JP3951852B2

JP3951852B2 - Engine control device

Info

Publication number: JP3951852B2
Application number: JP2002228790A
Authority: JP
Inventors: 義之進矢; 光夫人見; 孝司住田; 好徳林
Original assignee: Mazda Motor Corp
Current assignee: Mazda Motor Corp
Priority date: 2002-08-06
Filing date: 2002-08-06
Publication date: 2007-08-01
Anticipated expiration: 2022-08-06
Also published as: JP2004068698A

Description

【０００１】
【発明の属する技術分野】
本発明は、エンジンの制御装置に関し、より詳しくは、多気筒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、例えば特開平１０−２７４０８５号公報に示されるように、燃焼室内に直接燃料を噴射する燃料噴射弁を備え、低回転低負荷域等では上記燃料噴射弁から圧縮行程で燃料を噴射することにより成層燃焼を行わせ、これによって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。
【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、上記公報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】
【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浄化性能の確保のために上記リーンＮＯｘ触媒を必要とする。そして、高負荷域等の理論空燃比で運転される領域での排気浄化のために三元触媒も必要であって、この三元触媒に加えて上記リーンＮＯｘ触媒が設けられ、かつ、このリーンＮＯｘ触媒はＮＯｘ吸着量をある程度確保するために比較的大容量が必要となり、また、三元触媒と比べて高価であるため、コスト的に不利である。
【０００５】
しかも、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吸着量が増大するような所定の期間毎に、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り、これにより、リーン燃焼による燃費改善効果が目減りしてしまうことになる。
【０００６】
そこで、本願出願人は、かかる課題に鑑み、吸気、圧縮、膨張、排気の各行程からなるサイクルを行う多気筒エンジンにおいて、低負荷低回転域では、排気行程と吸気行程が重なる一対の気筒間において排気行程側の気筒である先行気筒から排出される既燃ガスをそのまま吸気行程側の気筒である後続気筒に導入し、この後続気筒から排出されるガスを三元触媒を備えた排気通路に導くようにするとともに、この２気筒接続状態にあるときに、上記先行気筒において理論空燃比よりも所定量大きいリーン空燃比とした状態で燃焼を行わせ、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とした状態で燃焼を行わせるように燃焼状態等を制御（特殊運転モードという）する一方、高負荷高回転域では、通常通り、各気筒毎を理論空燃比で燃焼を行わせるように燃焼状態等を制御（通常運転モードという）することを考えた（特願２００２−０２４５４８号）。
【０００７】
これによると、低負荷低回転域において特殊運転モードとされることにより、先行気筒ではリーン空燃比での燃焼が行われ、熱効率が高められるとともにポンピングロス（ポンプ損失）が低減されることにより大幅な燃費改善効果が得られ、また、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が供給されて理論空燃比とされた状態で燃焼が行われて、ポンピングロス低減による燃費効果が得られる。しかも、後続気筒から排出される理論空燃比の既燃ガスのみが三元触媒を備えた排気通路に導かれるため、三元触媒だけで充分に排気浄化性能が確保され、リーンＮＯｘ触媒も不要となる。
【０００８】
また、出願人は、上記のような燃焼状態等の制御において、特に特殊運転モード時に、先行気筒から後続気筒に高温の既燃ガスが導入されるのを利用して、後続気筒において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ことにより熱効率を高め、これにより燃費の改善を図ることも考えている（特願２００２−２９８３６号）。
【０００９】
ところで、上記のように各気筒の燃焼状態等を制御する場合には、特に特殊運転モードにおいて、気筒間の生成トルクに差が生じることが考えられる。すなわち、後続気筒については、先行気筒から排出される（押し出される）既燃ガスがそのまま送り込まれるため、先行気筒に比してポンピングロスの低減効果が大きくなる。また、上述のように、後続気筒において圧縮自己着火による燃焼が行われる場合には、先行気筒に比べて後続気筒の方が熱損失が少なくなる。従って、このようなポンピングロスを含む気筒間の熱効率の差異により、先行気筒と後続気筒との間にトルク差が生じることが考えられる。しかしながら、このような気筒間のトルク差は、ＮＶＨ性能（騒音振動防止性能）を低下させる要因の一つとなるため、何らかの対策を講じる必要がある。
