JP2004027962A - Controller for spark ignition-type four-cycle engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To remarkably improve the fuel consumption by applying both of lean combustion and compressive self-ignition as much as possible, and to improve the self-ignition particularly in the compressive self-ignition. <P>SOLUTION: The combustion is performed with a lean air-fuel ratio in advanced cylinders 2A, 2D, and the combustion is performed substantially with a theoretical air-fuel ratio in trailing cylinders 2B, 2C, in a state of keeping a two-cylinder connection state for guiding the burned gas discharged from the advanced cylinders 2A, 2D to the trailing cylinders 2B, 2C as it is, between the pair of cylinders where an exhaust stroke and an intake stroke are overlapped, in a partial load area of an engine. Whether the ignition for combustion particularly in the trailing cylinders 2B, 2C is the compressive self-ignition or the forcible ignition, is selected, and the fuel injection amount to each cylinder is controlled to adjust the air-fuel ratio in the advanced cylinders 2A, 21D to be a small value in comparison with the case of the forcible ignition, when the compressive self-ignition is selected. <P>COPYRIGHT: (C)2004,JPO

Description

上記吸入空気量制御手段４５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、特殊運転モードとされる運転領域Ａでは、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リーン空燃比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。
【００４８】
上記燃料制御手段４６は、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し、また上記点火制御手段４６は、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、特に上記モード設定手段４３により設定されるモードに応じ、燃焼状態の制御（燃料の制御及び点火の制御）が変更される。
【００４９】
すなわち、特殊運転モードが設定された場合、先行気筒２Ａ，２Ｄに対しては、空燃比が理論空燃比よりも大きいリーン空燃比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時期を設定し、かつ、圧縮上死点付近で強制点火を行わせるように点火時期を設定する。
【００５０】
一方、後続気筒２Ｂ，２Ｃに対しては、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対して燃料を供給し、後続気筒２Ｂ，２Ｃでの燃焼の際に実質的に理論空燃比となるように燃料噴射量を制御する。そして、この特殊運転モードの中で圧縮自己着火モードが選択されたときは、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時期を設定するとともに、圧縮自己着火を行わせるべく、強制点火を停止させる。また、強制点火モードが選択されたときは、圧縮行程で燃料を噴射するように噴射時期を設定するとともに、圧縮上死点付近の所定時期に強制点火を行わせるように点火時期を設定する。
【００５１】
さらにこの特殊運転モードにおいて、一対の気筒の両方に対する燃料噴射量の和が先行気筒２Ａ，２Ｄに導入される空気の量に対して理論空燃比となる量に調整されつつ、圧縮自己着火モードが選択された場合と強制点火モードが選択された場合とで、先行気筒２Ａ，２Ｄに対する燃料噴射量と後続気筒２Ｂ，２Ｃに対する燃料噴射量との割合が変更されることにより、先行気筒２Ａ，２Ｄの空燃比が変更される。そして、圧縮自己着火モードが選択された場合、先行気筒２Ａ，２Ｄの空燃比が理論空燃比よりも大（空気過剰率λが１＜λ）となり、かつ、強制点火モードが選択された場合と比べて小さくなるように燃料噴射量が調整される。
【００５２】
具体的には、強制点火モードでは先行気筒２Ａ，２Ｄに対する燃料噴射量が後続気筒２Ｂ，２Ｃに対する燃料噴射量と同程度もしくはそれ以下とされることにより先行気筒２Ａ，２Ｄの空燃比が理論空燃比の２倍程度（Ａ／Ｆ≒３０）もしくはそれより大とされるのに対し、圧縮自己着火モードでは先行気筒２Ａ，２Ｄに対する燃料噴射量が後続気筒２Ｂ，２Ｃに対する燃料噴射量よりも多くされて、先行気筒２Ａ，２Ｄの空燃比が例えばＡ／Ｆ≒２５とされる。
【００５３】
なお、圧縮自己着火モードにおける先行気筒２Ａ，２Ｄの空燃比は運転状態に応じて変更してもよい。このようにする場合には、後に詳述するように自己着火性向上及びノッキング抑制のため、特殊運転モード領域Ａ内の低速低負荷側や高速高負荷側で、同領域Ａの中速中負荷域と比べ、先行気筒２Ａ，２Ｄの空燃比を小さくする（先行気筒２Ａ，２Ｄに対する燃料供給量の割合を多くする）ことが好ましい。
【００５４】
また、通常運転モードが選択された場合には、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、例えばこの運転領域Ｂ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時期を設定し、かつ、各気筒２Ａ〜２Ｄとも強制点火を行わせるようにする。
【００５５】
なお、ＥＣＵ４０は、好ましくはさらに燃料のオクタン価を判別する手段を備え、後述のフローチャートに示すように、特殊運転モードにおける圧縮自己着火モードの場合に、この領域Ａのうちの低速、低負荷側の領域Ａ１にあるか否かということと、燃料のオクタン価とに応じ、先行気筒の空燃比の補正（燃料噴射量の補正）を行うようになっている。
【００５６】
上記ＥＣＵ４０に含まれる各手段による処理をフローチャートで示すと、図５のようになる。
【００５７】
このフローチャートの処理がスタートすると、先ずステップＳ１で各種信号が入力され、次にステップＳ２で、運転状態が特殊運転モード領域Ａにあるか否かが判定される。
【００５８】
特殊運転モード領域Ａにある場合には、ステップＳ３で吸・排気弁が特殊運転モードに制御され、つまり前述のような弁停止機構機構３５の制御により吸・排気流通状態が２気筒接続状態となるように制御されるとともに、ステップＳ４で特殊運転モードでの目標吸入空気量が得られるようにスロットル弁１７が制御される。さらに、ステップＳ５で、エンジンの温度状態が所定温度以下の低温であるか否かが判定される。
【００５９】
ステップＳ５の判定がＹＥＳの場合は、ステップＳ６で強制点火モードが選択されて、先行気筒２Ａ，２Ｄ及び後続気筒２Ｂ，２Ｃがともに強制点火とされるように各気筒２Ａ〜２Ｄの点火時期が設定されるとともに、ステップＳ７で燃料噴射量が演算される。この場合、先行気筒２Ａ，２Ｄでの空気過剰率λを１よりも大きい所定値λａ（例えばλａ≒２）とするとともに、後続気筒２Ｂ，２Ｃでの空気過剰率λをλ≒１とするように、各気筒２Ａ〜２Ｄに対する燃料噴射量が演算される。さらに、ステップＳ８で、先行気筒２Ａ，２Ｄでは成層燃焼のため圧縮行程噴射とし、後続気筒２Ｂ，２Ｃでも着火性確保のため圧縮行程噴射とするように、各気筒２Ａ〜２Ｄの燃料噴射時期が演算される。そして、ステップＳ９で、演算ないし設定された各気筒２Ａ〜２Ｄに対する燃料噴射量、燃料噴射時期及び点火時期に従って燃料噴射及び点火の制御が行われる。
【００６０】
上記ステップＳ５の判定がＮＯの場合、つまりエンジンが所定温度より高温の状態にある場合は、ステップＳ１０で圧縮自己着火モードが選択されて、先行気筒２Ａ，２Ｄのみ強制点火とされるように点火時期が設定されるとともに、ステップＳ１１で燃料噴射量が演算される。この場合、先行気筒２Ａ，２Ｄでの空気過剰率λを１＜λｂ＜λａとなる値とするとするとともに、後続気筒２Ｂ，２Ｃでの空気過剰率λをλ≒１とするように、各気筒２Ａ〜２Ｄに対する燃料噴射量が演算される。
【００６１】
続いてステップＳ１２で、特殊運転モード領域Ａのうちの領域Ａ１にあるか否かが判定され、その判定がＹＥＳの場合はステップＳ１３で燃料のオクタン価が高いか否かが判定され、ステップＳ１２の判定がＮＯの場合はステップＳ１４で燃料のオクタン価が低いか否かが判定される。
【００６２】
これらの判定に基づき、運転状態が上記領域Ａ１内にあって、燃料のオクタン価が高い場合には、ステップＳ１５で、オクタン価に応じて先行気筒２Ａ，２Ｄの空燃比がリッチ側に補正され、つまり、オクタン価が高いほど先行気筒２Ａ，２Ｄの空燃比が小さくなる（燃料噴射量が多くなる）ように補正される。また、運転状態が特殊運転モード領域Ａ内の上記領域Ａ１外にあって、燃料のオクタン価が低い場合にもステップＳ１５に移ってオクタン価に応じた補正が行われ、この場合にオクタン価が低いほど先行気筒２Ａ，２Ｄの空燃比が小さくなる（燃料噴射量が多くなる）ように補正される。
【００６３】
ステップＳ１５の処理が行われた場合、あるいはステップＳ１３またはステップＳ１４の判定がＮＯの場合は、ステップＳ８に移って各気筒２Ａ〜２Ｄの燃料噴射時期の演算が行われるが、この場合、先行気筒２Ａ，２Ｄでは成層燃焼のため圧縮行程噴射とされる一方、後続気筒２Ｂ，２Ｃでは吸気行程噴射とされる。そして、ステップＳ９で燃料噴射及び点火の制御が行われる。
【００６４】
また、上記ステップＳ２の判定がＮＯの場合、つまり運転状態が通常運転モード領域Ｂにある場合には、ステップＳ１６で吸・排気弁が通常運転モードに制御され、つまり前述のような弁停止機構機構３５の制御により吸・排気流通状態が各気筒独立状態となるように制御されるとともに、ステップＳ１７で通常運転モードでの目標吸入空気量が得られるようにスロットル弁１７が制御される。さらに、ステップＳ１８で、各気筒２Ａ〜２Ｄとも強制点火を行わせるように点火時期が設定されるとともに、ステップＳ１９で、各気筒２Ａ〜２Ｄでの空気過剰率λが１以下となるように各気筒２Ａ〜２Ｄに対する燃料噴射量が演算され、さらにステップＳ２０で、いずれの気筒も均一燃焼のため吸気行程噴射とするように燃料噴射時期が演算される。そして、ステップＳ９で燃料噴射及び点火の制御が行われる。
【００６５】
以上のような当実施形態の装置の作用を、図６〜図９を参照しつつ説明する。
【００６６】
低速低負荷側の特殊運転モード領域Ａでは、特殊運転モードとされ、前述のように第１排気弁３２ａ及び第１吸気弁３１ａが停止状態、第２排気弁３２ｂ及び第２吸気弁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８に示すようになり、先行気筒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。
【００６７】
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８中の矢印ａ）、先行気筒２Ａ，２Ｄでは空燃比が理論空燃比よりも大きいリーン空燃比となるように燃料噴射量が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、成層燃焼が行われる（図６、図７参照）。
【００６８】
