JP2004116403A - Control device for spark-ignition engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To effectively perform combustion by compression self-ignition in a wider operating region while suppressing the occurrence of knocking. <P>SOLUTION: The flow state of intake/exhaust is controlled so that burnt gas exhausted from preceding cylinders 2A, 2D is led into succeeding cylinders 2B, 2C in a partial load region of an engine. A combustion state control means 44 controls to perform combustion of the succeeding cylinders 2B, 2C by compression self-ignition in at least a partial operating region out of an operating region set to a special operation mode. In a high load side region in the compression self-ignition region, the combustion state control means 44 controls to lead fresh air into the succeeding cylinders 2B, 2C in addition to the burnt gas led out of the preceding cylinders 2A, 2D by making the air-fuel ratio of the preceding cylinders 2A, 2D relatively rich compared to a lower load side region and opening a fresh air lead-in intake valve 31a for leading the fresh air into the succeeding cylinders 2B, 2C. <P>COPYRIGHT: (C)2004,JPO

Description

【０００１】
【発明の属する技術分野】
本発明は、火花点火式エンジンの制御装置に関し、より詳しくは、多気筒のエンジンにおいて燃費改善及びエミッション向上のために各気筒の燃焼状態を制御する制御装置に関するものである。
【０００２】
【従来の技術】
従来から、火花点火式エンジンにおいて、各気筒内の混合気の空燃比を理論空燃比よりも大きいリーン空燃比とした状態で燃焼を行わせることにより燃費改善を図る技術が知られており、燃焼室内に直接燃料を噴射する燃料噴射弁を備え、低回転低負荷領域等で上記燃料噴射弁から圧縮行程で燃料を噴射して成層燃焼を行わせことにより、超リーン燃焼を実現するようにしたものが知られている（例えば、特許文献１参照）。
【０００３】
このようなエンジンにおいては、排気ガス浄化用の触媒として通常の三元触媒（ＨＣ，ＣＯ及びＮＯｘに対して理論空燃比付近で浄化性能の高い触媒）だけではリーン運転時にＮＯｘに対して充分な浄化性能が得られないため、下記特許文献１にも示されるように、酸素過剰雰囲気でＮＯｘを吸着して酸素濃度低下雰囲気でＮＯｘの離脱、還元を行うリーンＮＯｘ触媒を設けている。そして、このようなリーンＮＯｘ触媒を用いる場合、リーン運転中にリーンＮＯｘ触媒のＮＯｘ吸着量が増大したときは、例えば上記公報に示されるように主燃焼以外に膨張行程中に追加燃料を噴射することで排気ガスの空燃比をリッチ化するとともにＣＯを生成し、これによってＮＯｘの離脱、還元を促進するようにしている。
【０００４】
【特許文献１】
特開平１０−２９８３６号公報
【０００５】
【発明が解決しようとする課題】
上記のような従来のリーン運転を行うエンジンでは、リーン運転中のＮＯｘ浄化性能を確保するために、上記リーンＮＯｘ触媒を排気通路に設ける必要があり、コスト的に不利である。また、上記リーンＮＯｘ触媒の浄化性能を維持するためには、上述のようにＮＯｘ吸着量増大時にＮＯｘの離脱、還元のため追加燃料の供給等による一時的な空燃比のリッチ化を行う必要がある。さらに、使用燃料が硫黄分を多く含む場合には、上記リーンＮＯｘ触媒の硫黄被毒を解消するための触媒の加熱処理及び還元材供給等のリジェネレーション処理が必要となり、これらによって燃費改善効果が低下する。しかも、混合気の空燃比がある程度以上にリーンになると、燃焼速度が遅くなりすぎてその終期に近い燃焼が仕事に寄与しなくなるため、成層燃焼でのリーン化による燃費改善には限界があった。
【０００６】
また、燃費改善のための別の手法として、圧縮自己着火が研究されており、この圧縮自己着火は、ディーゼルエンジンと同様に圧縮行程終期に燃焼室内を高温・高圧にして燃料を自己着火させるようにするものであり、空燃比が超リーンの状態や多量のＥＧＲが導入されている状態でも、このような圧縮自己着火が行われれば燃焼室全体が一気に燃焼するため、仕事に寄与しない遅い燃焼が避けられて燃費改善に有利となる。
【０００７】
しかし、通常の火花点火式ガソリンエンジンでは、燃焼のために強制点火が必要であって、圧縮上死点付近での燃焼室内の温度、圧力が圧縮自己着火を生じさせる程度までには高められず、圧縮自己着火を行わせるには燃焼室内の温度または圧力を大幅に高めるための格別の工夫が必要となるが、従来、高負荷領域でのノッキング（燃焼室内で火炎が伝播する前に混合気が自然着火することによる異常燃焼）を避けつつ、燃費改善が要求される部分負荷領域で圧縮自己着火を生じさせる程度まで燃焼室内の温度または圧力を高めることが困難であった。
【０００８】
そこで、本出願人は、リーン燃焼と圧縮自己着火とを併用して大幅な燃費改善効果をもたせるべく、エンジンの部分負荷領域で、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に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にある後続気筒に気筒間ガス通路を介して導入される２気筒接続状態とするとともに、先行気筒では空燃比を理論空燃比よりも大きいリーン空燃比にして、強制点火により燃焼を行わせ、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圧縮自己着火により燃焼を行わせるようにした火花点火式エンジンの制御装置に関する技術を出願している（特願２００２−１８５２４２号）。
【０００９】
本発明は、このような技術に基づき、ノッキングの発生を抑制しつつ、さらに広い運転領域において後続気筒で効果的に圧縮自己着火による燃焼を行わせることができるようにし、燃費及びエミッションの改善効果を高めることができる火花点火式エンジンの制御装置を提供するものである。
【００１０】
【課題を解決するための手段】
請求項１に係る発明は、各気筒の燃焼サイクルが所定の位相差をもつように設定された多気筒の火花点火式エンジンにおいて、エンジンの部分負荷領域で、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に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に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ガスが排気通路に導かれるような２気筒接続状態としつつ、先行気筒の空燃比を理論空燃比よりも大きいリーン空燃比として燃焼を行わせ、この先行気筒から後続気筒に導入されたリーン空燃比の既燃ガスに燃料を供給して後続気筒の燃焼を行わせる特殊運転モードの制御を実行する運転モード制御手段とを備え、上記特殊運転モードとされる運転領域のうちの少なくとも一部の運転領域で、後続気筒を圧縮自己着火により燃焼を行わせるとともに、この圧縮自己着火領域における高負荷側領域では、それよりも低負荷側の領域に比べて先行気筒の空燃比を相対的にリッチとし、かつ後続気筒内に新気を導入する新気導入用吸気弁を開弁することにより、上記先行気筒から導出された既燃ガスに加えて新気を後続気筒内に導入させるように制御するものである。
【００１１】
この発明によると、エンジンの部分負荷領域で上記特殊運転モードとされるとともに後続気筒で圧縮自己着火により燃焼が行われる場合に、上記先行気筒ではリーン燃焼による熱効率向上及びポンピングロス低減による燃費改善効果が得られ、後続気筒では圧縮自己着火による燃焼効率の向上及びポンピングロス低減による燃費改善効果が得られる。そして、上記後続気筒の圧縮自己着火領域における高負荷側領域では、先行気筒の空燃比が比較的にリッチとされるとともに、これに対応して後続気筒内に導入される既燃ガス中の酸素濃度が低下した場合に、上記新気導入用吸気弁が開弁されて後続気筒に新気が導入されることにより、後続気筒の新気不足が解消されて上記後続気筒の圧縮自己着火が適正に実行され、かつ後続気筒内に導入される既燃ガス成分が増大されてノッキングの発生が効果的に防止されるとともに、エンジン出力が確保されることになる。
【００１２】
また、請求項２に係る発明は、上記請求項１記載の火花点火式エンジンの制御装置において、後続気筒の圧縮自己着火領域における低負荷側領域では、新気導入用吸気弁を閉弁状態に維持し、上記圧縮自己着火領域における高負荷側領域では、新気導入用吸気弁を後続気筒の吸気上死点付近で開弁するものである。
【００１３】
上記構成によれば、先行気筒の空燃比が比較的にリーンとされることにより、後続気筒内に導入される既燃ガス中の酸素濃度が充分に高い値に維持された後続気筒の圧縮自己着火領域における低負荷領域では、上記新気導入用吸気弁が閉弁状態に維持されることにより、後続気筒の空燃比がリーンな状態となることが防止される。また、上記圧縮自己着火領域における高負荷側領域では、新気導入用吸気弁が後続気筒の吸気上死点付近で開弁されることにより、後続気筒内に新気が効率よく導入されることになる。
【００１４】
また、請求項３に係る発明は、上記請求項２記載の火花点火式エンジンの制御装置において、後続気筒の圧縮自己着火領域における高負荷側領域で開弁した新気導入用吸気弁を、後続気筒の吸気行程途中で閉弁するものである。