【００１０】
本発明は以上のような課題を考慮してなされたものであり、リーン燃焼による燃費改善効果をもたせつつ、リーンＮＯｘ触媒を必要とせず三元触媒を用いるだけで、排気浄化性能を向上することすることができ、しかも、各気筒間の生成トルクの均衡を保つことができるエンジンの制御装置を提供するものである。
【００１１】
【課題を解決するための手段】
上記課題を解決するために、本発明は、各気筒にそれぞれ新気を導入する各気筒独立状態と、排気行程と吸気行程が重なる一対の気筒間において先行気筒の既燃ガスを気筒間ガス通路を介して後続気筒に導入する２気筒接続状態とに吸気および排気の流通経路が切換え可能に構成され、かつ、この流通経路を前記各気筒独立状態として各気筒においてそれぞれ独立して燃焼を行わせる通常運転モードと、前記２気筒接続状態として先行気筒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に導入して燃焼を行わせる特殊運転モードとに運転モードを切換え可能に構成される多気筒のエンジンの制御装置であって、前記特殊運転モードにあるときに、先行気筒への吸入空気量に基づき、先行気筒では理論空燃比よりも所定量だけ大きいリーン空燃比とした状態で燃焼を行わせ、かつ後続気筒では、先行気筒から導出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とした状態で燃焼を行わせ得るように先行気筒および後続気筒に対する燃料噴射量の総和を求める総燃料噴射量演算手段と、先行気筒と後続気筒との生成トルクが均等となるようにエンジンの運転状態に応じて後続気筒に対する先行気筒の空燃比の比率を設定する比率設定手段と、この比率設定手段で設定される比率と前記総燃料噴射量演算手段において求められる燃料噴射量の総和とに基づいて先行気筒および後続気筒に対する最終的な燃料噴射量を求める最終燃料噴射量演算手段とを備えているものである。
【００１２】
この発明によると、例えばエンジンの低負荷低回転域において、前記特殊運転モードの燃焼制御が実行されることにより、先行気筒ではリーン空燃比での燃焼が行われて、熱効率が高められるとともにポンピングロス（ポンプ損失）が低減されることにより大幅な燃費改善効果が得られ、かつ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が供給されて理論空燃比とされた状態で燃焼が行われることにより、少なくともポンピングロス低減による燃費効果が得られる。また、後続気筒から排出される理論空燃比の既燃ガスのみが排気通路に導かれるため、三元触媒だけで充分に排気浄化性能が確保される。一方、高負荷・高回転の運転領域では、通常運転モードに設定されることにより出力性能が確保される。
【００１３】
また、前記特殊運転モードでは、エンジンの運転状態に応じて、先行気筒と後続気筒との生成トルクが均等となるように各気筒に対する燃料噴射量が求められて噴射されることにより、ポンピングロスを含む熱効率の差異に起因する先後気筒間における生成トルク差の発生が解消、あるいは抑制され、ＮＶＨ性能（騒音振動防止性能）が高められる。
【００１４】
なお、比率設定手段で設定する空燃比の比率は、先行気筒および後続気筒のトルク差を実験的に求め、そのデータから定めてもよいし、また、先行気筒および後続気筒におけるポンプ損失又は熱効率に関連するパラメータに基づいて定めるようにしてもよい。
【００１５】
この発明においては、前記比率設定手段において求められる空燃比の比率に基づく先行気筒および後続気筒の燃焼状態が正常燃焼可能な範囲内にあるか否かを事前に判定する可燃判定手段をさらに備え、可燃判定手段における判定結果が肯定的な場合にのみ前記比率設定手段において設定された比率に基づいて燃料噴射量を演算するように前記最終燃料噴射量演算手段が構成されているのが好ましい。
【００１６】
すなわち、理論的に先後気筒間のトルクバランスを保ち得るような空燃比の比率が求められた場合であっても、当該比率によると先行気筒又は後続気筒において正常な燃焼を行うことが困難となる場合もあり、上記の構成によると、そのような比率が求められた場合には、該比率に基づく燃料噴射量の演算が禁止される。これにより、先後気筒間のトルクバランスを保ち得るような空燃比の比率に基づいて無制限に燃料噴射量が設定されることによる弊害、つまり失火や異常燃焼（ノッキング）の発生を未然に防止することが可能となる。
【００１７】
なお、可燃判定手段における判定結果が否定的である場合には、各気筒において正常燃焼が行われる範囲内で予め設定された比率に基づいて各気筒に対する最終燃料噴射量を演算するように前記最終燃料噴射量演算手段を構成すればよい。
【００１８】
このようにすれば、失火や異常燃焼（ノッキング）の発生を防止して、先行気筒および後続気筒での燃焼を正常に行わせることができる。
【００１９】