それから、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６、図７中の白抜き矢印及び図８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、理論空燃比となるように燃料噴射量が制御されつつ燃焼が行われる。
【００６９】
この場合、エンジンが所定温度より高温の状態にある時には、圧縮自己着火モードが選択され、図６に示すように、後続気筒２Ｂ，２Ｃでは吸気行程で燃料が噴射された後、圧縮行程の上死点付近で燃焼室内が充分に高温、高圧の状態となって、圧縮自己着火が良好に行われる。
【００７０】
すなわち、先行気筒２Ａ，２Ｄから排出された高温の既燃ガスが短い気筒間ガス通路２２を通って後続気筒２Ｂ，２Ｃに直ちに導入されるため、後続気筒２Ｂ，２Ｃでは吸気行程で燃焼室内の温度が高くなり、この状態からさらに圧縮行程で圧力、温度が上昇することにより、圧縮行程終期の上死点付近では混合気が自己着火し得る程度まで燃焼室内の温度が上昇する。しかも、上記既燃ガスは先行気筒２Ａ，２Ｄから排出されて後続気筒２Ｂ，２Ｃに導入されるまでの間に充分にミキシングされて均一に分布し、さらに吸気行程で噴射された燃料も圧縮行程終期までの間に燃焼室全体に均一に分散するため、理想的な同時圧縮自己着火条件を満たすような均一な混合気分布状態が得られる。そして、同時圧縮自己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。
【００７１】
このように、先行気筒２Ａ，２Ｄでは、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、成層燃焼を行わない通常のエンジンと比べて吸気負圧が小さくなることでポンピングロスが低減され、一方、後続気筒２Ｂ，２Ｃでは、空燃比が略理論空燃比とされつつ、均一な混合気分布状態で圧縮自己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、先行気筒２Ａ，２Ｄから押出されたガスが送り込まれるため先行気筒２Ａ，２Ｄよりもさらにポンピングロスが低減される。これらの作用により、燃費が大幅に改善される。
【００７２】
しかも、後続気筒２Ｂ，２Ｃから排気通路２０に排出されるガスは理論空燃比であるため、従来のリーンバーンエンジンのようにリーンＮＯｘ触媒を設ける必要がなく、三元触媒２４だけで充分に排気浄化性能が確保される。
【００７３】
そして、リーンＮＯｘ触媒を設ける必要がないことから、リーンＮＯｘ触媒のＮＯｘ吸蔵量増大時におけるＮＯｘの放出、還元のための一時的な空燃比のリッチ化を行う必要がなく、燃費改善の目減りが避けられる。さらに、リーンＮＯｘ触媒の硫黄被毒の問題が生じることもない。
【００７４】
また、先行気筒２Ａ，２Ｄでは理論空燃比の略２倍もしくはそれに近い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る。一方、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とともに、同時圧縮自己着火による急速燃焼が行われると可及的に酸素と窒素との反応が避けられることから、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。
【００７５】
また、後続気筒２Ｂ，２Ｃでの圧縮自己着火が先行気筒２Ａ，２Ｄから排出される既燃ガスの熱を利用して達成されるため、格別の加熱手段を用いたりエンジンの圧縮比を極端に高くしたりする必要がなく、容易に圧縮自己着火を達成することができる。
【００７６】
とくに、後続気筒２Ｂ，２Ｃでの燃焼が圧縮自己着火により行われる場合に、先行気筒２Ａ，２Ｄの空燃比が、理論空燃比より大きく、かつ、強制点火とされる場合と比べて小さくされることにより、圧縮自己着火が行われ易くなり、しかも、ノッキング抑制にも有利となる。
【００７７】
すなわち、このように先行気筒２Ａ，２Ｄの空燃比が設定されることにより、強制点火とされる場合と比べ、先行気筒２Ａ，２Ｄに対する燃料噴射量が多くなることにより、先行気筒２Ａ，２Ｄでの発熱量が増加し、先行気筒２Ａ，２Ｄから後続気筒２Ｂ，２Ｃへ導かれるガスの温度が上昇するため、比較的低速低負荷側でも圧縮自己着火が行われ易くなる。また、このように後続気筒２Ｂ，２Ｃに導入されるガスの温度は上昇するものの、後続気筒２Ｂ，２Ｃに導入されるガス中のＥＧＲに相当する既燃ガス成分が増大するとともに、後続気筒２Ｂ，２Ｃに対する燃料噴射量が少なくなることによって後続気筒２Ｂ，２Ｃでの燃焼により発生するエネルギーが小さくなるため、比較的高速高負荷側でもノッキングが抑制される。
【００７８】
とくに、特殊運転モード領域Ａ内の低速低負荷側や高速高負荷側で、同領域Ａの中速中負荷域と比べ、先行気筒２Ａ，２Ｄの空燃比を小さくする（先行気筒２Ａ，２Ｄに対する燃料供給量の割合を多くする）ように運転状態に応じて燃料噴射量を制御すれば、低速低負荷側での自己着火性向上及び高速高負荷側でのノッキング抑制の効果が高められる。
【００７９】
また、エンジンの運転状態及び燃料のオクタン価に応じ、特殊運転モード領域Ａのうちの高速、高負荷側の領域にあって、かつオクタン価が低い場合（耐ノック性の面で不利な情況にある場合）には、先行気筒２Ａ，２Ｄの空燃比を小さくするように燃料噴射量を制御すれば、効果的にノッキングが抑制される。さらに、特殊運転モード領域Ａのうちの低速、低負荷側の領域Ａ１にあって、かつオクタン価が高い場合（自己着火にとって不利な情況にある場合）にも、先行気筒２Ａ，２Ｄの空燃比を小さくするように燃料噴射量を制御すれば、充分に圧縮自己着火を行わせることができる。
【００８０】
また、エンジンが所定温度以下の低温状態にある場合には、先行気筒２Ａ，２Ｄの空燃比を小さくしても後続気筒２Ｂ，２Ｃに導入されるガスの温度及び後続気筒２Ｂ，２Ｃでの圧縮行程における温度上昇が充分に高くならずに自己着火が困難になる場合があり、また、先行気筒２Ａ，２Ｄの空燃比を小さくしすぎると後続気筒２Ｂ，２Ｃの燃焼安定性や燃費にとって不利になる。
【００８１】
そこで、このような場合には、強制点火モードとされ、図７に示すように、先行気筒２Ａ，２Ｄと同様に後続気筒２Ｂ，２Ｃでも圧縮上死点付近で点火プラグによる点火が行われ、強制点火で確実に着火、燃焼が行われる。
【００８２】
この強制点火によると、圧縮自己着火が良好に行われている場合と比べると、後続気筒２Ｂ，２Ｃでの燃焼効率が多少低下するが、それでも上記特殊運転モードとされている限り、先行気筒２Ａ，２Ｄでの成層リーンバーン、後続気筒２Ｂ，２Ｃでのポンピングロス低減等による燃費改善効果は得られるとともに、後続気筒２Ｂ，２Ｃから排出されるガスが理論空燃比であるため三元触媒だけで充分に排気浄化性能が確保される。とくに、強制点火モードとされる場合は先行気筒２Ａ，２Ｄの空燃比が理論空燃比の略２倍若しくはそれより大とされることにより、先行気筒２Ａ，２Ｄでのリーンバーンによる燃費改善効果及び後続気筒２Ｂ，２Ｃでの燃焼安定性が確保される。
【００８３】
一方、高速側ないし高負荷側の運転領域Ｂでは通常運転モードとされ、前述のように第１排気弁３２ａ及び第１吸気弁３１ａが作動状態、第２排気弁３２ｂ及び第２吸気弁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９に示すようになり、各気筒２Ａ〜２Ｄの吸気ポート３１，３１ａ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３１，３１ａに新気が導入されるとともに各気筒２Ａ〜２Ｄの排気ポート３１，３１ａから排気通路２０に既燃ガスが排出される。そしてこの場合は、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。
【００８４】
なお、本発明の装置の具体的構成は上記実施形態に限定されず、種々変更可能である。他の実施形態を以下に説明する。
【００８５】
▲１▼特殊運転モードのうちの圧縮自己着火モードとされるときに、上記実施形態では単に後続気筒２Ｂ，２Ｃに対する点火を停止しているが、後続気筒２Ｂ，２Ｃに対し、強制点火とする場合の点火時期よりも所定量リタードした時期にバックアップのための点火Ｓ´（図６中に破線で示す）を行わせるようにしてもよい。このバックアップのための点火Ｓ´は、圧縮上死点より後であって、圧縮上死点の近傍に設定すればよい。
【００８６】
このようにすれば、圧縮自己着火モードにおいて、何らかの原因で圧縮自己着火が良好に行われないような事態が生じた場合でも、上記バックアップのための点火Ｓ´により着火燃焼が行われ、トルク変動が避けられるとともに、エミッションの悪化が防止される。
【００８７】
▲２▼基本実施形態では、特殊運転モード領域Ａの全体で、エンジンの温度状態に応じて高温時に圧縮自己着火、低温時に強制点火とするように制御しているが、エンジンの低速域（例えば領域Ａ１）でのみエンジンの温度状態に応じて圧縮自己着火と強制点火とを選択し、特殊運転モード領域Ａのうちで自己着火が行われ易い高速、高負荷側の領域ではエンジンの温度状態に関わらず圧縮自己着火モードとするようにしてもよい。
【００８８】
▲３▼基本実施形態では弁停止機構を用いて２気筒接続状態と各気筒独立状態とに吸・排気流通状態を切換可能としているが、吸・排気通路及び気筒間ガス通路に開閉弁を設けてこれらの通路の開閉により２気筒接続状態と各気筒独立状態とに切換え得るようにしておいてもよい。
【００８９】
▲４▼本発明の装置は４気筒以外の多気筒エンジンにも適用可能である。そして、例えば６気筒等では１つの気筒の排気行程と別の気筒の吸気行程が完全に重なり合うことはないが、このような場合は、一方の気筒の排気行程が他方の気筒の吸気行程より先行するとともに、両行程が部分的に重なり合う２つの気筒を先行、後続の一対の気筒とすればよい。
【００９０】
【発明の効果】
以上のように本発明の制御装置によると、特殊運転モードとされた場合に、排気行程と吸気行程が重なる一対の気筒のうちの先行気筒ではリーン空燃比で燃焼を行わせ、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して、理論空燃比で燃焼を行わせるようにしているため、先行気筒でのリーン燃焼による熱効率向上およびポンピングロス低減、ならびに後続気筒でのポンピングロス低減等により、燃費を改善することができ、しかも、後続気筒における燃焼の際の空燃比が実質的に理論空燃比となるようにしているため、排気通路での排気ガスの浄化を三元触媒だけで充分に行うことでき、リーンＮＯｘ触媒が不要となる。
【００９１】
そして、上記特殊運転モードでの制御時に後続気筒での燃焼を圧縮自己着火で行わせるか強制点火で行わせるかをエンジンの状態に応じて選択するようにしているため、圧縮自己着火が選択された場合に、後続気筒での圧縮自己着火による燃焼効率の向上により、燃費改善効果がさらに高められる。
【００９２】
特に本発明では、後続気筒における燃焼のための着火が圧縮自己着火とされる場合、先行気筒の空燃比が理論空燃比よりも大きく、かつ、強制点火とされる場合と比べて小さい値となるように燃料供給量を制御しているため、先行気筒から後続気筒へ導入されるガスの温度を高めて自己着火性を向上するとともに、ノッキング抑制作用を高め、後続気筒での圧縮自己着火を良好に行わせることができる。
【図面の簡単な説明】
【図１】本発明の一実施形態による制御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】制御系統のブロック図である。
【図４】運転状態に応じた制御を行うための運転領域設定の一例を示す説明図である。
【図５】ＥＣＵによる制御の具体例を示すフローチャートである。
【図６】特殊運転モードのうちの圧縮自己着火モードにあるときの、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７】特殊運転モードのうちの強制点火モードにあるときの、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８】特殊運転モードにあるとき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９】通常運転モードにあるとき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。
【符号の説明】
１　エンジン本体
２Ａ〜２Ｄ　気筒
７　点火プラグ
８　点火回路
９　燃料噴射弁
１１　吸気ポート
１１ａ　第１吸気ポート
１１ｂ　第２吸気ポート
１２　排気ポート
１２ａ　第１排気ポート
１２ｂ　第２排気ポート