【００１５】
上記構成によれば、上記圧縮自己着火領域における高負荷側領域では、新気導入用吸気弁が後続気筒の吸気上死点付近で開弁されることにより、後続気筒内に新気が効率よく導入された後、後続気筒の吸気行程途中で上記新気導入用吸気弁が閉弁されて新気の導入が停止されることにより、先行気筒から導出された既燃ガスが後続気筒内にスムーズに導入されることになる。
【００１６】
また、請求項４に係る発明は、上記請求項１記載の火花点火式エンジンの制御装置において、後続気筒の圧縮自己着火領域における高負荷側領域では、後続気筒の既燃ガス導入弁を吸気行程の途中で開弁し、かつこの既燃ガス導入弁の開弁時期よりも前に、新気導入用吸気弁を開弁するものである。
【００１７】
上記構成によれば、上記圧縮自己着火領域における高負荷側領域では、後続気筒の吸気行程途中で既燃ガス導入弁が開弁される前に、上記新気導入用吸気弁が閉弁されることにより、後続気筒内に新気が効率よく導入されるとともに、その後に後続気筒の吸気行程途中で上記既燃ガス導入弁が閉弁されることにより、先行気筒から導出された既燃ガスが後続気筒内に導入されることになる。
【００１８】
また、請求項５に係る発明は、上記請求項１記載の火花点火式エンジンの制御装置において、後続気筒の圧縮自己着火領域における高負荷側領域では、それよりも低負荷側の領域に比べ、先行気筒の空燃比がリッチになるのに対応して後続気筒に導入される総ガス量に対する新気導入量の割合を高めるように制御するものである。
【００１９】
上記構成によれば、後続気筒の圧縮自己着火領域における高負荷側領域では、先行気筒の空燃比が比較的にリッチとされるとともに、これに対応して後続気筒内に導入される既燃ガス中の酸素濃度が低下した場合に、後続気筒に導入される総ガス量に対する新気導入量の割合が高められることにより、後続気筒の新気不足が効果的に解消されて上記後続気筒の圧縮自己着火が適正に実行され、かつ後続気筒内の温度上昇が抑制されてノッキングの発生が効果的に防止されることになる。
【００２０】
また、請求項６に係る発明は、上記請求項１記載の火花点火式エンジンの制御装置において、少なくとも後続気筒の圧縮自己着火領域では、後続気筒から排出される排気ガス中の酸素濃度が、理論空燃比の燃焼状態に対応した値となるように後続気筒の空燃比を制御するものである。
【００２１】
上記構成によれば、少なくとも後続気筒の圧縮自己着火領域では、後続気筒から排出される排気ガス中の酸素濃度を、理論空燃比の燃焼状態に対応した値とする後続気筒の空燃比制御が実行されることにより、先行気筒でリーンな空燃比で燃焼が行われつつ、理論空燃比で燃焼した後続気筒の既燃ガスのみが排気通路に導出されることになる。
【００２２】
【発明の実施の形態】
図１は本発明の一実施形態によるエンジンの概略構成を示し、図２はエンジン本体１の一つの気筒とそれに対して設けられた吸・排気弁等の構造を概略的に示している。これらの図において、エンジン本体１は複数の気筒を有し、図示の実施形態では４つの気筒２Ａ〜２Ｄを有している。各気筒２Ａ〜２Ｄにはピストン３が嵌挿され、ピストン３の上方に燃焼室４が形成されている。
【００２３】
各気筒２Ａ〜２Ｄに設けられた燃焼室４の頂部には点火プラグ７が装備され、そのプラグ先端が燃焼室４内に臨んでいる。この点火プラグ７には、電子制御による点火時期のコントロールが可能な点火回路８が接続されている。
【００２４】
燃焼室４の側方部には、燃焼室４内に燃料を直接噴射する燃料噴射弁９が設けられている。この燃料噴射弁９は、図略のニードル弁及びソレノイドを内蔵し、後述のパルス信号が入力されることにより、そのパルス入力時期にパルス幅に対応する時間だけ駆動されて開弁し、その開弁時間に応じた量の燃料を噴射するように構成されている。なお、図外の燃料ポンプ及び燃料供給通路等を備えるとともに、圧縮行程での燃焼室内の圧力よりも高い燃料圧力を与え得る燃料供給系統を介して、上記燃料噴射弁９に燃料が供給されるように構成されている。
【００２５】
また、各気筒２Ａ〜２Ｄの燃焼室４に対して吸気ポート１１、１１ａ，１１ｂ及び排気ポート１２、１２ａ，１２ｂが開口し、これらのポートに吸気通路１５、排気通路２０等が接続されるとともに、各ポートが吸気弁３１、３１ａ，３１ｂ及び排気弁３２、３２ａ，３２ｂにより開閉されるようになっている。
【００２６】
そして、吸気、圧縮、膨張及び排気の各行程からなる燃焼サイクルが各気筒２Ａ〜２Ｄ毎に所定の位相差をもって行われるように構成され、４気筒エンジンの場合に、気筒列方向の一端側から１番気筒２Ａ、２番気筒２Ｂ、３番気筒２Ｃ及び４番気筒２Ｄと呼ぶと、図６に示すように、上記燃焼サイクルが１番気筒２Ａ、３番気筒２Ｃ、４番気筒２Ｄ、２番気筒２Ｂの順にクランク角で１８０°ずつの位相差をもって行われるようになっている。なお、図６において、ＥＸは排気行程、ＩＮは吸気行程であり、また、Ｆは燃料噴射、Ｓは強制点火を表し、図中の星マークは圧縮自己着火が行われることを表している。
【００２７】
排気行程と吸気行程が重なる一対の気筒間には、排気行程と吸気行程が重なるときの排気行程側の気筒（当明細書ではこれを先行気筒と呼ぶ）から吸気行程側の気筒（当明細書ではこれを後続気筒と呼ぶ）へ既燃ガスをそのまま導くことができるように、気筒間ガス通路２２が設けられている。当実施形態の４気筒エンジンでは、図６に示すように１番気筒２Ａの排気行程（ＥＸ）と２番気筒２Ｂの吸気行程（ＩＮ）とが重なり、また４番気筒２Ｄの排気行程（ＥＸ）と３番気筒２Ｃの吸気行程（ＩＮ）が重なるので、１番気筒２Ａ及び２番気筒２Ｂと、４番気筒２Ｄ及び３番気筒２Ｃとがそれぞれ一対をなし、１番気筒２Ａ及び４番気筒２Ｄが先行気筒となり、かつ２番気筒２Ｂ及び３番気筒２Ｃが後続気筒となる。
【００２８】
各気筒２Ａ〜２Ｄの吸・排気ポートと、これに接続される吸気通路１５、排気通路２０及び気筒間ガス通路２２は、具体的には次のように構成されている。
【００２９】
先行気筒である１番気筒２Ａ及び４番気筒２Ｄには、それぞれ、新気を導入するための吸気ポート１１と、既燃ガス（排気ガス）を排気通路２０に送り出すための第１排気ポート１２ａと、既燃ガスを後続気筒に導出するための第２排気ポート１２ｂとが配設されている。また、後続気筒である２番気筒２Ｂ及び３番気筒２Ｃには、それぞれ新気を導入するための第１吸気ポート１１ａと、先行気筒からの既燃ガスを導入するための第２吸気ポート１１ｂと、既燃ガスを排気通路に送り出すための排気ポート１２とが配設されている。
【００３０】
図１に示す例では、１番，４番気筒（先行気筒）２Ａ，２Ｄにおける吸気ポート１１及び２番，３番気筒（後続気筒）２Ｂ，２Ｃにおける第１吸気ポート１１ａが、１気筒当り２個ずつ、燃焼室の左半部側に並列的に設けられる一方、１番，４番気筒２Ａ，２Ｄ（先行気筒）における第１排気ポート１２ａ及び第２排気ポート１２ｂならびに２番，３番気筒（後続気筒）２Ｂ，２Ｃにおける第２吸気ポート１１ｂ及び排気ポート１２が、燃焼室の右半部側に並列的に設けられている。
【００３１】
１番，４番気筒２Ａ，２Ｄにおける吸気ポート１１及び２番，３番気筒２Ｂ，２Ｃにおける第１吸気ポート１１ａには、吸気通路１５における気筒別の分岐吸気通路１６の下流端が接続されている。各分岐吸気通路１６の下流端近傍には、共通の軸を介して互いに連動する多連スロットル弁１７が設けられており、この多連スロットル弁１７は制御信号に応じてアクチュエータ１８により駆動されることにより、吸入空気量を調節するようになっている。なお、吸気通路１５における集合部よりも上流の共通吸気通路には、吸気流量を検出するエアフローセンサ１９が設けられている。
【００３２】
１番，４番気筒２Ａ，２Ｄにおける第１排気ポート１２ａ及び２番，３番気筒２Ｂ，２Ｃにおける排気ポート１２には、排気通路２０における気筒別の分岐排気通路２１の上流端が接続されている。また、１番気筒２Ａと２番気筒２Ｂとの間及び３番気筒２Ｃと４番気筒２Ｄとの間には、それぞれ気筒間ガス通路２２が設けられている。そして、先行気筒である１番，４番気筒２Ａ，２Ｄの第２排気ポート１２ｂに、上記気筒間ガス通路２２の上流端が接続されるとともに、後続気筒である２番，３番気筒２Ｂ，２Ｃの第２吸気ポート１１ｂに、上記気筒間ガス通路２２の下流端が接続されている。
【００３３】
上記気筒間ガス通路２２は、互いに隣接する気筒間を接続する比較的短い通路であり、先行気筒２Ａ，２Ｄから排出されるガスがこの通路２２を通る間における放熱量が比較的小さく抑えられるようになっている。
【００３４】
排気通路２０における分岐排気通路２１の下流の集合部には排気ガス中の酸素濃度を検出することにより空燃比を検出するＯ_２センサ２３が設けられている。さらに、このＯ_２センサ２３の設置部の下流側における排気通路２０には、排気浄化用の三元触媒２４が設けられている。この三元触媒２４は、一般に知られているように、排気ガスの空燃比が理論空燃比（つまり空気過剰率λ＝１）付近にあるときにＨＣ，ＣＯ及びＮＯｘに対して高い浄化性能を示す触媒である。
【００３５】
各気筒２Ａ〜２Ｄの吸・排気ポートを開閉する吸・排気弁とこれらに対する動弁機構は、次のようになっている。
【００３６】
１番，４番気筒（先行気筒）２Ａ，２Ｄにおける吸気ポート１１、第１排気ポート１２ａ及び第２排気ポート１２ｂにはそれぞれ吸気弁３１、第１排気弁３２ａ及び第２排気弁３２ｂが設けられ、また、２番，３番気筒（後続気筒）２Ｂ，２Ｃにおける第１吸気ポート１１ａ、第２吸気ポート１１ｂ及び排気ポート１２にはそれぞれ第１吸気弁３１ａ、第２吸気弁３１ｂ及び排気弁３２が設けられている。そして、各気筒２Ａ〜２Ｄの吸気行程や排気行程が上述のような所定の位相差をもって行われるように、これら吸・排気弁がそれぞれカムシャフト３３，３４等からなる動弁機構により所定のタイミングで開閉するように駆動される。
【００３７】
さらに、これらの吸・排気弁のうちで第１排気弁３２ａ、第２排気弁３２ｂ、第１吸気弁３１ａ及び第２吸気弁３１ｂに対しては、各弁を作動状態と停止状態とに切換える弁停止機構３５が設けられている。この弁停止機構３５は、従来から知られているため詳しい図示は省略するが、例えば、カムシャフト３３，３４のカムと弁軸との間に介装されたタペットに作動油の給排が可能な油圧室が設けられ、この油圧室に作動油が供給されている状態ではカムの作動が弁に伝えられて弁が開閉作動され、油圧室から作動油が排出されたときにはカムの作動が弁に伝えられなくなることで弁が停止されるようになっている。
【００３８】