また、前記発明においては、特殊運転モードにあるときに、後続気筒における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運転状態に応じて選択する着火制御手段を備え、前記強制点火による燃焼か圧縮自己着火による燃焼かに応じて、設定する空燃比の比率を変更するように前記比率設定手段が構成されているのが好ましい。
【００２０】
このようにすれば、前記特殊運転モードであって、後続気筒で圧縮自己着火により燃焼が行われる場合には、後続気筒における燃焼効率がさらに向上し、より一層の燃費改善効果が得られる。また、このように後続気筒において圧縮自己着火が行われる場合には、後続気筒の熱効率が向上するため先行気筒と後続気筒のトルク差が大きくなるが、圧縮自己着火が行われない場合とは異なる空燃比の前記比率、つまり後続気筒の熱効率の向上を加味した比率が設定されることにより、圧縮自己着火が行われない場合と同様に、先後気筒間における生成トルク差の発生が良好に解消、あるいは抑制される。
【００２１】
【発明の実施の形態】
以下、図面に基づいて本発明の実施の形態について説明する。
【００２２】
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の燃焼室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、この燃料噴射弁９には、図外の燃料ポンプにより燃料供給通路等を介して燃料が供給され、かつ、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ように燃料供給系統が構成されている。
【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、各気筒２Ａ〜２Ｄが所定の位相差をもって吸気、圧縮、膨張、排気の各行程からなる燃焼サイクルを行うようになっており、４気筒エンジンの場合、気筒列方向一端側から１番気筒２Ａ、２番気筒２Ｂ、３番気筒２Ｃ、４番気筒２Ｄと呼ぶと、図６に示すように上記サイクルが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燃焼サイクルが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、Ｆは燃料噴射、Ｓは点火をそれぞれ表している。
【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と２番気筒２Ｂ、及び４番気筒２Ｄと３番気筒２Ｃ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、２番気筒２Ｂ及び３番気筒２Ｃが後続気筒となる。
【００２８】
各気筒の吸・排気ポートとこれに接続される吸気通路、排気通路及び気筒間ガス通路は、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２Ａ，２Ｄにおける吸気ポート１１および２番，３番気筒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および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、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および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、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に気筒間ガス通路２２の下流端が接続されている。
【００３３】
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理論空燃比検出用の排気ガス濃度検出手段であるＯ₂センサ２３が設けられ、さらにその下流の排気通路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が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。また、Ｏ₂センサ２３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もので、特に理論空燃比付近で出力が急変するλＯ₂センサにより構成されている。
【００３４】
上記気筒間ガス通路２２には、排気ガス中の酸素濃度の変化（空燃比の変化）に対して出力がリニアに変化するリニアＯ₂センサ２５が設けられている。
【００３５】
各気筒の吸・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。すなわち、１番，４番気筒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。
【００３６】
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。
【００３７】
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、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がそれぞれ設けられている（図３参照）。
【００３８】