１５　吸気通路
２０　排気通路
２２　気筒間ガス通路
３５　弁停止機構
４０　ＥＣＵ
４１　運転状態判別手段
４２　温度状態判別手段
４３　モード設定手段
４４　弁停止機構制御手段
４５　吸入空気量制御手段
４６　燃料制御手段
４７　着火制御手段[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a control device for a spark ignition type four-cycle engine, and more particularly to a device for controlling the combustion state of each cylinder in a multi-cylinder engine in order to improve fuel efficiency and emission.
[0002]
[Prior art]
Conventionally, in a spark ignition type engine, there is known a technology for improving fuel efficiency by performing combustion in a state in which an air-fuel ratio of an air-fuel mixture in each cylinder is set to a lean air-fuel ratio larger than a stoichiometric air-fuel ratio. As disclosed in JP-A-10-274085, a fuel injection valve for directly injecting fuel into a combustion chamber is provided, and in a low-speed low-load region or the like, stratified combustion is performed by injecting fuel from the fuel injection valve in a compression stroke. It is known to perform super-lean combustion.
[0003]
In such an engine, an ordinary three-way catalyst (a catalyst having a high purification performance near the stoichiometric 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 for adsorbing NOx in an oxygen-excess atmosphere and desorbing and reducing NOx in an oxygen-low concentration atmosphere is provided as shown in the above-mentioned publication. When such a lean NOx catalyst is used, if the amount of NOx adsorbed by the lean NOx catalyst increases during the lean operation, additional fuel is injected during the expansion stroke other than the main combustion, for example, as described in the above publication. As a result, the air-fuel ratio of the exhaust gas is enriched and CO is generated, thereby promoting the separation and reduction of NOx.
[0004]
[Problems to be solved by the invention]
In the engine performing the conventional lean operation as described above, the lean NOx catalyst is required to secure NOx purification performance during the lean operation, which is disadvantageous in cost. Further, in order to maintain the purification performance of the lean NOx catalyst, it is necessary to temporarily enrich the air-fuel ratio by supplying additional fuel for releasing and reducing NOx when the NOx adsorption amount increases as described above. In addition, when the fuel used contains a large amount of sulfur, it is necessary to regenerate the lean NOx catalyst by heating the catalyst and supplying a reducing material in order to eliminate the sulfur poisoning of the lean NOx catalyst. I do.
[0005]
In addition, when the air-fuel ratio becomes lean to a certain degree or more, the combustion speed becomes too slow and the combustion near the end of the combustion does not contribute to the work, so that there is a limit to the improvement in fuel efficiency by leaning in stratified combustion.
[0006]
As another method for improving fuel efficiency, compression self-ignition has been studied.This compression self-ignition is performed by raising the temperature and pressure in the combustion chamber at the end of the compression stroke to make the fuel self-ignite at the end of the compression stroke, similar to a diesel engine. Even if the air-fuel ratio is super lean or a large amount of EGR is introduced, if such compression self-ignition is performed, the entire combustion chamber will burn at once, and slow combustion that does not contribute to work will occur. Avoided, which is advantageous for improving fuel efficiency. However, in a normal spark ignition type engine (gasoline engine), forced ignition is required for combustion, and a special device for greatly increasing the temperature or pressure in the combustion chamber to perform compression self-ignition is required. Therefore, it has been difficult to increase the temperature or pressure in the combustion chamber to such an extent that compression self-ignition occurs in a partial load region where improvement in fuel consumption is required while avoiding knocking in a high load region.
[0007]
Therefore, the present applicant is in the exhaust stroke between a pair of cylinders where the exhaust stroke and the intake stroke overlap in the partial load region of the engine in order to achieve a significant fuel efficiency improvement by using the lean combustion and the compression self-ignition together. The burned gas discharged from the preceding cylinder is directly connected to the succeeding cylinder in the intake stroke via the inter-cylinder gas passage in a two-cylinder connection state, and the leading cylinder has an air-fuel ratio of lean air larger than the stoichiometric air-fuel ratio. A fuel-ignition-type engine, in which combustion is performed by forced ignition and fuel is supplied to burned gas having a lean air-fuel ratio introduced from a preceding cylinder in a succeeding cylinder and combustion is performed by compression self-ignition. We have applied for a technology relating to a control device (Japanese Patent Application No. 2002-29836).
[0008]
According to the present invention, based on such a technique, in the partial load region of the engine, in the two-cylinder connection state, lean combustion is performed in the preceding cylinder and combustion by compression self-ignition is performed in the succeeding cylinder as much as possible. The present invention provides a control device for a spark-ignition type four-stroke engine, which can improve the self-ignitability of the engine and effectively improve the fuel efficiency and emission.
