burned gas discharge valve for discharging burned gas of the preceding cylinder to the inter-cylinder gas passage when at least the two cylinders are connected;
  In at least a predetermined region on the low load side of the operation region in which the two cylinders are connected, the opening timing of the burned gas introduction valve is set to be delayed from the opening timing of the burned gas discharge valve. The spark ignition engine control device according to claim 6, wherein the control device is a spark ignition engine.

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