上記第１排気弁３２ａの弁停止機構３５と第１吸気弁３１ａの弁停止機構３５とに対する作動油給排用の通路３６には、第１コントロール弁３７が設けられ、また第２排気弁３２ｂの弁停止機構３５と第２吸気弁３１ｂの弁停止機構３５とに対する作動油給排用の通路３８には、第２コントロール弁３９が設けられている（図３参照）。
【００３９】
図３は、駆動、制御系統の構成を示している。この図において、マイクロコンピュータ等からなるエンジン制御用のＥＣＵ（コントロールユニット）４０には、エアフローセンサ１９及びＯ_２センサ２３からの信号が入力され、さらに運転状態を判別するためにエンジン回転数を検出する回転数センサ４７及びアクセル開度（アクセルペダル踏込み量）を検出するアクセル開度センサ４８等からの信号も入力されている。また、上記ＥＣＵ４０から、各燃料噴射弁９と、多連スロットル弁１７のアクチュエータ１８と、上記第１，第２のコントロール弁３９とに対して制御信号が出力されるようになっている。
【００４０】
上記ＥＣＵ４０は、運転状態判別手段４１、弁停止機構制御手段４２、吸入空気量制御手段４３及び燃焼状態制御手段４４を備えている。
【００４１】
運転状態判別手段４１は、図４に示すようにエンジンの運転領域が低負荷低回転側の運転領域Ａ（部分負荷領域）と、高負荷側ないし高回転側の運転領域Ｂとに分けられた制御用マップを有し、上記回転数センサ４７及びアクセル開度センサ４８等からの信号により調べられるエンジンの運転状態（エンジン回転数及びエンジン負荷）が上記運転領域Ａ，Ｂのいずれの領域にあるかを判別する。そして、この判別に基づき、低負荷低回転側の運転領域Ａでは、排気行程にある先行気筒２Ａ，２Ｄから排出される既燃ガスを、そのまま吸気行程にある後続気筒２Ｂ，２Ｃに導入して燃焼させる特殊運転モードが選択され、高負荷側ないし高回転側の運転領域Ｂでは、各気筒２Ａ〜２Ｄをそれぞれ独立させ燃焼させる通常運転モードが選択されるようになっている。
【００４２】
さらに運転状態判別手段４１は、上記特殊運転モードが選択される運転領域Ａにある場合に、この領域Ａのうちの高負荷側領域Ａ２、またはそれよりも低負荷側の領域Ａ１のいずれにあるかを判別するように構成されている。
【００４３】
弁停止機構制御手段４２は、特殊運転モードでは気筒間ガス通路２２を介して先行気筒の既燃ガスを後続気筒に導入させる２気筒接続状態とし、通常運転モードでは各気筒にそれぞれ新気を導入させる各気筒独立状態とするように吸・排気流通状態を変更すべく弁停止機構３５を制御するもので、具体的には運転状態が運転領域Ａ，Ｂのいずれにあるかに応じ、上記各コントロール弁３７，３９を制御することにより、原則として各弁停止機構３５を次のように制御する。
【００４４】

【００４５】
上記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理論空燃比に対応した値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。
【００４６】
上記燃焼状態制御手段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、特に運転状態が図４中の運転領域Ａにある場合と運転領域Ｂにある場合とで燃焼状態の制御（燃料噴射の制御及び点火の制御）が変更される。
【００４７】
すなわち、運転状態が低負荷低回転側の運転領域Ａにある場合、特殊運転モードでの制御状態として、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、実質的に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し、かつ、圧縮自己着火を行わせるべく、強制点火を停止させる。
【００４８】
さらに、上記運転領域Ａにおいて、先行気筒及び後続気筒からなる一対の気筒に対する燃料噴射量の和が先行気筒に導入される新気量に対して理論空燃比となる量に調整されるとともに、後続気筒でノッキングが発生するのを防止しつつ、圧縮自己着火が良好に行われるように、先行気筒（１番、４番気筒）２Ａ，２Ｄに対する燃料噴射量と、後続気筒（２番、３番気筒）２Ｂ，２Ｃに対する燃料噴射量との割合が運転状態に応じて制御される。
【００４９】
具体的には、上記運転領域Ａの低負荷側領域Ａ１では、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を略同一、ないしは後続気筒２Ｂ，２Ｃ側の燃料噴射量を少し多くすることにより、先行気筒２Ａ，２Ｄでの燃焼の際の空燃比が理論空燃比の２倍程度（Ａ／Ｆ≒３０、空気過剰率λで表せばλ＝２程度）、ないしは理論空燃比の２倍より大（空気過剰率λがλ＞２）となるようにする。この結果、エンジン負荷が低いために燃料の総噴射量が相対的に少ない値に設定され、これに起因して後続気筒２Ｂ，２Ｃの失火が発生し易い傾向にある上記低負荷側の領域Ａ１で、後続気筒２Ｂ，２Ｃに対する燃料噴射量が過度に少ない値に設定されることが防止され、上記失火の発生が防止されることになる。
【００５０】
これに対して上記運転領域Ａの高負荷側領域Ａ２では、先行気筒２Ａ，２Ｄに対する燃料噴射量を後続気筒２Ｂ，２Ｃに対する燃料噴射量よりも多くすることにより、先行気筒での燃焼の際の空燃比が理論空燃比の２倍より小（空気過剰率λが１＜λ＜２）となるようにし、例えばＡ／Ｆ≒２５となるように制御することにより、上記低負荷側の領域Ａ１に比べて先行気筒２Ａ，２Ｄの空燃比を相対的にリッチとする。この結果、エンジン負荷が高いために燃料の総噴射量が相対的に多い値に設定されることにより、後続気筒２Ｂ，２Ｃの温度が過度に高くなるとともに、これに対応して後続気筒２Ｂ，２Ｃでノッキングが発生し易い傾向にある上記高負荷側領域Ａ２で、後続気筒２Ｂ，２Ｃに多量の既燃ガスが導入され、そのＥＧＲ効果によって上記ノッキングの発生が防止されることになる。
【００５１】
また、上記のように運転領域Ａの高負荷側領域Ａ２で先行気筒２Ａ，２Ｄに対する燃料噴射量を後続気筒２Ｂ，２Ｃに対する燃料噴射量よりも多く設定した場合には、後続気筒２Ｂ，２Ｃに導入される既燃ガス中の酸素濃度が低下し、後続気筒２Ｂ，２Ｃに噴射された燃料を燃焼させることが不可能となって上記特殊運転モードの制御を実行することができなくなる懸念がある。このため、上記運転領域Ａの高負荷側領域Ａ２では、後続気筒２Ｂ，２Ｃ内に新気を導入する新気導入用吸気弁（第１吸気弁３１ａ）を一時的に開弁することにより、上記先行気筒２Ａ，２Ｄから導出された既燃ガスに加えて新気を後続気筒２Ｂ，２Ｃに導入させるように制御する。
【００５２】
すなわち、上記運転領域Ａの高負荷側領域Ａ２では、第１吸気弁３１ａを後続気筒２Ｂ，２Ｃの吸気上死点付近で開弁した後、この後続気筒２Ｂ，２Ｃの吸気行程途中で上記第１吸気弁３１ａを閉弁状態とする。また、この第１吸気弁３１ａが閉弁状態となる直前までは、後続気筒２Ｂ，２Ｃの既燃ガス導入弁（第２吸気弁３１ｂ）を閉弁状態に維持した後に、この既燃ガス導入弁を開弁することにより上記先行気筒２Ａ，２Ｄから導出された既燃ガスを後続気筒２Ｂ，２Ｃに導入させる。
【００５３】
一方、エンジンの運転状態が高負荷側ないし高回転側の運転領域Ｂにある場合には、通常運転モードでの制御として、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における大部分の領域で理論空燃比とし、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制御する。
【００５４】
以上のような当実施形態の装置の作用を、図５〜図８を参照しつつ説明する。低負荷低回転側の運転領域Ａでは、上記弁停止機構制御手段４２及び吸入空気量制御手段４３等からなる運転モード制御手段により、特殊運転モードの制御が実行され、原則として前述のように第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７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（図７中の矢印ｂ）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（図７中の矢印ｃ）ような２気筒接続状態とされる。
【００５５】
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７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くて、理論空燃比の略２倍ないしそれより小さい値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、リーン空燃比での成層燃焼が行われる（図６参照）。
【００５６】
また、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から導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中の白抜き矢印及び図７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、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の圧力、温度の上昇により圧縮自己着火が行われる。
【００５７】
この場合、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。
【００５８】
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から押出された既燃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。
【００５９】
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。
【００６０】
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。とくに、特殊運転モードでの先行気筒（１番、４番気筒２Ａ，２Ｄ）に対する燃料噴射量と、後続気筒（２番、３番気筒２Ｂ，２Ｃ）に対する燃料噴射量との割合が運転状態に応じて前述のように調整されることにより、広い運転領域に亘って適正に圧縮自己着火を行わせることができる。
【００６１】