[0009]
[Means for Solving the Problems]
The invention according to claim 1 is a multi-cylinder spark ignition type four-stroke engine in which the combustion cycle of each cylinder is performed with a predetermined phase difference, and the intake, exhaust and combustion of the engine in a partial load region of the engine. The control mode for the state is a special operation mode. In this special operation mode, the burned gas discharged from the preceding cylinder in the exhaust stroke between the pair of cylinders in which the exhaust stroke and the intake stroke overlap is the subsequent cylinder in the intake stroke as it is. The lean air-fuel ratio of the preceding cylinder is larger than the stoichiometric air-fuel ratio while the two cylinders are connected so that the gas discharged from the succeeding cylinder is guided to the exhaust passage through the inter-cylinder gas passage. In the following cylinder, fuel is supplied to burned gas having a lean air-fuel ratio introduced from the preceding cylinder, and combustion is performed substantially at the stoichiometric air-fuel ratio. An ignition control means for selecting according to the state of the engine whether to perform combustion in the subsequent cylinder by compression self-ignition or forced ignition during control in the special operation mode. In the control in the special operation mode, when the combustion in the subsequent cylinder is performed by compression self-ignition according to the selection by the ignition control means, the air-fuel ratio of the preceding cylinder is larger than the stoichiometric air-fuel ratio, and And a fuel control means for controlling the fuel supply amount so as to have a smaller value than when the combustion in the subsequent cylinder is performed by forced ignition.
[0010]
According to the present invention, when the combustion is performed by the compression self-ignition in the subsequent cylinder when the special operation mode is set and the subsequent cylinder is used, a fuel efficiency improvement effect is obtained in the preceding cylinder by improving thermal efficiency by lean combustion and reducing pumping loss. An improvement in fuel efficiency can be obtained by improving combustion efficiency and reducing pumping loss by compression self-ignition. Further, since the gas discharged from the subsequent cylinder to the exhaust passage has a stoichiometric air-fuel ratio, the exhaust gas can be sufficiently purified only by the three-way catalyst.
[0011]
Further, when the succeeding cylinder is set to the compression self-ignition, the fuel supply amount to the preceding cylinder is increased so that the air-fuel ratio in the preceding cylinder becomes a smaller value than the case where the subsequent cylinder is set to the forced ignition. The temperature of the gas introduced from the first cylinder to the subsequent cylinder is increased to improve the self-ignition property, and the knocking suppressing effect is enhanced by increasing the burned gas component corresponding to EGR in the gas.
[0012]
In the present invention, a means for judging the temperature state of the engine is provided, and the ignition control means controls the compression self-ignition and the forced ignition in accordance with the temperature state of the engine even in the same operation region during control in the special operation mode. It is preferable to select In this case, during the control in the special operation mode, the ignition control means performs the forced ignition when the engine temperature is lower than a predetermined temperature even after the engine is warmed up, and the compression self-ignition when the engine temperature is higher than the predetermined temperature. What is necessary is just to control so that it may become ignition.
[0013]
In this way, even if the engine temperature is relatively low even after the engine is warmed up, if the engine is in a situation where self-ignition is difficult, forced ignition is performed, and if the engine temperature becomes somewhat high, compression self-ignition is effectively performed. .
[0014]
The fuel control device further includes means for determining the octane number of the fuel, and the fuel control means controls the combustion in the subsequent cylinder during control in the high speed, high load side region of the operation region set as the special operation mode. Is performed by compression self-ignition, it is preferable to correct the fuel supply amount so that the lower the octane number, the smaller the air-fuel ratio of the preceding cylinder, according to the determination of the octane number.
[0015]
In this way, the combustion state is controlled such that the lower the octane number indicating the antiknock property of the fuel, the more the knocking due to the combustion by the compression self-ignition in the subsequent cylinder is suppressed. That is, when the air-fuel ratio of the preceding cylinder is reduced by increasing the fuel supply amount to the preceding cylinder, the temperature of the gas introduced into the succeeding cylinder increases, but the burned gas corresponding to the EGR in the gas to the succeeding cylinder increases Under the condition that the abrupt combustion in the following cylinder is alleviated by the increase in the component and the air-fuel ratio of the following cylinder is set to approximately the stoichiometric air-fuel ratio, the amount of fuel supplied to the following cylinder is increased by the amount of fuel supply to the preceding cylinder. The supply amount is reduced, the combustion energy in the subsequent cylinder is reduced, and knocking in the subsequent cylinder is suppressed by these actions. Then, the compression auto-ignition region can be expanded by the amount by which knocking is suppressed in the region on the high-speed, high-load side.
[0016]
Further, when the combustion in the subsequent cylinder is performed by compression self-ignition during control by the mode in the low speed, low load side region of the operation region that is the special operation mode, according to the determination of the octane number, It is also preferable to correct the fuel supply amount so that the higher the octane number, the smaller the air-fuel ratio of the preceding cylinder.
[0017]
In this way, the self-ignition property is improved in a situation where self-ignition is difficult. That is, in the low-speed, low-load range, the fuel supply amount is small and the heat generation amount is small, which is disadvantageous for self-ignition. Further, if the octane number of the fuel is high, the self-ignition is more disadvantageous due to the high anti-knock property. It becomes. In such a case, if the air-fuel ratio of the preceding cylinder is reduced by increasing the fuel supply amount to the preceding cylinder, the amount of heat generated in the preceding cylinder increases, and the temperature of the gas introduced into the succeeding cylinder increases. However, even under the above-mentioned circumstances that are disadvantageous to the self-ignition, the compression self-ignition is sufficiently performed.
[0018]
Further, the ignition control means may select the compression self-ignition and the forced ignition in a low-speed region of an operation region set as the special operation mode. That is, since the temperature rise in the combustion chamber during compression is essentially smaller in the low speed range than in the middle and high speed ranges, if conditions that are disadvantageous to compression self-ignition such as low engine temperature are applied in this range, , Forced ignition.
[0019]
Further, when the combustion in the subsequent cylinder is performed by forced ignition during the control in the special operation mode, the fuel control means has an air-fuel ratio of the preceding cylinder larger than twice the stoichiometric air-fuel ratio. It is preferable to control the fuel supply amount as described above. By doing so, the effect of improving fuel efficiency by lean combustion in the preceding cylinder is enhanced as compared with the case where the air-fuel ratio of the preceding cylinder is not more than twice the stoichiometric air-fuel ratio, while the temperature of the gas introduced into the succeeding cylinder is lower. However, combustion is ensured by forced ignition.
[0020]
Further, the ignition control means, when performing the compression self-ignition during the control in the special operation mode, performs backup for the subsequent cylinder at a timing retarded by a predetermined amount from the ignition timing when performing the forced ignition. It is preferable to cause ignition for In this case, the ignition timing for the backup may be set after the compression top dead center and near the compression top dead center.
[0021]
With this configuration, when the compression self-ignition is performed, even if a situation occurs in which the compression self-ignition is not properly performed for some reason, the ignition and the combustion are performed by the backup ignition. In addition, it is possible to avoid deterioration of emission and the like.
[0022]
BEST MODE FOR CARRYING OUT THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
[0023]
FIG. 1 shows a schematic structure of an engine according to an 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 2A to 2D, and a combustion chamber 4 is formed above the piston 3.
[0024]
A spark plug 7 is provided at the top of the combustion chamber 4 of each cylinder 2, and the tip of the plug faces the combustion chamber 4. The ignition plug 7 is connected to an ignition circuit 8 capable of controlling the ignition timing by electronic control.
[0025]
A fuel injection valve 9 for directly injecting fuel into the combustion chamber 4 is provided on a side portion of the combustion chamber 4. The fuel injection valve 9 has a built-in needle valve and a solenoid (not shown). When a pulse signal described later is input, the fuel injection valve 9 is driven for a time corresponding to the pulse width at the pulse input time, and opens. It is configured to inject an amount of fuel according to the valve time. Fuel is supplied to the fuel injection valve 9 through a fuel supply passage or the like by a fuel pump (not shown), and a fuel supply system is provided so as to provide a fuel pressure higher than the pressure in the combustion chamber during the compression stroke. Is configured.
[0026]
In addition, intake ports 11, 11a, 11b and exhaust ports 12, 12a, 12b are opened to the combustion chambers 4 of the cylinders 2A to 2D, and these ports are connected to an intake passage 15, an exhaust passage 20, and the like. Each port is opened and closed by intake valves 31, 31a, 31b and exhaust valves 32, 32a, 32b.
[0027]
Each cylinder performs a cycle including intake, compression, expansion, and exhaust strokes with a predetermined phase difference. In the case of a four-cylinder engine, the first cylinder 2A and the second cylinder 2A are arranged from one end in the cylinder row direction. When the cylinders are referred to as cylinders 2B, 3C, and 4D, as shown in FIGS. 5 and 6, the above-described cycle is performed in the order of the first cylinder 2A, the third cylinder 2C, the fourth cylinder 2D, and the second cylinder 2B. This is performed with a phase difference of 180 ° at each angle. 6 and 7 show the stroke, fuel injection timing, ignition timing, etc. of each cylinder in a four-cycle four-cylinder engine. As will be described in detail later, FIG. FIG. 7 shows a case where the subsequent cylinder is set to forced ignition in the special operation mode. In these figures, EX indicates an exhaust stroke, IN indicates an intake stroke, F indicates fuel injection, S indicates forced ignition, and the star mark in the figures indicates that compression self-ignition is performed.