図３はエンジンの駆動、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、Ｏ₂センサ２３及びリニアＯ₂センサ２５からの信号が入力されるとともに、エンジンの冷却水温度を検出する水温センサ５１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５２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５３等からの信号も入力されている。また、この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３７，３９とに対して制御信号が出力されている。
【００３９】
上記ＥＣＵ４０は、その機能構成として運転状態判別手段４１、温度状態判別手段４２、モード設定手段４３、弁停止機構制御手段４４、吸入空気量制御手段４５、燃料制御手段４６および点火制御手段４７等を備えている。
【００４０】
運転状態判別手段４１は、図５に示すようにエンジンの運転領域が低速低負荷側の領域Ａと高速側ないし高負荷側の領域Ｂとに分けられた制御用マップを有し、低速低負荷側の領域Ａを特殊運転モード領域、高速側ないし高負荷側の領域Ｂを通常運転モード領域とする。そして、上記回転数センサ５２及びアクセル開度センサ５３等からの信号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）が上記領域Ａ，Ｂのいずれにあるかを判別する。
【００４１】
温度状態判別手段４２は、水温センサ５１からの信号によってエンジンの温度状態を判別するものであり、水温（エンジン温度）が所定値以下の低温時か所定温度より高い高温時かを判別するようになっている。
【００４２】
モード設定手段４３は、運転状態判別手段４１による判別に基づき、前記特殊運転モード領域Ａでは、排気行程にある先行気筒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に導入して燃焼させる特殊運転モードを選択し、前記通常運転モード領域領域Ｂでは、各気筒をそれぞれ独立させて燃焼させる通常運転モードを選択する。
【００４３】
弁停止機構制御手段４４は、モード設定手段４３によるモードの設定に応じ、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（１番、４番気筒）２Ａ，２Ｄの既燃ガスを後続気筒（２番、３番気筒）２Ｂ，２Ｃに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、運転状態が領域Ａ，Ｂのいずれかにあるかに応じ、上記コントロール弁３７，３９を制御することにより、各弁停止機構３５を次のように制御する。
領域Ａ：（特殊運転モード）
第１排気弁３２ａ及び第１吸気弁３１ａを停止状態
第２排気弁３２ｂ及び第２吸気弁３１ｂを作動状態
領域Ｂ：（通常運転モード）
第１排気弁３２ａ及び第１吸気弁３１ａを作動状態
第２排気弁３２ｂ及び第２吸気弁３１ｂを停止状態
【００４４】
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２Ｂ，２Ｃに対する分岐吸気通路１６からの吸気導入が遮断された状態で、先行気筒２Ａ，２Ｄ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が先行気筒２Ａ，２Ｄに供給されるようにスロットル開度を調節する。
【００４５】
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、点火制御手段４７は、運転状態に応じた点火時期の制御を行う。そして、モード設定手段４３により設定されるモードに応じて、燃焼状態の制御（燃料の制御および点火の制御）が変更される。
【００４６】
すなわち、特殊運転モードが設定された場合には、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、好ましくは理論空燃比の略２倍もしくはそれ以上とするように燃料噴射量を制御するとともに、圧縮行程で燃料を噴射して成層燃焼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。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燃料を供給して理論空燃比とするように燃料噴射量を制御する。また、既燃ガスが多い状況下で着火、燃焼が可能なように噴射時期を設定し、例えば着火性確保のため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。
【００４７】
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。
【００４８】
なお、特殊運転モードが選択された場合には、一対の気筒の両方に対する燃料噴射量の総和が先行気筒２Ａ，２Ｄに導入される空気量に対して理論空燃比となる量であって、かつ先行気筒２Ａ，２Ｄの生成トルクと後続気筒２Ｂ，２Ｃの生成トルクとが均等になるように、先行気筒２Ａ，２Ｄに対する燃料噴射量と後続気筒２Ａ，２Ｄに対する燃料噴射量の割合が調整される。
【００４９】