すなわち、特殊運転モードとされる運転領域Ａのうちの高負荷側領域Ａ２では、低負荷側領域Ａ１に比べて先行気筒２Ａ，２Ｄに対する燃料噴射量が多く設定されることにより、先行気筒２Ａ，２Ｄの空燃比を相対的にリッチ、つまり理論空燃比の２倍より小さい値となるように制御される。これにより、先行気筒２Ａ，２Ｄの空燃比が理論空燃比の２倍（先行気筒と後続気筒とが同じ噴射量）とされる場合と比べ、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ことにより、そのＥＧＲ効果によってノッキングが抑制される。
【００６２】
そして、上記高負荷側領域Ａ２で先行気筒２Ａ，２Ｄの空燃比が理論空燃比の２倍より小さい値となるように設定されることにより、後続気筒２Ｂ，２Ｃに導入される既燃ガス中の新気量が減少することになるが、この場合に、上記先行気筒２Ａ，２Ｄから導出された既燃ガスに加えて新気を後続気筒２Ｂ，２Ｃ内に導入させるように構成したため、上記高負荷側領域Ａ２において後続気筒２Ｂ，２Ｃの新気不足が解消されて上記圧縮自己着火が適正に実行されることになる。
【００６３】
具体的には、図５に示すように、上記後続気筒２Ｂ，２Ｃの吸気上死点（ＩＴＤＣ）付近で、後続気筒２Ｂ，２Ｃの既燃ガス導入弁（第２吸気弁３１ｂ）を閉弁状態に維持しつつ、新気導入用吸気弁（第１吸気弁３１ａ）を開弁状態とすることにより、吸気通路１５及び分岐通路１６を介して導入された新気を上記後続気筒２Ｂ，２Ｃ内に供給するように構成したため、後続気筒２Ｂ，２Ｃにおいて圧縮自己着火を実行するための新気量を確保することができる。そして、後続気筒２Ｂ，２Ｃの吸気行程途中で上記第１吸気弁３１ａを閉弁状態とするとともに、その前に後続気筒２Ｂ，２Ｃの第２吸気弁３１ｂを開弁状態とすることにより、先行気筒２Ａ，２Ｄから導出された既燃ガスを後続気筒２Ｂ，２Ｃ内に導入させることができる。
【００６４】
上記のように圧縮自己着火領域Ａにおける高負荷側領域Ａ２で、新気導入用吸気弁（第１吸気弁３１ａ）を後続気筒２Ｂ，２Ｃの吸気上死点付近（ＩＴＤＣ）で開弁状態とすることにより、先行気筒２Ａ，２Ｄから導出された既燃ガスが後続気筒２Ｂ，２Ｃ内に導入される前に、後続気筒２Ｂ，２Ｃ内に比較的温度の低い新気を効率よく導入させることができる。しかも、後続気筒２Ｂ，２Ｃの圧縮自己着火領域（部分負荷領域）Ａにおける低負荷側領域Ａ１では、新気導入用吸気弁（第１吸気弁３１ａ）を閉弁状態に維持するように構成したため、先行気筒２Ａ，２Ｄの空燃比が比較的にリーンとされることにより、後続気筒２Ｂ，２Ｃ内に導入される既燃ガス中の酸素濃度が充分に高い値に維持された後続気筒２Ｂ，２Ｃの圧縮自己着火領域Ａにおける低負荷領域Ａ１において、後続気筒２Ｂ，２Ｃに新気が導入されることに起因して後続気筒２Ｂ，２Ｃの空燃比がリーンになるのを防止することができる。
【００６５】
また、後続気筒２Ｂ，２Ｃの圧縮自己着火領域Ａにおける高負荷側領域Ａ２で開弁した上記新気導入用吸気弁（第１吸気弁３１ａ）を、後続気筒２Ｂ，２Ｃの吸気行程途中で閉弁状態とすることにより、後続気筒２Ｂ，２Ｃ内に新気を効率よく導入させた後に、この新気の導入を停止させることにより、先行気筒２Ａ，２Ｃから導出された既燃ガスを後続気筒２Ｂ，２Ｃ内にスムーズに導入させることができる。
【００６６】
さらに、上記実施形態に示すように、後続気筒２Ｂ，２Ｃの圧縮自己着火領域Ａにおける高負荷側領域Ａ２では、後続気筒２Ｂ，２Ｃの既燃ガス導入弁（第２吸気弁３１ｂ）を吸気行程の途中で開弁し、かつこの既燃ガス導入弁（第２吸気弁３１ｂ）の開弁時期よりも前、例えば後続気筒２Ｂ，２Ｃの吸気上死点付近（ＩＴＤＣ）で、新気導入用吸気弁（第１吸気弁３１ａ）を開弁するように構成した場合には、上記圧縮自己着火領域Ａにおける高負荷側領域Ａ２で、後続気筒２Ｂ，２Ｃ内に新気を効率よく導入させることができるとともに、その後に上記新気導入用吸気弁（第１吸気弁３１ａ）を閉弁状態とすることにより、先行気筒２Ａ，２Ｄから導出された既燃ガスを後続気筒２Ｂ，２Ｃ内に効率よく導入させることができる。
【００６７】
すなわち、図５に破線で示すように、既燃ガス導入弁（第２吸気弁３１ｂ）を後続気筒２Ｂ，２Ｃの吸気上死点ＩＴＤＣ付近で開弁状態とすることもできるが、この場合には、吸気通路１５から供給された新気と、気筒間ガス通路２２を介して供給された既燃ガスとが同時に後続気筒２Ｂ，２Ｃ内に導入されることにより、新気の導入量が低減されることになる。このため、上記のように後続気筒２Ｂ，２Ｃの吸気行程途中まで既燃ガス導入弁（第２吸気弁３１ｂ）を閉止状態に維持することにより、後続気筒２Ｂ，２Ｃ内に新気が効率よく導入されるように構成することが好ましい。また、後続気筒２Ｂ，２Ｃの吸気行程途中まで既燃ガス導入弁を閉止状態に維持するように構成した場合には、先行気筒２Ａ，２Ｄにおける内部ＥＧＲ量を増大させることができるため、先行気筒２Ａ，２Ｄの内部温度を上昇させて圧縮自己着火させることができるという利点がある。
【００６８】
また、後続気筒２Ｂ，２Ｃの圧縮自己着火領域Ａにおける高負荷側領域Ａ２では、それよりも低負荷側の領域Ａ１に比べ、先行気筒２Ａ，２Ｄの空燃比がリッチになるのに対応して後続気筒２Ｂ，２Ｃ内に導入される総ガス量に対する新気導入量の割合を高める制御を実行するように構成した場合には、後続気筒２Ｂ，２Ｃの圧縮自己着火領域Ａにおける高負荷側領域Ａ２で、先行気筒２Ａ，２Ｄの空燃比が比較的にリッチに設定されるのに対応して後続気筒２Ｂ，２Ｃ内に導入される既燃ガス中の酸素濃度が低下した場合に、後続気筒２Ｂ，２Ｃに導入される総ガス量に対する新気導入量の割合を高めることにより、後続気筒２Ｂ，２Ｃの新気不足を効果的に解消して上記後続気筒２Ｂ，２Ｃの圧縮自己着火を適正に実行しつつ、エンジン出力を充分に確保することができ、かつ後続気筒２Ｂ，２Ｃ内の温度上昇を抑制してノッキングの発生を効果的に防止できるという利点がある。
【００６９】
また、少なくとも後続気筒２Ｂ，２Ｃの圧縮自己着火領域Ａで、後続気筒２Ｂ，２Ｃから排出される排気ガス中の酸素濃度が、理論空燃比の燃焼状態に対応した値となるように後続気筒２Ｂ，２Ｃの空燃比を制御するように構成した場合には、先行気筒２Ａ，２Ｄでリーンな空燃比で燃焼が行われつつ、理論空燃比で燃焼した後続気筒２Ｂ，２Ｃの既燃ガスのみが排気通路２０に導出されることになる。したがって、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。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。
【００７０】
一方、高負荷側ないし高回転側の運転領域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８に示すようになり、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１２，１２ａに新気が導入されるとともに各気筒２Ａ〜２Ｄの排気ポート１２，１２ａ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。
【００７１】
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。すなわち、上記各実施形態では、特殊運転モードとされる運転領域Ａの全域で、後続気筒２Ｂ，２Ｃを圧縮自己着火により燃焼させるようにしているが、特殊運転モードとされる運転領域Ａのうちの一部、例えば燃焼室内の温度、圧力が圧縮自己着火可能な状態に達しにくい極低速低負荷の領域では、後続気筒２Ｂ，２Ｃに対して所定の点火時期に点火プラグ７による点火を行わせ、強制点火により燃焼させるようにしてもよい。あるいはまた、エンジン温度が低いときに、後続気筒２Ｂ，２Ｃを強制点火により燃焼させるようにしてもよい。
【００７２】
また、上記基本実施形態では弁停止機構３５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。
【００７３】
さらに、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。
【００７４】
【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両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、圧縮自己着火により燃焼を行わせるようにしているため、先行気筒ではリーン燃焼による熱効率向上及びポンピングロス低減により、また後続気筒では圧縮自己着火による燃焼効率の向上及びポンピングロス低減により、燃費を改善することができる。
【００７５】
そして、特に本発明では、上記後続気筒の圧縮自己着火領域における高負荷側領域では、先行気筒の空燃比が比較的にリッチとされることによりエンジン出力が確保されるとともに、これに対応して後続気筒内に導入される既燃ガス中の酸素濃度が低下した場合に、上記新気導入用吸気弁が開弁されて後続気筒に新気が導入されるように構成されているため、後続気筒の新気不足を解消して上記後続気筒の圧縮自己着火を適正に実行し、かつ後続気筒内に導入される既燃ガス成分を増大させてノッキングの発生を効果的に防止することができる。このため、上記圧縮自己着火領域を大幅に拡大することができる。
【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】先行気筒及び後続気筒の燃焼サイクル及び開弁タイミング等を示す説明図である。
【図６】各気筒の排気行程、吸気行程、燃料噴射時期及び点火時期等を示す図である。
【図７】低負荷低回転時の実質的な新気及びガスの流通経路を示す説明図である。
【図８】高負荷高回転時の実質的な新気及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
９　燃料噴射弁
１５　吸気通路
２０　排気通路
２２　気筒間ガス通路
３５　弁停止機構
３１ａ　第１吸気弁（新気導入用吸気弁）
３１ｂ　第２吸気弁（既燃ガス導入弁）
４０　ＥＣＵ
４１　運転状態判別手段
４２　弁停止機構制御手段（運転モード制御手段）
４３　吸入空気量制御手段
４４　燃焼状態制御手段[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a control device for a spark ignition type engine, and more particularly, to a control device for controlling a combustion state of each cylinder in a multi-cylinder engine for improving fuel efficiency and emission.