[0028]
Between a pair of cylinders where the exhaust stroke and the intake stroke overlap, between the cylinder on the exhaust stroke side (hereinafter referred to as a preceding cylinder in this specification) and the cylinder on the intake stroke side (this specification) when the exhaust stroke and the intake stroke overlap. In this case, an inter-cylinder gas passage 22 is provided so that the burned gas can be directly guided to the subsequent cylinder. In the four-cylinder engine of the present embodiment, as shown in FIGS. 6 and 7, the exhaust stroke (EX) of the first cylinder 2A and the intake stroke (IN) of the second cylinder 2B overlap, and the exhaust of the fourth cylinder 2D. Since the stroke (EX) and the intake stroke (IN) of the third cylinder 2C overlap, 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. The fourth cylinder 2D is the preceding cylinder, the second cylinder 2B, and the third cylinder 2C are the succeeding cylinders.
[0029]
The intake / exhaust ports of each cylinder and the intake passage, exhaust passage and inter-cylinder gas passage connected thereto are specifically configured as follows.
[0030]
The first cylinder 2A and the fourth cylinder 2D, which are the preceding cylinders, have an intake port 11 for introducing fresh air and a first exhaust port 12a for sending burned gas (exhaust gas) to an exhaust passage, respectively. , And a second exhaust port 12b for leading burned gas to a subsequent cylinder. A second intake port 11a for introducing fresh air and a second intake port for introducing burned gas from the preceding cylinder are respectively provided to the second cylinder 2B and the third cylinder 2C, which are subsequent cylinders. 11b and an exhaust port 32 for sending burned gas to an exhaust passage are provided.
[0031]
In the example shown in FIG. 1, the number of intake ports 11 in the first and fourth cylinders 2A and 2D and the number of first intake ports 11a in the second and third cylinders 2B and 2C are two for each cylinder and 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. 12 are provided in parallel on the right half side of the combustion chamber.
[0032]
The downstream end of the cylinder-specific branch intake passage 16 in the intake passage 15 is connected to 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. I have. In the vicinity of the downstream end of each branch intake passage 16, a multiple throttle valve 17 interlocking with each other via a common shaft is provided, and the multiple throttle valve 17 is driven by an actuator 18 according to a control signal, The intake air volume is adjusted. Note that an airflow sensor 19 for detecting an intake air flow rate is provided in a common intake passage upstream of the collecting portion in the intake passage 15.
[0033]
The 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. I have. Further, inter-cylinder gas passages 22 are provided between the first cylinder 2A and the second cylinder 2B and between the third cylinder 2C and the fourth cylinder 2D, respectively, and the first and fourth cylinders 2A and 2A, which are the preceding cylinders, are provided. 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 ports 11b of the second and third cylinders 2B and 2C that are subsequent cylinders. Is connected.
[0034]
The inter-cylinder gas passage 22 is a relatively short passage connecting between adjacent cylinders, and heat radiation while gas discharged from the preceding cylinder passes through the passage 22 is relatively small. .
[0035]
An O-fuel ratio is detected by detecting the oxygen concentration in the exhaust gas at a collecting portion of the exhaust passage 20 downstream of the branch exhaust passage 21. ₂ A sensor 23 is provided. Further O ₂ A three-way catalyst 24 is provided in the exhaust passage 21 downstream of the sensor 23 for purifying exhaust gas. As is generally known, the three-way catalyst 24 purifies HC, CO, and NOx when the air-fuel ratio of the exhaust gas is near the stoichiometric air-fuel ratio (that is, when the excess air ratio λ is λ = 1). It is a catalyst that shows performance.
[0036]
The intake / exhaust valves for opening and closing the intake / exhaust ports of each cylinder and the valve operating mechanism for these valves are as follows.
[0037]
The intake port 11, the first exhaust port 12a, and the second exhaust port 12b of the first and fourth cylinders 2A, 2D are provided with an intake valve 31, a first exhaust valve 32a, and a second exhaust valve 32b, respectively. A first intake valve 31a, a second intake valve 31b, and an exhaust valve 32 are provided at the first intake port 11a, the second intake port 11b, and the exhaust port 12, respectively, of the third and third cylinders 2B, 2C. Then, these intake and exhaust valves are opened and closed at predetermined timings by a valve mechanism including the camshafts 33 and 34 so that the intake stroke and the exhaust stroke of each cylinder are performed with the above-described predetermined phase difference. Driven as follows.
[0038]
Further, among these intake / exhaust valves, for the first exhaust valve 32a, the second exhaust valve 32b, the first intake valve 31a, and the second intake valve 31b, each valve is switched between an operating state and a stopped state. A valve stop mechanism 35 is provided. The valve stop mechanism 35 is conventionally known, so a detailed illustration thereof is omitted. For example, hydraulic oil can be supplied and discharged to and from a tappet interposed between the cams of the camshafts 33 and 34 and the valve shaft. When the hydraulic oil is supplied to the hydraulic chamber, the operation of the cam is transmitted to the valve to open and close the valve, and when the hydraulic oil is discharged from the hydraulic chamber, the operation of the cam is controlled by the valve. The valve is stopped because it cannot be communicated to.
[0039]
A first control valve 37 is provided in a passage 36 for supplying and discharging hydraulic oil to the valve stop mechanism 35 of the first exhaust valve 32a and the valve stop mechanism 35 of the first intake valve 31a, and a valve stop of the second exhaust valve 32b. A second control valve 39 is provided in a passage 38 for supplying and discharging hydraulic oil to the mechanism 35 and the valve stop mechanism 35 of the second intake valve 31b (see FIG. 3).
[0040]
FIG. 3 shows the configuration of the drive and control system. In this figure, an engine control ECU (control unit) 40 including a microcomputer or the like includes an airflow sensor 19 and an O ₂ A signal from a sensor 23 is inputted, a signal from a water temperature sensor for detecting a temperature of a cooling water of the engine is inputted, and a rotation speed sensor 52 for detecting an engine speed and an accelerator opening to further determine an operation state. A signal from an accelerator opening sensor 53 or the like for detecting (accelerator pedal depression amount) is also input. Control signals are output from the ECU 40 to each of the fuel injection valves 9, the actuator 18 of the multiple throttle valve 17, and the first and second control valves 39.
[0041]
The ECU 40 includes an operation state determination unit 41, a temperature state determination unit 42, a mode setting unit 43, a valve stop mechanism control unit 44, an intake air amount control unit 45, a fuel control unit 46, and an ignition control unit 47.
[0042]
As shown in FIG. 4, the operating state determining means 41 generates a control map in which the operating region of the engine is divided into a low-speed low-load region A (partial load region) and a high-speed or high load region B. The area A on the low speed and low load side is a special operation mode area, and the area B on the high speed or high load side is a normal operation mode area. Then, it is determined which of the regions A and B the operating state (engine speed and engine load) of the engine, which is checked by signals from the rotation speed sensor 52 and the accelerator opening sensor 53, is.
[0043]
Further, when the operating state is in the special operation mode area A, the operating state determining means 41 determines whether the operating state is in the area A1 on the low-speed and low-load side of the area A or in the other area. Has become.
[0044]
The temperature state determination means 42 determines the temperature state of the engine based on a signal from the water temperature sensor 51, and determines whether the water temperature (engine temperature) is low or lower than a predetermined value or higher than a predetermined temperature. It has become.
[0045]
The mode setting means 43 introduces the burned gas discharged from the preceding cylinder in the exhaust stroke as it is into the subsequent cylinder in the intake stroke in the special operation mode area A based on the determination by the operating state determination means 41. The special operation mode in which combustion is performed is selected. In the normal operation mode region B, the normal operation mode in which each cylinder is independently operated and combustion is selected. Further, when the special operation mode is set, the forced ignition mode in which combustion in the subsequent cylinder is performed by forced ignition at the time of the low temperature is selected based on the determination of the temperature state of the engine by the temperature state determination means 42. A compression self-ignition mode in which combustion in the cylinder is performed by compression self-ignition is selected.
[0046]
According to the mode setting by the mode setting means 43, the valve stop mechanism control means 44 sets the two-cylinder connection state in which the burned gas of the preceding cylinder is introduced into the succeeding cylinder through the inter-cylinder gas passage 22 in the special operation mode, In the normal operation mode, the valve stop mechanism control 35 is controlled to change the intake / exhaust flow state so that each cylinder is in an independent state in which fresh air is introduced into each cylinder. , B, the respective valve stop mechanisms 35 are controlled as follows by controlling the respective control valves 37, 39.
[0047]

The intake air amount control means 45 controls the opening degree of the throttle valve 17 (throttle opening degree) by controlling the actuator 18, and obtains a target intake air amount from a map or the like according to an operation state. The throttle opening is controlled according to the target intake air amount. In this case, in the operation region A where the special operation mode is set, the gas introduced from the preceding cylinder in a state where the intake of the intake air from the branch intake passage 16 is cut off in the subsequent cylinders (the second and third cylinders 2B and 2C). Since the combustion is performed while the ratio of the excess air in the air to the newly supplied fuel is set to the lean air-fuel ratio, the amount of air necessary for combustion of the fuel in accordance with the required torque of the preceding and succeeding two cylinders ( The throttle opening is adjusted such that the stoichiometric air-fuel ratio of the amount of fuel for the two cylinders is supplied to the preceding cylinders (the first and fourth cylinders 2A and 2D).