この点にき、図４に基づき燃料制御手段４６のより詳細な構成について説明する。すなわち、燃料制御手段４６は、その機能構成として、同図に示すように総燃料噴射量演算部５５、最終燃料噴射量演算部５６、トルクバランス空燃比設定部５７、可燃判定部５８および分割比演算部５９を有している。
【００５０】
総燃料噴射量演算部５５は、エアフローセンサ１９により検出される吸入空気量に基づいて燃料噴射弁９から噴射する燃料噴射量を求めるものであり、特に、特殊運転モードが選択された場合には、先行気筒に対する燃料噴射量と後続気筒に対する燃料噴射量の総和（総燃料噴射量）を求める。この場合には、上述のように燃料噴射量の総和が先行気筒２Ａ，２Ｄに導入される空気量に対して理論空燃比となるように燃料噴射量が求められる。
【００５１】
最終燃料噴射量演算部５６は、最終的に制御する燃料噴射量を決定するもので、通常運転モードが選択された場合には、総燃料噴射量演算部５５において求められる総燃料噴射量をそのまま最終的な燃料噴射量とする。一方、特殊運転モードが選択された場合には、総燃料噴射量と後述する分配比とから先行気筒および後続気筒に対する燃料噴射量をそれぞれ演算し、これを最終的な燃料噴射量とする。
【００５２】
トルクバランス空燃比設定部５７、可燃判定部５８および分割比演算部５９は、特殊運転モードが選択された場合に機能するもので、トルクバランス空燃比設定部５７は、回転数センサ５２及びアクセル開度センサ５３等からの信号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）に応じて予め記憶されたマップから先行気筒の空燃比を求める。マップは、例えば、ポンピングロスを含む熱効率の差に起因して先行気筒と後続気筒との間に生じる生成トルクの差を予め実験的に求め、このトルク差が「０」となり得るように設計上の空燃比（つまり、理論空燃比よりも大きいリーン空燃比、好ましくは理論空燃比の略２倍もしくはそれ以上となる空燃比であって設計上求められる値）を補正し、これをエンジンの運転状態に対応付けたものである。
【００５３】
可燃判定部５８は、トルクバランス空燃比設定部５７で求められた空燃比の比率に基づいて燃焼が正常に行われ得るか否かを水温センサ５１からの信号により判別されるエンジン温度とマップ等から事前に判定するものであり、その結果を分割比演算部５９に出力するものである。
【００５４】
分割比演算部５９は、前記マップから求められる先行気筒の空燃比と後続気筒の空燃比（理論空燃比）との比率から先行・後続気筒に対する燃料（前記総燃料噴射量）の分配比を定めるもので、可燃判定部５８での判定結果が肯定的である場合、つまり正常燃焼が行われると判定された場合には、トルクバランス空燃比設定部５７で求められた空燃比の比率に基づいて燃料の分割比を求め、その結果を最終燃料噴射量演算部５６に出力する。一方、可燃判定部５８での判定結果が否定的である場合、すなわち失火やノッキングなどが生じる虞れがあるような場合には、先行・後続気筒において正常燃焼が行われ得る範囲内で予め設定された分配比、例えば先行気筒の空燃比を設計上の値（トルク差に基づく補正前の値）とした場合の分配比を最終燃料噴射量演算部５６に出力するものである。なお、当実施形態では、この分割比演算部５９および前記最終燃料噴射量演算部５６により本願発明の最終燃料噴射量手段が構成されている。
【００５５】
次に、以上のような実施形態の装置の作用を、図６〜図８を参照しつつ説明する。
【００５６】
低負荷低回転側の運転領域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７に示すように、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から排出される既燃ガスのみが三元触媒２４を備えた排気通路２０に導かれるような２気筒接続状態とされる。
【００５７】
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、リーン空燃比での成層燃焼が行われる（図６参照）。
【００５８】
その後、先行気筒２Ａ，２Ｄの排気行程と後続気筒２Ｂ，２Ｃの吸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気筒間ガス通路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、適当なタイミング（例えば圧縮行程）で燃料が噴射され、かつ、所定点火時期に点火が行われて燃焼が行われる（図６参照）。そして、後続気筒２Ｂ，２Ｃでの燃焼後の既燃ガスは、三元触媒２４を備えた排気通路２０に排出される（図７中の矢印ｃ）。
【００５９】
このように、先行気筒２Ａ，２Ｄではリーン空燃比での成層燃焼が行われること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、これらの相乗効果で大幅に燃費が改善される。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燃焼が行われることにより、先行気筒２Ａ，２Ｄのようにリーン空燃比で成層燃焼が行われるものと比べると熱効率では多少劣るものの、先行気筒２Ａ，２Ｄから押し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、これにより燃費改善効果が充分に得られる。