[0002]
[Prior art]
Conventionally, in a spark ignition type engine, a technology for improving fuel efficiency by performing combustion in a state where 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 has been known. A fuel injection valve that injects fuel directly into the room is provided, and super-lean combustion is realized by injecting fuel in the compression stroke from the fuel injection valve and performing stratified combustion in a low-speed low-load region or the like. Some are known (for example, see Patent Document 1).
[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 no purification performance can 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 Patent Document 1 below. 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]
[Patent Document 1]
JP-A-10-29836
[0005]
[Problems to be solved by the invention]
In the engine performing the conventional lean operation as described above, it is necessary to provide the lean NOx catalyst in the exhaust passage in order 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. is there. Further, when the used fuel contains a large amount of sulfur, it is necessary to perform a heating treatment of the catalyst to eliminate the sulfur poisoning of the lean NOx catalyst and a regenerative treatment such as a supply of a reducing material. descend. In addition, when the air-fuel ratio of the air-fuel mixture becomes lean to a certain degree or more, the combustion speed becomes too slow, and the combustion near the end does not contribute to the work, so there is a limit to the improvement of fuel efficiency by leaning in stratified combustion. .
[0006]
As another technique 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, as in diesel engines, to self-ignite the fuel. 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 burns at a stretch, so that slow combustion that does not contribute to work is performed. Is avoided, which is advantageous for improving fuel efficiency.
[0007]
However, in a normal spark ignition type gasoline engine, forced ignition is required for combustion, and the temperature and pressure in the combustion chamber near the compression top dead center cannot be increased to the extent that compression self-ignition occurs. In order to perform compression self-ignition, it is necessary to take special measures to significantly increase the temperature or pressure in the combustion chamber. Conventionally, however, knocking in a high-load region (mixing air before the flame propagates in the combustion chamber) However, 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 efficiency is required, while avoiding abnormal combustion caused by spontaneous ignition).
[0008]
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 have a significant fuel efficiency improvement effect by using both the lean combustion and the compression self-ignition. 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. 2002-185242).
[0009]
The present invention makes it possible to effectively perform combustion by compression self-ignition in a subsequent cylinder in a wider operating range while suppressing knocking based on such technology, thereby improving fuel efficiency and emission. It is intended to provide a control device for a spark-ignition engine capable of increasing the engine speed.
[0010]
[Means for Solving the Problems]
The invention according to claim 1 is a multi-cylinder spark ignition type engine in which the combustion cycle of each cylinder is set to have a predetermined phase difference, wherein a pair of exhaust strokes and intake strokes overlap in a partial load region of the engine. The burned gas discharged from the preceding cylinder in the exhaust stroke between the cylinders is directly introduced into the subsequent cylinder in the intake stroke via the inter-cylinder gas passage, and the gas discharged from this succeeding cylinder is guided to the exhaust passage. The combustion is performed with the air-fuel ratio of the preceding cylinder set to a lean air-fuel ratio larger than the stoichiometric air-fuel ratio while the two-cylinder connection state is established, and fuel is supplied from this preceding cylinder to burned gas having a lean air-fuel ratio introduced into the succeeding cylinder. Operation mode control means for performing control of a special operation mode for causing the subsequent cylinder to burn, and operating at least a part of the operation region in the special operation mode. In the compression self-ignition region, the air-fuel ratio of the preceding cylinder is made relatively richer in the high load side region than in the lower load side region. By opening a fresh air introduction intake valve for introducing fresh air into the succeeding cylinder, control is performed such that fresh air is introduced into the succeeding cylinder in addition to the burned gas derived from the preceding cylinder. Things.
[0011]
According to the present invention, when the special operation mode is set in the partial load region of the engine and combustion is performed by compression self-ignition in the subsequent cylinder, the preceding cylinder has improved thermal efficiency by lean combustion and improved fuel efficiency by reducing pumping loss. In the subsequent cylinder, the combustion efficiency is improved by the compression self-ignition and the fuel efficiency is improved by reducing the pumping loss. In the high load side region in the compression self-ignition region of the succeeding cylinder, the air-fuel ratio of the preceding cylinder is made relatively rich, and the oxygen in the burned gas introduced into the succeeding cylinder is correspondingly increased. When the concentration is reduced, the intake valve for fresh air is opened to introduce fresh air to the succeeding cylinder, whereby the lack of fresh air in the succeeding cylinder is resolved and the compression self-ignition of the succeeding cylinder is properly performed. And the burned gas component introduced into the succeeding cylinder is increased to effectively prevent the occurrence of knocking and ensure the engine output.
[0012]
According to a second aspect of the present invention, in the control apparatus for a spark ignition engine according to the first aspect, the intake valve for introducing fresh air is closed in a low load side region in a compression self-ignition region of a subsequent cylinder. In the high-load region in the compression self-ignition region, the intake valve for introducing fresh air is opened near the intake top dead center of the subsequent cylinder.
[0013]
According to the above configuration, since the air-fuel ratio of the preceding cylinder is relatively lean, the compression self-compression of the succeeding cylinder in which the oxygen concentration in the burned gas introduced into the succeeding cylinder is maintained at a sufficiently high value. In the low load range in the ignition range, the air-fuel ratio of the subsequent cylinder is prevented from becoming lean by maintaining the intake valve for fresh air introduction in the closed state. Further, in the high load side region in the compression self-ignition region, the fresh air introduction intake valve is opened near the intake top dead center of the subsequent cylinder, so that fresh air is efficiently introduced into the subsequent cylinder. become.
[0014]
According to a third aspect of the present invention, in the control apparatus for a spark ignition engine according to the second aspect, the intake valve for introducing fresh air, which is opened in the high load side region in the compression auto-ignition region of the subsequent cylinder, The valve is closed during the intake stroke of the cylinder.
[0015]
According to the above configuration, in the high-load side region in the compression self-ignition region, the fresh air introduction intake valve is opened near the intake top dead center of the subsequent cylinder, so that fresh air is efficiently introduced into the subsequent cylinder. After being introduced, the intake valve for fresh air is closed during the intake stroke of the succeeding cylinder to stop the introduction of fresh air, so that the burned gas derived from the preceding cylinder smoothly flows into the succeeding cylinder. Will be introduced.
[0016]
According to a fourth aspect of the present invention, in the control apparatus for a spark ignition type engine according to the first aspect, the burned gas introduction valve of the subsequent cylinder is moved to the intake stroke in the high load side region in the compression self-ignition region of the subsequent cylinder. The intake valve for fresh air introduction is opened before the timing of opening the burned gas introduction valve.
[0017]
According to the above configuration, in the high load side region in the compression self-ignition region, the fresh air introduction intake valve is closed before the burned gas introduction valve is opened during the intake stroke of the subsequent cylinder. As a result, fresh air is efficiently introduced into the succeeding cylinder, and the burned gas introduction valve is closed during the intake stroke of the succeeding cylinder. It will be introduced into the following cylinder.
[0018]
Further, according to a fifth aspect of the present invention, in the control apparatus for a spark ignition engine according to the first aspect, the high load side region in the compression self-ignition region of the subsequent cylinder is smaller than the lower load side region. The control is performed to increase the ratio of the fresh air introduction amount to the total gas amount introduced into the succeeding cylinder in response to the richness of the air-fuel ratio of the preceding cylinder.
[0019]
According to the above configuration, in the high load side region in the compression self-ignition region of the subsequent cylinder, the air-fuel ratio of the preceding cylinder is made relatively rich, and the burned gas introduced into the subsequent cylinder correspondingly When the oxygen concentration in the inside decreases, the ratio of the amount of fresh air introduced to the total amount of gas introduced into the succeeding cylinder is increased, whereby the shortage of fresh air in the succeeding cylinder is effectively eliminated, and the compression of the succeeding cylinder is reduced. The self-ignition is properly executed, and the rise in the temperature in the following cylinder is suppressed, so that the occurrence of knocking is effectively prevented.
[0020]
According to a sixth aspect of the present invention, in the control apparatus for a spark ignition engine according to the first aspect, at least in a compression self-ignition region of the subsequent cylinder, the oxygen concentration in the exhaust gas discharged from the subsequent cylinder is theoretically reduced. The air-fuel ratio of the following cylinder is controlled so that the air-fuel ratio becomes a value corresponding to the combustion state.
[0021]
According to the above configuration, at least in the compression self-ignition region of the subsequent cylinder, the air-fuel ratio control of the subsequent cylinder is performed so that the oxygen concentration in the exhaust gas discharged from the subsequent cylinder is a value corresponding to the combustion state of the stoichiometric air-fuel ratio. As a result, only the burned gas of the succeeding cylinder burned at the stoichiometric air-fuel ratio is led to the exhaust passage, while the preceding cylinder burns at the lean air-fuel ratio.
[0022]
BEST MODE FOR CARRYING OUT THE INVENTION
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.
[0023]
A spark plug 7 is provided at the top of the combustion chamber 4 provided in each of the cylinders 2A to 2D, and the tip of the plug faces the inside of the combustion chamber 4. The ignition plug 7 is connected to an ignition circuit 8 capable of controlling the ignition timing by electronic control.
[0024]
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. Note that fuel is supplied to the fuel injection valve 9 through a fuel supply system that includes a fuel pump and a fuel supply passage (not shown) and that can provide a fuel pressure higher than the pressure in the combustion chamber in the compression stroke. It is configured as follows.
[0025]
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.
[0026]
The combustion cycle including the intake, compression, expansion, and exhaust strokes is performed with a predetermined phase difference for each of the cylinders 2A to 2D. In the case of a four-cylinder engine, the combustion cycle starts from one end in the cylinder row direction. When called the first cylinder 2A, the second cylinder 2B, the third cylinder 2C, and the fourth cylinder 2D, as shown in FIG. 6, the combustion cycle is the first cylinder 2A, the third cylinder 2C, the fourth cylinder 2D, This is performed with a phase difference of 180 ° at each crank angle in the order of the second cylinder 2B. In FIG. 6, EX indicates an exhaust stroke, IN indicates an intake stroke, F indicates fuel injection, S indicates forced ignition, and the star mark in the figure indicates that compression self-ignition is performed.
[0027]
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 this embodiment, as shown in FIG. 6, 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 make a pair, respectively, and the first cylinder 2A and the fourth cylinder 2C. The cylinder 2D is a leading cylinder, and the second cylinder 2B and the third cylinder 2C are succeeding cylinders.