[0048]
The fuel control means 46 controls the fuel injection amount and the injection timing from the fuel injection valve 9 provided in each of the cylinders 2A to 2D according to the operating state of the engine. The control of the ignition timing and the control of the stop of the ignition are performed. The control of the combustion state (control of fuel and control of ignition) is changed according to the mode set by the mode setting means 43 in particular.
[0049]
That is, when the special operation mode is set, for the preceding cylinders 2A and 2D, the fuel injection amount is controlled so that the air-fuel ratio becomes a lean air-fuel ratio larger than the stoichiometric air-fuel ratio, and the fuel is supplied in the compression stroke. The injection timing is set so as to cause stratification of the air-fuel mixture by injection, and the ignition timing is set so as to perform forced ignition near the compression top dead center.
[0050]
On the other hand, fuel is supplied to the succeeding cylinders 2B and 2C with respect to burned gas having a lean air-fuel ratio introduced from the preceding cylinders 2A and 2D, and substantially combusted in the subsequent cylinders 2B and 2C. The fuel injection amount is controlled so as to achieve the stoichiometric air-fuel ratio. Then, when the compression self-ignition mode is selected in the special operation mode, the injection timing is set so as to inject fuel in the intake stroke so as to equalize the air-fuel mixture, and the compression self-ignition is performed. , Stop the forced ignition. When the forced ignition mode is selected, the injection timing is set so as to inject the fuel in the compression stroke, and the ignition timing is set so that the forced ignition is performed at a predetermined timing near the compression top dead center.
[0051]
Further, in this special operation mode, the compression self-ignition mode is set while the sum of the fuel injection amounts for both the pair of cylinders is adjusted to the stoichiometric air-fuel ratio with respect to the amount of air introduced into the preceding cylinders 2A and 2D. The ratio between the fuel injection amount for the preceding cylinders 2A and 2D and the fuel injection amount for the succeeding cylinders 2B and 2C is changed between the case where the selected cylinder is selected and the case where the forced ignition mode is selected, so that the preceding cylinders 2A and 2D are changed. Is changed. When the compression self-ignition mode is selected, the air-fuel ratio of the preceding cylinders 2A and 2D becomes larger than the stoichiometric air-fuel ratio (excess air ratio λ is 1 <λ), and the forced ignition mode is selected. The fuel injection amount is adjusted to be smaller.
[0052]
Specifically, in the forced ignition mode, the fuel injection amount for the preceding cylinders 2A, 2D is set to be equal to or less than the fuel injection amount for the following cylinders 2B, 2C, so that the air-fuel ratios of the preceding cylinders 2A, 2D become stoichiometric. In contrast, in the compression self-ignition mode, the fuel injection amount for the preceding cylinders 2A and 2D is larger than the fuel injection amount for the succeeding cylinders 2B and 2C, while the fuel injection ratio is about twice (A / F ≒ 30) or more. Then, the air-fuel ratio of the preceding cylinders 2A and 2D is set to, for example, A / F ≒ 25.
[0053]
Note that the air-fuel ratio of the preceding cylinders 2A and 2D in the compression self-ignition mode may be changed according to the operating state. In this case, as will be described in detail later, in order to improve the self-ignition property and to suppress knocking, the low-speed low-load side and the high-speed high load side in the special operation mode area A are used. It is preferable that the air-fuel ratio of the preceding cylinders 2A and 2D is made smaller (the ratio of the fuel supply amount to the preceding cylinders 2A and 2D is increased) as compared with the region.
[0054]
When the normal operation mode is selected, the fuel injection amount is controlled so that the air-fuel ratio of each of the cylinders 2A to 2D is equal to or lower than the stoichiometric air-fuel ratio. The stoichiometric air-fuel ratio is set in the region, and the stoichiometric air-fuel ratio is set to be richer in the full open load and the operating region in the vicinity thereof. In this case, the injection timing is set such that fuel is injected into each of the cylinders 2A to 2D in the intake stroke so as to equalize the air-fuel mixture, and the cylinders 2A to 2D are also forcedly ignited. To
[0055]
The ECU 40 preferably further includes means for determining the octane number of the fuel, and as shown in a flowchart described later, in the case of the compression auto-ignition mode in the special operation mode, the low-speed, low-load side of this area A. The correction of the air-fuel ratio of the preceding cylinder (correction of the fuel injection amount) is performed according to whether or not the fuel cell is in the region A1 and the octane value of the fuel.
[0056]
FIG. 5 is a flowchart showing the processing performed by each unit included in the ECU 40.
[0057]
When the process of this flowchart is started, first, various signals are input in step S1, and then, in step S2, it is determined whether or not the operation state is in the special operation mode area A.
[0058]
If it is in the special operation mode area A, the intake / exhaust valve is controlled to the special operation mode in step S3, that is, the intake / exhaust flow state becomes the two-cylinder connection state by the control of the valve stop mechanism mechanism 35 as described above. The throttle valve 17 is controlled so that the target intake air amount in the special operation mode is obtained in step S4. Further, in step S5, it is determined whether the temperature state of the engine is a low temperature equal to or lower than a predetermined temperature.
[0059]
If the determination in step S5 is YES, the forced ignition mode is selected in step S6, and the ignition timing of each of the cylinders 2A to 2D is set so that the preceding cylinders 2A and 2D and the following cylinders 2B and 2C are both forcedly ignited. At the same time, the fuel injection amount is calculated in step S7. In this case, the excess air ratio λ in the preceding cylinders 2A, 2D is set to a predetermined value λa larger than 1 (eg, λa ≒ 2), and the excess air ratio λ in the succeeding cylinders 2B, 2C is set to λ ≒ 1. Next, the fuel injection amount for each of the cylinders 2A to 2D is calculated. Further, in step S8, the fuel injection timing of each of the cylinders 2A to 2D is set such that the preceding cylinders 2A and 2D perform the compression stroke injection for stratified combustion, and the succeeding cylinders 2B and 2C also perform the compression stroke injection for ensuring ignitability. It is calculated. Then, in step S9, control of fuel injection and ignition is performed in accordance with the calculated or set fuel injection amount, fuel injection timing, and ignition timing for each of the cylinders 2A to 2D.
[0060]
If the determination in step S5 is NO, that is, if the engine is at a temperature higher than the predetermined temperature, the compression auto-ignition mode is selected in step S10, and ignition is performed so that only the preceding cylinders 2A and 2D are forcibly ignited. The timing is set, and the fuel injection amount is calculated in step S11. In this case, the excess air ratio λ in the preceding cylinders 2A and 2D is set to a value satisfying 1 <λb <λa, and the excess air ratio λ in the subsequent cylinders 2B and 2C is set to λ ≒ 1. The fuel injection amounts for 2A to 2D are calculated.
[0061]
Subsequently, in step S12, it is determined whether or not it is in the area A1 of the special operation mode area A. If the determination is YES, it is determined in step S13 whether or not the octane value of the fuel is high. If the determination is NO, it is determined in step S14 whether the octane number of the fuel is low.
[0062]
Based on these determinations, if the operating state is within the region A1 and the octane number of the fuel is high, the air-fuel ratios of the preceding cylinders 2A and 2D are corrected to rich according to the octane number in step S15. Is corrected so that the higher the octane number, the smaller the air-fuel ratio of the preceding cylinders 2A, 2D (the larger the fuel injection amount). Also, when the operation state is outside the above-mentioned area A1 in the special operation mode area A and the octane number of the fuel is low, the process proceeds to step S15 and the correction according to the octane number is performed. The correction is performed so that the air-fuel ratio of the cylinders 2A and 2D decreases (the fuel injection amount increases).
[0063]
If the processing in step S15 has been performed, or if the determination in step S13 or step S14 is NO, the process proceeds to step S8 to calculate the fuel injection timing of each of the cylinders 2A to 2D. In 2A and 2D, compression stroke injection is performed due to stratified combustion, while in subsequent cylinders 2B and 2C, intake stroke injection is performed. Then, control of fuel injection and ignition is performed in step S9.
[0064]
If the determination in step S2 is NO, that is, if the operation state is in the normal operation mode region B, the intake and exhaust valves are controlled to the normal operation mode in step S16, that is, the valve stop mechanism as described above. The control of the mechanism 35 controls the intake / exhaust flow state so as to be an independent state for each cylinder, and the throttle valve 17 is controlled in step S17 so as to obtain the target intake air amount in the normal operation mode. Further, in step S18, the ignition timing is set so that the cylinders 2A to 2D are also forcedly ignited. In step S19, each of the cylinders 2A to 2D is set so that the excess air ratio λ is 1 or less. The fuel injection amount for each of the cylinders 2A to 2D is calculated, and further, in step S20, the fuel injection timing is calculated so that all cylinders perform the intake stroke injection for uniform combustion. Then, control of fuel injection and ignition is performed in step S9.
[0065]
The operation of the apparatus of the present embodiment as described above will be described with reference to FIGS.
[0066]
In the special operation mode area A on the low speed and low load side, the special operation mode is set, and the first exhaust valve 32a and the first intake valve 31a are stopped, and the second exhaust valve 32b and the second intake valve 31b operate as described above. As a result, the substantial fresh air and gas flow paths are as shown in FIG. 8, and the burned gas discharged from the preceding cylinders 2A and 2D is passed through the inter-cylinder gas passage 22 as it is to the subsequent cylinders. The two-cylinder connection state is such that only the gas that is introduced into the following cylinders 2B and 2C is introduced into the exhaust passage 20 while being introduced into the following cylinders 2B and 2C.
[0067]
In this state, fresh air is introduced into the preceding cylinders 2A and 2D from the intake passage 15 in the intake stroke (arrow a in FIG. 8), and the lean air-fuel ratio of the preceding cylinders 2A and 2D is larger than the stoichiometric air-fuel ratio. The fuel is injected in the compression stroke while the fuel injection amount is controlled such that the stratified combustion is performed at a predetermined ignition timing (see FIGS. 6 and 7).
[0068]
Then, during a period in which the intake strokes of the preceding cylinders 2A, 2D and the exhaust strokes of the following cylinders 2B, 2C overlap, the burned gas discharged from the preceding cylinders 2A, 2D is introduced into the following cylinders 2B, 2C through the gas passage 22. (Open arrows in FIGS. 6 and 7 and arrow b in FIG. 8). Then, in the subsequent cylinders 2B and 2C, fuel is supplied to the burned gas having the lean air-fuel ratio introduced from the preceding cylinders 2A and 2D, and combustion is performed while controlling the fuel injection amount so as to achieve the stoichiometric air-fuel ratio. .
[0069]
In this case, when the engine is at a temperature higher than the predetermined temperature, the compression auto-ignition mode is selected, and as shown in FIG. 6, in the subsequent cylinders 2B and 2C, after the fuel is injected in the intake stroke, the compression stroke is increased. In the vicinity of the dead center, the combustion chamber is in a sufficiently high temperature and high pressure state, and the compression self-ignition is performed favorably.
[0070]
That is, the high-temperature burned gas discharged from the preceding cylinders 2A, 2D is immediately introduced into the succeeding cylinders 2B, 2C through the short inter-cylinder gas passage 22, so that the succeeding cylinders 2B, 2C take the inside of the combustion chamber during the intake stroke. As the temperature rises and the pressure and temperature further increase in the compression stroke from this state, the temperature in the combustion chamber rises to the extent that the air-fuel mixture can self-ignite near the top dead center at the end of the compression stroke. In addition, the burned gas is sufficiently mixed and uniformly distributed before being discharged from the preceding cylinders 2A and 2D and introduced into the following cylinders 2B and 2C, and the fuel injected during the intake stroke is also compressed. Since it is uniformly dispersed throughout the combustion chamber by the end, a uniform mixture distribution that satisfies the ideal simultaneous compression auto-ignition condition can be obtained. Then, the combustion is rapidly performed by the simultaneous compression self-ignition, thereby significantly improving the thermal efficiency.
[0071]
As described above, in the preceding cylinders 2A and 2D, the thermal efficiency is improved by the stratified combustion in the lean state, and the pumping loss is reduced by reducing the intake negative pressure as compared with a normal engine that does not perform the stratified combustion. In the subsequent cylinders 2B and 2C, while the air-fuel ratio is set to substantially the stoichiometric air-fuel ratio, the compression self-ignition is performed in a uniform mixture distribution state, thereby increasing the thermal efficiency, and the gas extruded from the preceding cylinders 2A and 2D. The pumping loss is further reduced as compared with the preceding cylinders 2A and 2D. These actions greatly improve fuel economy.
[0072]
In addition, since the gas discharged from the subsequent cylinders 2B and 2C to the exhaust passage 20 has the stoichiometric air-fuel ratio, it is not necessary to provide a lean NOx catalyst as in the conventional lean burn engine, and the three-way catalyst 24 is sufficiently exhausted. Purification performance is ensured.
[0073]
Since there is no need to provide a lean NOx catalyst, there is no need to temporarily enrich the air-fuel ratio for the release and reduction of NOx when the NOx storage amount of the lean NOx catalyst increases, and the reduction in fuel efficiency is reduced. can avoid. Further, the problem of sulfur poisoning of the lean NOx catalyst does not occur.
[0074]
Further, in the preceding cylinders 2A and 2D, the lean air-fuel ratio is set to be approximately twice the stoichiometric air-fuel ratio or close to the stoichiometric air-fuel ratio, so that the NOx generation amount can be suppressed relatively small. On the other hand, in the succeeding cylinders 2B and 2C, the burned gas is introduced from the preceding cylinders 2A and 2D to be in a state equivalent to that a large amount of EGR is performed, and rapid combustion by simultaneous compression self-ignition is performed. As a result, the reaction between oxygen and nitrogen is avoided as much as possible, so that the generation of NOx is sufficiently suppressed. From such a point, it is advantageous for improving the emission.
[0075]
Further, since the compression self-ignition in the succeeding cylinders 2B and 2C is achieved by utilizing the heat of the burned gas discharged from the preceding cylinders 2A and 2D, a special heating means is used or the compression ratio of the engine is extremely reduced. Compression self-ignition can be easily achieved without having to raise the height.
[0076]
In particular, when the combustion in the following cylinders 2B, 2C is performed by compression self-ignition, the air-fuel ratio of the preceding cylinders 2A, 2D is made larger than the stoichiometric air-fuel ratio and smaller than that in the case of forced ignition. This facilitates compression self-ignition, and is advantageous in suppressing knocking.
[0077]
That is, by setting the air-fuel ratio of the preceding cylinders 2A and 2D in this way, the fuel injection amount to the preceding cylinders 2A and 2D is increased as compared with the case where the forced ignition is performed, so that the preceding cylinders 2A and 2D And the temperature of the gas guided from the preceding cylinders 2A, 2D to the succeeding cylinders 2B, 2C rises, so that compression self-ignition is easily performed even at a relatively low speed and low load side. Although the temperature of the gas introduced into the subsequent cylinders 2B and 2C rises in this way, the burned gas component corresponding to the EGR in the gas introduced into the subsequent cylinders 2B and 2C increases, and the following cylinder 2B , 2C, the energy generated by the combustion in the subsequent cylinders 2B, 2C is reduced, so that knocking is suppressed even at a relatively high speed and high load.
[0078]
In particular, the air-fuel ratio of the preceding cylinders 2A and 2D is made smaller on the low speed low load side and the high speed high load side in the special operation mode area A as compared with the medium speed medium load area of the same area A (for the preceding cylinders 2A and 2D). If the fuel injection amount is controlled according to the operating state so as to increase the ratio of the fuel supply amount), the effect of improving the self-ignition property at low speed and low load and suppressing knocking at high speed and high load can be enhanced.
[0079]
Further, depending on the operating state of the engine and the octane number of the fuel, the octane number is low in the special operation mode area A on the high-speed and high-load side (in a situation where knock resistance is disadvantageous). In (), if the fuel injection amount is controlled so as to reduce the air-fuel ratio of the preceding cylinders 2A and 2D, knocking is effectively suppressed. Further, even in the low-speed, low-load region A1 of the special operation mode region A and when the octane number is high (in a situation that is disadvantageous for self-ignition), the air-fuel ratio of the preceding cylinders 2A and 2D is also reduced. If the fuel injection amount is controlled to be small, the compression self-ignition can be sufficiently performed.
[0080]
Further, when the engine is in a low temperature state below a predetermined temperature, even if the air-fuel ratio of the preceding cylinders 2A, 2D is reduced, the temperature of the gas introduced into the succeeding cylinders 2B, 2C and the compression in the succeeding cylinders 2B, 2C. In some cases, the temperature rise in the stroke is not sufficiently high and self-ignition becomes difficult. If the air-fuel ratio of the preceding cylinders 2A and 2D is too small, the combustion stability and fuel efficiency of the following cylinders 2B and 2C are disadvantageously reduced. Become.
[0081]
Therefore, in such a case, the forced ignition mode is set, and as shown in FIG. 7, in the succeeding cylinders 2B and 2C as well as in the preceding cylinders 2A and 2D, ignition is performed by a spark plug near the compression top dead center, Ignition and combustion are reliably performed by forced ignition.
[0082]
According to the forced ignition, the combustion efficiency in the subsequent cylinders 2B and 2C is slightly reduced as compared with the case where the compression self-ignition is performed well. However, as long as the special operation mode is set, the preceding cylinder 2A is used. , 2D, and fuel efficiency improvement by reducing the pumping loss in the following cylinders 2B, 2C, etc., and the gas discharged from the following cylinders 2B, 2C has the stoichiometric air-fuel ratio, so that only the three-way catalyst is used. Sufficient exhaust purification performance is ensured. In particular, when the forced ignition mode is set, the air-fuel ratio of the preceding cylinders 2A, 2D is set to approximately twice or more than the stoichiometric air-fuel ratio, thereby improving the fuel efficiency by lean burn in the preceding cylinders 2A, 2D and Combustion stability in the following cylinders 2B and 2C is ensured.
[0083]
On the other hand, in the operation region B on the high speed side or the high load side, the normal operation mode is set, the first exhaust valve 32a and the first intake valve 31a are in the operating state, and the second exhaust valve 32b and the second intake valve 31b are By being in the stopped state, the flow paths of the fresh air and gas are substantially as shown in FIG. 9, and the intake ports 31, 31a and the exhaust ports 12a, 12 of the cylinders 2A to 2D are independent, and the intake passage From 15, fresh air is introduced into the intake ports 31, 31 a of the cylinders 2 A to 2 D, and burned gas is discharged from the exhaust ports 31, 31 a of the cylinders 2 A to 2 D to the exhaust passage 20. In this case, the output performance is ensured by controlling the intake air amount and the fuel injection amount so as to be stoichiometric air-fuel ratio or richer.