[0028]
The intake / exhaust ports of the cylinders 2A to 2D and the intake passage 15, exhaust passage 20 and inter-cylinder gas passage 22 connected thereto are specifically configured as follows.
[0029]
The first cylinder 2A and the fourth cylinder 2D, which are the leading cylinders, have an intake port 11 for introducing fresh air and a first exhaust port 12a for sending burned gas (exhaust gas) to the exhaust passage 20, 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 11b for introducing burned gas from the preceding cylinder are provided to the second cylinder 2B and the third cylinder 2C, which are subsequent cylinders, respectively. And an exhaust port 12 for sending burned gas to an exhaust passage.
[0030]
In the example shown in FIG. 1, the intake ports 11 in the first and fourth cylinders (preceding cylinders) 2A and 2D and the first intake ports 11a in the second and third cylinders (subsequent cylinders) 2B and 2C have two The first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders 2A and 2D (preceding cylinders) and the second and third cylinders are provided in parallel on the left half side of the combustion chamber. (Following cylinders) The second intake port 11b and the exhaust port 12 in 2B and 2C are provided in parallel on the right half side of the combustion chamber.
[0031]
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. Near the downstream end of each branch intake passage 16, there are provided multiple throttle valves 17 which are interlocked with each other via a common shaft, and the multiple throttle valves 17 are driven by actuators 18 according to control signals. Thus, the intake air amount is adjusted. Note that an airflow 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.
[0032]
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. 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. The upstream end of the inter-cylinder gas passage 22 is connected to the second exhaust ports 12b of the first and fourth cylinders 2A and 2D, which are the preceding cylinders, and the second and third cylinders 2B and 2B which are the following cylinders. The downstream end of the inter-cylinder gas passage 22 is connected to the 2C second intake port 11b.
[0033]
The inter-cylinder gas passage 22 is a relatively short passage connecting between adjacent cylinders, and the amount of heat released while the gas discharged from the preceding cylinders 2A, 2D passes through the passage 22 is relatively small. It has become.
[0034]
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. Furthermore, this O ₂ A three-way catalyst 24 for purifying exhaust gas is provided in the exhaust passage 20 on the downstream side of the installation part of the sensor 23. As is generally known, the three-way catalyst 24 has high purification performance for HC, CO and NOx when the air-fuel ratio of the exhaust gas is near the stoichiometric air-fuel ratio (that is, the excess air ratio λ = 1). It is a catalyst shown.
[0035]
The intake / exhaust valves for opening and closing the intake / exhaust ports of the cylinders 2A to 2D and the valve mechanism for these valves are as follows.
[0036]
The intake port 11, the first exhaust port 12a and the second exhaust port 12b of the first and fourth cylinders (preceding cylinders) 2A and 2D are provided with an intake valve 31, a first exhaust valve 32a and a second exhaust valve 32b, respectively. The first intake port 11a, the second intake port 11b, and the exhaust port 12 of the second and third cylinders (subsequent cylinders) 2B, 2C are respectively provided with a first intake valve 31a, a second intake valve 31b, and an exhaust valve 32. Is provided. The intake and exhaust valves are operated at predetermined timings by a valve mechanism including camshafts 33 and 34 so that the intake and exhaust strokes of the cylinders 2A to 2D are performed with the above-described predetermined phase difference. It is driven to open and close.
[0037]
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.
[0038]
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 the second exhaust valve 32b A second control valve 39 is provided in a passage 38 for supplying and discharging hydraulic oil to the valve stop mechanism 35 of the second intake valve 31b and the valve stop mechanism 35 of the second intake valve 31b (see FIG. 3).
[0039]
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 the sensor 23 is input, and signals from a rotation speed sensor 47 for detecting an engine rotation speed and an accelerator opening sensor 48 for detecting an accelerator opening (accelerator pedal depression amount) for determining an operating state are also provided. Has been entered. The ECU 40 outputs control signals 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.
[0040]
The ECU 40 includes an operation state determination unit 41, a valve stop mechanism control unit 42, an intake air amount control unit 43, and a combustion state control unit 44.
[0041]
As shown in FIG. 4, the operating state determining means 41 divides the operating region of the engine into an operating region A (partial load region) on the low-load low-rotation side and an operating region B on the high-load side or the high-rotation side. The operating state (engine speed and engine load) of the engine, which has a control map and is checked by signals from the rotation speed sensor 47, the accelerator opening sensor 48, and the like, is in any of the operation regions A and B. Is determined. Then, based on this determination, in the low-load, low-rotation-side operation region A, the burned gas discharged from the preceding cylinders 2A, 2D in the exhaust stroke is directly introduced into the subsequent cylinders 2B, 2C in the intake stroke. The special operation mode in which combustion is performed is selected, and in the operation region B on the high load side or the high rotation side, the normal operation mode in which each of the cylinders 2A to 2D is independently burned is selected.
[0042]
Further, when the operating state determining means 41 is in the operating area A in which the special operation mode is selected, the operating state determining means 41 is in either the high load side area A2 of this area A or the lower load side area A1. Is determined.
[0043]
In the special operation mode, the valve stop mechanism control means 42 is in a two-cylinder connection state in which the burned gas of the preceding cylinder is introduced into the succeeding cylinder via the inter-cylinder gas passage 22, and in the normal operation mode, fresh air is introduced into each cylinder. The valve stop mechanism 35 is controlled so as to change the intake / exhaust flow state so as to make each cylinder independent state. Specifically, according to which of the operation states A and B the operation state is, By controlling the control valves 37 and 39, each valve stop mechanism 35 is controlled in principle as follows.
[0044]

[0045]
The intake air amount control means 43 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 in accordance with an operating 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 a value corresponding to the stoichiometric air-fuel ratio, the combustion required for the combustion of the fuel according to the required torque of the preceding and succeeding two cylinders is performed. The throttle opening is adjusted so that the amount of air (the amount of air that becomes the stoichiometric air-fuel ratio with respect to the amount of fuel for the two cylinders) is supplied to the preceding cylinders (the first and fourth cylinders 2A and 2D). You.
[0046]
The combustion state control means 44 includes a fuel injection control means 45 and an ignition control means 46. The fuel injection control means 45 controls the amount of fuel injection from the fuel injection valves 9 provided in the cylinders 2A to 2D. The injection timing is controlled according to the operation state of the engine, and the ignition control means 46 controls the ignition timing and the ignition stop according to the operation state. In particular, the control of the combustion state (the control of the fuel injection and the control of the ignition) is changed between the case where the operation state is in the operation area A and the case where the operation state is in the operation area B in FIG.
[0047]
That is, when the operation state is in the operation area A on the low-load, low-rotation side, the stoichiometric air-fuel ratio is set as the control state in the special operation mode for the preceding cylinders (the first and fourth cylinders) 2A and 2D. The fuel injection amount is controlled so as to have a lean air-fuel ratio larger than the fuel ratio, the injection timing is set so as to inject fuel in the compression stroke to perform stratification of the air-fuel mixture, and the vicinity of the compression top dead center is set. The ignition timing is set so as to perform forced ignition. On the other hand, fuel is supplied to the subsequent cylinders (second and third cylinders) 2B and 2C with respect to the burned gas having the lean air-fuel ratio introduced from the preceding cylinder so that the stoichiometric air-fuel ratio is substantially achieved. In addition to controlling the fuel injection amount, the injection timing is set so as to inject the fuel in the intake stroke, and the forced ignition is stopped to perform the compression self-ignition.
[0048]
Further, in the operation region A, the sum of the fuel injection amounts for the pair of cylinders including the leading cylinder and the trailing cylinder is adjusted to an amount that becomes the stoichiometric air-fuel ratio with respect to the fresh air amount introduced to the leading cylinder, and In order to prevent the occurrence of knocking in the cylinders and perform the compression auto-ignition satisfactorily, the fuel injection amounts to the preceding cylinders (the first and fourth cylinders) 2A and 2D and the subsequent cylinders (the second and third cylinders) The ratio of the fuel injection amount to the (cylinder) 2B, 2C is controlled according to the operating state.
[0049]
Specifically, in the low-load-side region A1 of the operation region A, the fuel injection amount for the preceding cylinders 2A and 2D and the fuel injection amount for the succeeding cylinders 2B and 2C are substantially the same, or the fuel injection amount for the succeeding cylinders 2B and 2C is By slightly increasing the injection amount, the air-fuel ratio at the time of combustion in the preceding cylinders 2A and 2D is about twice the stoichiometric air-fuel ratio (A / F ≒ 30, about λ = 2 when expressed by the excess air ratio λ), Or, it is set to be larger than twice the stoichiometric air-fuel ratio (excess air ratio λ> 2). As a result, the total fuel injection amount is set to a relatively small value because the engine load is low, and as a result, the low-load side area A1 in which misfires in the subsequent cylinders 2B and 2C tend to occur easily. Therefore, the fuel injection amount for the subsequent cylinders 2B and 2C is prevented from being set to an excessively small value, and the occurrence of the misfire is prevented.
[0050]
On the other hand, in the high load side region A2 of the operation region A, the fuel injection amount for the preceding cylinders 2A and 2D is made larger than the fuel injection amount for the succeeding cylinders 2B and 2C, so that the combustion in the preceding cylinder is performed. By controlling the air-fuel ratio to be smaller than twice the stoichiometric air-fuel ratio (the excess air ratio λ is 1 <λ <2), for example, to control A / F ≒ 25, the low-load side region A1 is controlled. , The air-fuel ratio of the preceding cylinders 2A and 2D is made relatively rich. As a result, the temperature of the subsequent cylinders 2B, 2C becomes excessively high by setting the total fuel injection amount to a relatively large value due to the high engine load, and the subsequent cylinders 2B, 2B, In the high load side area A2 where knocking tends to occur in 2C, a large amount of burned gas is introduced into the subsequent cylinders 2B and 2C, and the occurrence of the knocking is prevented by the EGR effect.