[0084]
Note that the specific configuration of the device of the present invention is not limited to the above embodiment, but can be variously modified. Another embodiment will be described below.
[0085]
{Circle around (1)} When the compression self-ignition mode of the special operation mode is set, the ignition of the subsequent cylinders 2B and 2C is simply stopped in the above embodiment, but the forced ignition is performed for the subsequent cylinders 2B and 2C. The ignition S ′ for backup (indicated by a broken line in FIG. 6) may be performed at a timing retarded by a predetermined amount from the ignition timing in this case. The ignition S 'for this backup may be set after the compression top dead center and near the compression top dead center.
[0086]
In this way, even in the case where the compression self-ignition is not performed well for some reason in the compression self-ignition mode, the ignition combustion is performed by the backup S ′ and the torque fluctuation is performed. Is avoided, and deterioration of emission is prevented.
[0087]
{Circle over (2)} In the basic embodiment, control is performed such that compression self-ignition is performed at high temperature and forced ignition is performed at low temperature in the entire special operation mode area A in accordance with the temperature state of the engine. Only in the region A1), the compression self-ignition and the forced ignition are selected according to the temperature condition of the engine. In the special operation mode region A, in the high-speed, high-load region where the self-ignition is easily performed, the engine temperature condition is set. Regardless, the compression auto-ignition mode may be set.
[0088]
(3) In the basic embodiment, the intake / exhaust flow state can be switched between the two-cylinder connection state and the individual cylinder independent state by using a valve stop mechanism, but open / close valves are provided in the intake / exhaust passage and the inter-cylinder gas passage. By opening and closing these passages, it may be possible to switch between the two-cylinder connection state and each cylinder independent state.
[0089]
{Circle around (4)} The device of the present invention is applicable to multi-cylinder engines other than four cylinders. For example, in the case of six cylinders, the exhaust stroke of one cylinder and the intake stroke of another cylinder do not completely overlap. In such a case, the exhaust stroke of one cylinder precedes the intake stroke of the other cylinder. At the same time, the two cylinders in which both strokes partially overlap may be a pair of preceding and succeeding cylinders.
[0090]
【The invention's effect】
As described above, according to the control device of the present invention, when the special operation mode is set, the combustion is performed at the lean air-fuel ratio in the preceding cylinder of the pair of cylinders in which the exhaust stroke and the intake stroke overlap, and in the subsequent cylinder, Since fuel is supplied to burned gas with a lean air-fuel ratio introduced from the cylinder and combustion is performed at the stoichiometric air-fuel ratio, improvement in thermal efficiency and reduction in pumping loss due to lean combustion in the preceding cylinder, and in subsequent cylinders The fuel consumption can be improved by reducing the pumping loss at the same time, and the air-fuel ratio at the time of combustion in the subsequent cylinder is substantially set to the stoichiometric air-fuel ratio. Can be sufficiently performed only with the three-way catalyst, and the lean NOx catalyst is not required.
[0091]
Then, when controlling in the special operation mode, whether to perform combustion in the subsequent cylinder by compression self-ignition or by forced ignition is selected according to the state of the engine, so that compression self-ignition is selected. In such a case, the fuel efficiency improvement effect is further enhanced by improving the combustion efficiency by the compression self-ignition in the subsequent cylinder.
[0092]
In particular, in the present invention, when the ignition for combustion in the subsequent cylinder is compression self-ignition, the air-fuel ratio of the preceding cylinder is larger than the stoichiometric air-fuel ratio and smaller than that in the case of forced ignition. The fuel supply is controlled as described above, so that the temperature of the gas introduced from the preceding cylinder to the succeeding cylinder is raised to improve the self-ignition property, and the knocking suppression action is enhanced, and the compression self-ignition in the succeeding cylinder is improved. Can be performed.
[Brief description of the drawings]
FIG. 1 is a schematic plan view of an entire engine including a control device according to an embodiment of the present invention.
FIG. 2 is a schematic sectional view of an engine body and the like.
FIG. 3 is a block diagram of a control system.
FIG. 4 is an explanatory diagram showing an example of an operation area setting for performing control according to an operation state.
FIG. 5 is a flowchart illustrating a specific example of control by the ECU.
FIG. 6 is a diagram showing an exhaust stroke, an intake stroke, a fuel injection timing, an ignition timing, and the like of each cylinder when in a compression self-ignition mode of the special operation mode.
FIG. 7 is a diagram showing an exhaust stroke, an intake stroke, a fuel injection timing, an ignition timing, and the like of each cylinder when in a forced ignition mode of the special operation mode.
FIG. 8 is an explanatory diagram showing a substantial fresh air and gas distribution route in a special operation mode.
FIG. 9 is an explanatory diagram showing a substantial fresh air and gas distribution route in a normal operation mode.
[Explanation of symbols]
1 Engine body
2A-2D cylinder
7 Spark plug
8 Ignition circuit
9 Fuel injection valve
11 Intake port
11a 1st intake port
11b Second intake port
12 Exhaust port
12a First exhaust port
12b Second exhaust port
15 Intake passage
20 Exhaust passage
22 Gas passage between cylinders
35 Valve stop mechanism
40 ECU
41 Operating state determination means
42 Temperature state determination means
43 Mode setting means
44 Valve stop mechanism control means
45 Intake air amount control means
46 Fuel control means
47 Ignition control means

Claims (9)

In a multi-cylinder spark-ignition four-cycle engine in which the combustion cycle of each cylinder is performed with a predetermined phase difference,
The control mode for the intake / exhaust and combustion states of the engine in the partial load range of the engine is a special operation mode. In this special operation mode, exhaust is performed from a preceding cylinder in the exhaust stroke between a pair of cylinders where the exhaust stroke and the intake stroke overlap. While the burned gas is directly introduced into the succeeding cylinder in the intake stroke through the inter-cylinder gas passage, and the gas discharged from the succeeding cylinder is guided to the exhaust passage, the two-cylinder connection state is established. In this case, combustion is performed at a lean air-fuel ratio whose air-fuel ratio is greater than the stoichiometric air-fuel ratio, and in the succeeding cylinders, fuel is supplied to burned gas having a lean air-fuel ratio introduced from the preceding cylinder to burn at substantially the stoichiometric air-fuel ratio. A control device that causes
Ignition control means for selecting whether to perform combustion in the subsequent cylinder by compression self-ignition or forced ignition by control in the special operation mode in accordance with the state of the engine,
At the time of control in the special operation mode, when the combustion in the subsequent cylinder is performed by compression self-ignition according to the selection by the ignition control means, the air-fuel ratio of the preceding cylinder is larger than the stoichiometric air-fuel ratio, and A control device for a spark ignition type four-stroke engine, comprising: fuel control means for controlling a fuel supply amount such that a value of the combustion in a succeeding cylinder is reduced as compared with a case in which combustion is performed by forced ignition. Along with a means for determining the temperature state of the engine, the ignition control means, during control in the special operation mode, selects between the compression self-ignition and the forced ignition according to the temperature state of the engine even in the same operation region. 2. The control device according to claim 1, wherein the control is performed. During the control in the special operation mode, the ignition control means performs the forced ignition when the engine temperature is lower than a predetermined temperature even after the engine is warmed up, and performs the compression self-ignition when the engine temperature is higher than the predetermined temperature. 3. The control device for a spark-ignition type four-stroke engine according to claim 2, wherein the control is performed as follows. In addition to the means for determining the octane number of the fuel, the fuel control means controls the compression of the combustion in the subsequent cylinder during the control in the high-speed, high-load side of the operation region in the special operation mode. 4. The fuel supply system according to claim 1, wherein, when the ignition is performed by self-ignition, the fuel supply amount is corrected such that the lower the octane number is, the smaller the air-fuel ratio of the preceding cylinder is, according to the determination of the octane number. A control device for a spark-ignition type four-stroke engine according to the above description. In addition to the means for determining the octane number of the fuel, the fuel control means controls the compression of the combustion in the subsequent cylinder during the control in the low speed, low load side region of the operation region in the special operation mode. 4. The fuel supply system according to claim 1, wherein, when the ignition is performed by self-ignition, the fuel supply amount is corrected such that the higher the octane number, the smaller the air-fuel ratio of the preceding cylinder becomes, according to the determination of the octane number. A control device for a spark-ignition type four-stroke engine according to the above description. The said ignition control means performs selection of the said compression self-ignition and the said forced ignition in the low speed area | region of the driving | operation area | region set as the said special operation mode, The Claim 1 characterized by the above-mentioned. Control system for spark-ignition four-stroke engine. The fuel control means may control the air-fuel ratio of the preceding cylinder to be a value larger than twice the stoichiometric air-fuel ratio when combustion in the subsequent cylinder is performed by forced ignition during control in the special operation mode. The control apparatus for a spark ignition type four-stroke engine according to any one of claims 1 to 6, wherein a fuel supply amount is controlled. The ignition control means, when performing the compression self-ignition during the control in the special operation mode, performs backup for the subsequent cylinder at a timing retarded by a predetermined amount from the ignition timing when the forced ignition is performed. The control device for a spark-ignition four-cycle engine according to any one of claims 1 to 7, wherein the ignition is performed. 9. The spark ignition system according to claim 8, wherein the ignition timing for backup in the case of the compression self-ignition is set after compression top dead center and near compression top dead center. Control device for cycle engine.

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