[0051]
Further, as described above, when the fuel injection amount for the preceding cylinders 2A, 2D is set to be larger than the fuel injection amount for the following cylinders 2B, 2C in the high load side region A2 of the operation region A, the following cylinders 2B, 2C are set. There is a concern that the oxygen concentration in the burned gas introduced decreases, the fuel injected into the subsequent cylinders 2B, 2C cannot be burned, and the control in the special operation mode cannot be executed. . For this reason, in the high load side region A2 of the operation region A, a fresh air introduction intake valve (first intake valve 31a) for introducing fresh air into the subsequent cylinders 2B and 2C is temporarily opened, Control is performed such that fresh air is introduced into the succeeding cylinders 2B and 2C in addition to the burned gas derived from the preceding cylinders 2A and 2D.
[0052]
That is, in the high load side region A2 of the operation region A, after the first intake valve 31a is opened near the intake top dead center of the subsequent cylinders 2B and 2C, the first intake valve 31a is opened during the intake stroke of the subsequent cylinders 2B and 2C. One intake valve 31a is closed. Until immediately before the first intake valve 31a is closed, the burned gas introduction valves (second intake valves 31b) of the subsequent cylinders 2B and 2C are maintained in a closed state. By opening the valve, the burned gas derived from the preceding cylinders 2A, 2D is introduced into the succeeding cylinders 2B, 2C.
[0053]
On the other hand, when the operation state of the engine is in the operation range B on the high load side or the high rotation side, as the control in the normal operation mode, the air-fuel ratio of each of the cylinders 2A to 2D is set to the stoichiometric air-fuel ratio or lower. For example, the stoichiometric air-fuel ratio is controlled in most of the operation region B, and the stoichiometric air-fuel ratio is set to be richer than the stoichiometric air-fuel ratio in the fully open load and the operation 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 during the intake stroke so as to equalize the air-fuel mixture, and the cylinders 2A to 2D are also forcedly ignited. To control.
[0054]
The operation of the apparatus of the present embodiment as described above will be described with reference to FIGS. In the operation region A on the low-load, low-rotation side, the special operation mode is controlled by the operation mode control means including the valve stop mechanism control means 42 and the intake air amount control means 43. When the first exhaust valve 32a and the first intake valve 31a are in the stopped state and the second exhaust valve 32b and the second intake valve 31b are in the activated state, the substantial fresh air and gas flow paths are as shown in FIG. The burned gas discharged from the preceding cylinders (No. 1 and No. 4 cylinders) 2A and 2D is directly introduced into the succeeding cylinders (No. 2 and No. 3 cylinders) 2B and 2C via the inter-cylinder gas passage 22. Along with (the arrow b in FIG. 7), only the gas exhausted from the subsequent cylinders 2B and 2C is led to the exhaust passage 20 (the arrow c in FIG. 7) so that the two cylinders are connected.
[0055]
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. 7), and the air-fuel ratio of the preceding cylinders 2A and 2D is larger than the stoichiometric air-fuel ratio. Fuel is injected in the compression stroke while controlling the fuel injection amount so as to be approximately twice or less than the air-fuel ratio, and ignition is performed at a predetermined ignition timing, so that stratified combustion at a lean air-fuel ratio is performed. (See FIG. 6).
[0056]
Further, during a period in which the intake strokes of the preceding cylinders 2A, 2D and the exhaust strokes of the succeeding cylinders 2B, 2C overlap, the burned gas derived from the preceding cylinders 2A, 2D is introduced into the following cylinders 2B, 2C through the gas passage 22. (The white arrow in FIG. 6 and the arrow b in FIG. 7). 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 the fuel injection amount is controlled so that the stoichiometric air-fuel ratio is obtained. After the fuel is injected, the compression self-ignition is performed by increasing the pressure and temperature in the combustion chamber near the top dead center of the compression stroke.
[0057]
In this case, the high-temperature burned gas discharged from the preceding cylinders 2A and 2D is immediately introduced into the succeeding cylinders 2B and 2C through the short inter-cylinder gas passage 22, so that the succeeding cylinders 2B and 2C perform the intake stroke in the combustion chamber. When the pressure and the temperature further rise 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.
[0058]
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 air-fuel mixture distribution state, so that the thermal efficiency is enhanced, and the burned fuel extruded from the preceding cylinders 2A and 2D Since the gas is supplied, the pumping loss is further reduced as compared with the preceding cylinders 2A and 2D. These actions greatly improve fuel economy.
[0059]
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.
[0060]
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. In particular, the ratio of the fuel injection amount to the preceding cylinder (the first and fourth cylinders 2A and 2D) and the fuel injection amount to the succeeding cylinders (the second and third cylinders 2B and 2C) in the special operation mode becomes the operating state. By performing adjustment as described above, compression self-ignition can be appropriately performed over a wide operating range.
[0061]
That is, in the high load side region A2 of the operation region A which is set to the special operation mode, the fuel injection amount for the preceding cylinders 2A and 2D is set to be larger than that in the low load side region A1. The 2D air-fuel ratio is controlled to be relatively rich, that is, a value smaller than twice the stoichiometric air-fuel ratio. As a result, the temperature of the gas introduced into the following cylinders 2B and 2C is lower than in the case where the air-fuel ratio of the preceding cylinders 2A and 2D is twice the stoichiometric air-fuel ratio (the preceding cylinder and the succeeding cylinder have the same injection amount). Although the temperature rises, the burned gas component corresponding to the EGR in the gas introduced into the subsequent cylinders 2B and 2C increases, so that knocking is suppressed by the EGR effect.
[0062]
By setting the air-fuel ratio of the preceding cylinders 2A, 2D to be a value smaller than twice the stoichiometric air-fuel ratio in the high load side region A2, the burned gas introduced into the following cylinders 2B, 2C is set. However, in this case, in addition to the burned gas derived from the preceding cylinders 2A and 2D, fresh air is introduced into the succeeding cylinders 2B and 2C. In the high load side region A2, the shortage of fresh air in the subsequent cylinders 2B and 2C is resolved, and the compression self-ignition is properly performed.
[0063]
Specifically, as shown in FIG. 5, near the intake top dead center (ITDC) of the following cylinders 2B and 2C, the burned gas introduction valves (second intake valves 31b) of the following cylinders 2B and 2C are closed. By keeping the fresh air introduction intake valve (first intake valve 31a) in the open state while maintaining the state, fresh air introduced through the intake passage 15 and the branch passage 16 is supplied to the subsequent cylinders 2B and 2C. , It is possible to secure a fresh air amount for executing the compression self-ignition in the subsequent cylinders 2B and 2C. The first intake valve 31a is closed during the intake stroke of the subsequent cylinders 2B and 2C, and the second intake valve 31b of the subsequent cylinders 2B and 2C is opened before the first intake valve 31a. The burned gas derived from the cylinders 2A and 2D can be introduced into the subsequent cylinders 2B and 2C.
[0064]
As described above, in the high load side region A2 in the compression self-ignition region A, the fresh air introduction intake valve (the first intake valve 31a) is opened near the intake top dead center (ITDC) of the subsequent cylinders 2B and 2C. Thereby, before the burned gas derived from the preceding cylinders 2A, 2D is introduced into the succeeding cylinders 2B, 2C, fresh air having a relatively low temperature is efficiently introduced into the following cylinders 2B, 2C. Can be. Moreover, in the low load side region A1 in the compression self-ignition region (partial load region) A of the subsequent cylinders 2B and 2C, the intake valve for introducing fresh air (the first intake valve 31a) is configured to be kept closed. Since the air-fuel ratio of the preceding cylinders 2A and 2D is relatively lean, the oxygen concentration of the burned gas introduced into the following cylinders 2B and 2C is maintained at a sufficiently high value. In the low load region A1 in the compression self-ignition region A of 2C, it is possible to prevent the air-fuel ratio of the following cylinders 2B, 2C from becoming lean due to the introduction of fresh air into the following cylinders 2B, 2C. .
[0065]
Further, the intake valve (first intake valve 31a) for fresh air introduction opened in the high load side region A2 in the compression self-ignition region A of the subsequent cylinders 2B and 2C is closed during the intake stroke of the subsequent cylinders 2B and 2C. By setting the valve state, fresh air is efficiently introduced into the succeeding cylinders 2B, 2C, and then the introduction of the fresh air is stopped, so that the burned gas derived from the preceding cylinders 2A, 2C is discharged. It can be smoothly introduced into 2B and 2C.
[0066]
Furthermore, as shown in the above embodiment, in the high load side region A2 in the compression self-ignition region A of the subsequent cylinders 2B and 2C, the burned gas introduction valve (second intake valve 31b) of the subsequent cylinders 2B and 2C is set to the intake stroke. During the opening of the burned gas introduction valve (second intake valve 31b), for example, near the intake top dead center (ITDC) of the subsequent cylinders 2B and 2C, for fresh air introduction. When the intake valve (first intake valve 31a) is configured to be opened, fresh air is efficiently introduced into the subsequent cylinders 2B and 2C in the high load side region A2 in the compression self-ignition region A. Then, the burned gas derived from the preceding cylinders 2A and 2D is efficiently transferred into the succeeding cylinders 2B and 2C by closing the fresh air introduction intake valve (first intake valve 31a). Can be well introduced.
[0067]
That is, as shown by the broken line in FIG. 5, the burned gas introduction valve (second intake valve 31b) can be opened near the intake top dead center ITDC of the subsequent cylinders 2B and 2C. Means that the fresh air supplied from the intake passage 15 and the burned gas supplied via the inter-cylinder gas passage 22 are simultaneously introduced into the subsequent cylinders 2B and 2C, thereby reducing the amount of fresh air introduced. Will be done. Therefore, by maintaining the burned gas introduction valve (the second intake valve 31b) in the closed state until the middle of the intake stroke of the subsequent cylinders 2B and 2C as described above, fresh air can be efficiently introduced into the subsequent cylinders 2B and 2C. Preferably, it is configured to be introduced. Further, when the burned gas introduction valve is maintained in a closed state until the middle of the intake stroke of the subsequent cylinders 2B and 2C, the amount of internal EGR in the preceding cylinders 2A and 2D can be increased. There is an advantage that compression self-ignition can be performed by increasing the internal temperatures of 2A and 2D.
[0068]
Also, in the high load side region A2 in the compression self-ignition region A of the subsequent cylinders 2B, 2C, the air-fuel ratio of the preceding cylinders 2A, 2D becomes richer than in the lower load side region A1. When the control for increasing the ratio of the fresh air introduction amount to the total gas amount introduced into the succeeding cylinders 2B, 2C is performed, the high load side region in the compression auto-ignition region A of the following cylinders 2B, 2C is performed. At A2, if the oxygen concentration in the burned gas introduced into the subsequent cylinders 2B, 2C decreases in response to the air-fuel ratio of the preceding cylinders 2A, 2D being set relatively rich, the subsequent cylinders By increasing the ratio of the fresh air introduction amount to the total gas amount introduced into the 2B, 2C, the shortage of fresh air in the following cylinders 2B, 2C is effectively eliminated, and the compression self-ignition of the following cylinders 2B, 2C is properly performed. Running while the engine Force sufficiently can be secured, and the following cylinders 2B, there is an advantage that the occurrence of knocking by suppressing temperature rise in the 2C can be effectively prevented.
[0069]
Further, at least in the compression auto-ignition region A of the subsequent cylinders 2B and 2C, the subsequent cylinder 2B is set so that the oxygen concentration in the exhaust gas discharged from the subsequent cylinders 2B and 2C becomes a value corresponding to the combustion state of the stoichiometric air-fuel ratio. , 2C is controlled, only the burned gas of the following cylinders 2B, 2C burned at the stoichiometric air-fuel ratio while burning at the lean air-fuel ratio in the preceding cylinders 2A, 2D. It is led to the exhaust passage 20. Therefore, unlike the conventional lean burn engine, there is no need to provide a lean NOx catalyst, and the three-way catalyst 24 alone ensures sufficient exhaust gas purification performance. 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.
[0070]
On the other hand, in the operation region B on the high load side or the high rotation 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 set as described above. Is in a stopped state, the actual flow paths of fresh air and gas are as shown in FIG. 8, and the intake ports 11, 11a and the exhaust ports 12a, 12 of the cylinders 2A to 2D are independent, and the intake Fresh air is introduced into the intake ports 12, 12a of the cylinders 2A to 2D from the passage 15, and burned gas is discharged from the exhaust ports 12, 12a of the cylinders 2A to 2D 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.
[0071]
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. That is, in each of the above-described embodiments, the subsequent cylinders 2B and 2C are burned by compression self-ignition in the entire operation region A which is set to the special operation mode. For example, in a region of extremely low speed and low load where the temperature and pressure in the combustion chamber are hardly attained in a state where compression self-ignition is possible, ignition of the subsequent cylinders 2B and 2C by the ignition plug 7 is performed at a predetermined ignition timing. Alternatively, combustion may be performed by forced ignition. Alternatively, when the engine temperature is low, the subsequent cylinders 2B and 2C may be burned by forced ignition.
[0072]
Further, in the above-described basic embodiment, the intake / exhaust flow state can be switched between the two-cylinder connection state and the individual cylinder independent state by using the valve stop mechanism 35, but the open / close valve is provided in the intake / exhaust passage and the inter-cylinder gas passage. By providing these passages, it is possible to switch between the two-cylinder connection state and the individual cylinder independent state by opening and closing these passages.
[0073]
Further, the device of the present invention is applicable to a multi-cylinder engine other than the four-cylinder engine. 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.
[0074]
【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 leading cylinder of the two cylinders in which the exhaust stroke and the intake stroke overlap, and the leading cylinder in the subsequent cylinder. Fuel is supplied to the burned gas with a lean air-fuel ratio introduced from, and combustion is performed by compression self-ignition, so that in the preceding cylinders, thermal efficiency is improved by lean combustion and pumping loss is reduced, and in subsequent cylinders, The fuel efficiency can be improved by improving the combustion efficiency and reducing the pumping loss by the compression self-ignition.
[0075]
In particular, in the present invention, in the high load side region in the compression self-ignition region of the subsequent cylinder, the engine output is ensured by making the air-fuel ratio of the preceding cylinder relatively rich, and correspondingly, When the oxygen concentration in the burned gas introduced into the succeeding cylinder decreases, the intake valve for fresh air introduction is opened to introduce fresh air into the succeeding cylinder. It is possible to eliminate the shortage of fresh air in the cylinder, appropriately execute the compression self-ignition of the subsequent cylinder, and effectively prevent the occurrence of knocking by increasing the burned gas component introduced into the subsequent cylinder. . For this reason, the compression self-ignition region can be greatly expanded.
[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 an explanatory diagram showing a combustion cycle, a valve opening timing, and the like of a preceding cylinder and a following cylinder.
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.
FIG. 7 is an explanatory diagram showing a substantial flow path of fresh air and gas at low load and low rotation.
FIG. 8 is an explanatory view showing a substantial fresh air and gas flow path at the time of high load and high rotation.
[Explanation of symbols]
1 Engine body
2A-2D cylinder
9 Fuel injection valve
15 Intake passage
20 Exhaust passage
22 Gas passage between cylinders
35 Valve stop mechanism
31a First intake valve (intake valve for introducing fresh air)
31b Second intake valve (burned gas introduction valve)
40 ECU
41 Operating state determination means
42 Valve stop mechanism control means (operation mode control means)
43 Intake air amount control means
44 Combustion state control means

Claims (6)

各気筒の燃焼サイクルが所定の位相差をもつように設定された多気筒の火花点火式エンジンにおいて、
エンジンの部分負荷領域で、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に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に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ガスが排気通路に導かれるような２気筒接続状態としつつ、先行気筒の空燃比を理論空燃比よりも大きいリーン空燃比として燃焼を行わせ、この先行気筒から後続気筒に導入されたリーン空燃比の既燃ガスに燃料を供給して後続気筒の燃焼を行わせる特殊運転モードの制御を実行する運転モード制御手段とを備え、
上記特殊運転モードとされる運転領域のうちの少なくとも一部の運転領域で、後続気筒を圧縮自己着火により燃焼を行わせるとともに、この圧縮自己着火領域における高負荷側領域では、それよりも低負荷側の領域に比べて先行気筒の空燃比を相対的にリッチとし、かつ後続気筒内に新気を導入する新気導入用吸気弁を開弁することにより、上記先行気筒から導出された既燃ガスに加えて新気を後続気筒内に導入させるように制御することを特徴とする火花点火式エンジンの制御装置。In a multi-cylinder spark ignition engine set such that the combustion cycle of each cylinder has a predetermined phase difference,
In a partial load region of the engine, burned gas discharged from a preceding cylinder in an exhaust stroke between a pair of cylinders in which an exhaust stroke and an intake stroke overlap is directly introduced into a succeeding cylinder in an intake stroke through an inter-cylinder gas passage. While the two cylinders are connected so that the gas discharged from the succeeding cylinder is guided to the exhaust passage, combustion is performed with the air-fuel ratio of the preceding cylinder set to a lean air-fuel ratio larger than the stoichiometric air-fuel ratio. Operating mode control means for performing control of a special operation mode for supplying fuel to burned gas having a lean air-fuel ratio introduced to the cylinder and performing combustion of a subsequent cylinder,
In at least a part of the operation region set as the special operation mode, the subsequent cylinder is caused to burn by compression self-ignition, and in the high load side region in the compression self-ignition region, a lower load is applied. By setting the air-fuel ratio of the preceding cylinder relatively rich as compared to the region on the side and opening the fresh air introduction intake valve for introducing fresh air into the succeeding cylinder, the burned fuel derived from the preceding cylinder is opened. A control device for a spark-ignition engine, which controls so that fresh air in addition to gas is introduced into a subsequent cylinder. 後続気筒の圧縮自己着火領域における低負荷側領域では、新気導入用吸気弁を閉弁状態に維持し、上記圧縮自己着火領域における高負荷側領域では、新気導入用吸気弁を後続気筒の吸気上死点付近で開弁することを特徴とする請求項１記載の火花点火式エンジンの制御装置。In the low load side region in the compression self-ignition region of the subsequent cylinder, the intake valve for fresh air introduction is maintained in the closed state, and in the high load side region in the compression self-ignition region, the intake valve for fresh air introduction is 2. The control device for a spark ignition engine according to claim 1, wherein the valve is opened near an intake top dead center. 後続気筒の圧縮自己着火領域における高負荷側領域で開弁した新気導入用吸気弁を、後続気筒の吸気行程途中で閉弁することを特徴とする請求項２記載の火花点火式エンジンの制御装置。3. The control of a spark ignition engine according to claim 2, wherein the intake valve for introducing fresh air, which is opened in the high load side region in the compression self-ignition region of the subsequent cylinder, is closed during the intake stroke of the subsequent cylinder. apparatus. 後続気筒の圧縮自己着火領域における高負荷側領域では、後続気筒の既燃ガス導入弁を吸気行程の途中で開弁し、かつこの既燃ガス導入弁の開弁時期よりも前に、新気導入用吸気弁を開弁することを特徴とする請求項１記載の火花点火式エンジンの制御装置。In the high load side region in the compression self-ignition region of the subsequent cylinder, the burned gas introduction valve of the subsequent cylinder is opened during the intake stroke, and the fresh gas is introduced before the burned gas introduction valve opens. 2. The control device for a spark ignition type engine according to claim 1, wherein the introduction intake valve is opened. 後続気筒の圧縮自己着火領域における高負荷側領域では、それよりも低負荷側の領域に比べ、先行気筒の空燃比がリッチになるのに対応して後続気筒に導入される総ガス量に対する新気導入量の割合を高めるように制御することを特徴とする請求項１記載の火花点火式エンジンの制御装置。In the high-load region in the compression self-ignition region of the succeeding cylinder, the new gas amount corresponding to the richer air-fuel ratio of the preceding cylinder becomes larger than that in the lower-load region. The control device for a spark ignition engine according to claim 1, wherein the control is performed so as to increase a ratio of an air introduction amount. 少なくとも後続気筒の圧縮自己着火領域では、後続気筒から排出される排気ガス中の酸素濃度が、理論空燃比の燃焼状態に対応した値となるように後続気筒の空燃比を制御することを特徴とする請求項１記載の火花点火式エンジンの制御装置。At least in the compression self-ignition region of the subsequent cylinder, the air-fuel ratio of the subsequent cylinder is controlled such that the oxygen concentration in the exhaust gas discharged from the subsequent cylinder has a value corresponding to the combustion state of the stoichiometric air-fuel ratio. The control device for a spark ignition engine according to claim 1.

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