【００６０】
しかも、後続気筒２Ｂ，２Ｃから排気通路２０に排出される既燃ガスは理論空燃比に対応した値となるため、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こととなる。そして、このようにリーン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。
【００６１】
また、先行気筒２Ａ，２Ｄでは理論空燃比の略２倍もしくはそれ以上のリーン空燃比とされることでＮＯｘ発生量が比較的少なく抑えられ、後続気筒２Ｂ，２Ｃでは、先行気筒２Ａ，２Ｄから既燃ガスが導入されることで多量のＥＧＲが行われているのと同等の状態となることからＮＯｘの発生が充分に抑制される。このような点からもエミッションの向上に有利となる。
【００６２】
さらに、先行気筒２Ａ，２Ｄおよび後続気筒２Ｂ，２Ｃに対しては、エンジンの運転状態に応じて気筒間の生成トルクが均等となるように燃料噴射量が制御され（総燃料が分配され）、これによりＮＶＨ性能（騒音振動防止性能）が良好に高められる。すなわち、先行気筒２Ａ，２Ｄと後続気筒２Ｂ，２Ｃの間には、上記のようにポンピングロスを含む熱効率の差が存在するため、先行気筒２Ａ，２Ｄおよび後続気筒２Ｂ，２Ｃに対して燃料噴射量を均等に制御すると、気筒間の生成トルク差により振動や騒音が発生することが考えられるが、上記実施形態の装置によると、燃料制御手段４６により上述したように燃焼噴射量が制御されるため、先行気筒２Ａ，２Ｄと後続気筒２Ｂ，２Ｃとの間にトルク差が生じることが殆どなく、従って、トルク差に起因する振動等の発生が有効に防止されることとなる。
【００６３】
その上、燃料制御手段４６において燃料噴射量（分配比）を設定する過程では、先行・後続気筒において正常に燃焼が行われるか否かを自演に判定し（可燃判定部５８での判定）、正常燃焼が行われない虞れがある場合には、正常燃焼が可能な範囲で燃料噴射量を決定するようにしているので、生成トルクが均等になるように燃料噴射量を制御することによる弊害、例えば失火やノッキングの発生を有効に防止することができる。つまり、生成トルクの均一化を優先すると、先行気筒又は後続気筒の空燃比が正常燃焼可能な範囲をこえるような範囲で燃料噴射量（分配比）が設定されることも考えられ、この場合には、失火やノッキングなどを招くことが考えられる。しかし、この実施形態の装置によると、上記のような判定が事前に行われた上で燃料噴射量が決定されるので、常に、正常燃焼が可能な範囲で先行・後続気筒に対する燃料噴射量が制御されることとなる。従って、失火やノッキングの発生を防止して、正常な運転状態を確保することができる。
【００６４】
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、実質的に各気筒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（λ≦１）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。
【００６５】
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。
【００６６】
例えば、特殊運転モードにおいて、エンジンの温度状態の判別に基づき、低温時には後続気筒での燃焼を強制点火により行わせる強制点火モードとし、高温時には後続気筒での燃焼を圧縮自己着火により行わせる圧縮自己着火モードとし、とくに圧縮自己着火モードにおいては吸気行程で燃料を噴射するように制御してもよい。すなわち、特殊運転モードでは、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では吸気行程で燃料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に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。従って、高温時には後続気筒での燃焼を圧縮自己着火モードとして圧縮自己着火とすることにより燃焼を急速に行わせ、これにより熱効率を大幅に向上させることができる。
【００６７】
なお、このような圧縮自己着火モードにおいても、先行気筒２Ａ，２Ｄおよび後続気筒２Ｂ，２Ｃに対して気筒間の生成トルクが均等となるように燃料噴射量が制御されることによりＮＶＨ性能（騒音振動防止性能）が高められ、その結果、良好な運転状態を確保することが可能となるが、この場合には、トルクバランス空燃比設定部５７に空燃比の比率を求めるためのマップとして強制点火モードとは別に圧縮自己着火モードのものを記憶させておき、選択されたモードに対応するマップに基づいて空燃比の比率を設定するように構成することが必要となる。つまり、圧縮自己着火では、熱損失が少なく、また急速燃焼により気筒内温度が上昇し難いため、強制点火の場合に比べて熱効率が良くなる。そのため、圧縮自己着火モードでは、強制点火モードに比べて先行気筒２Ａ，２Ｄと後続気筒２Ｂ，２Ｃとの間の生成トルク差がさらに大きくなると考えられ、強制点火の場合と共通のマップに基づいて空燃比の比率を求めるだけでは、先行・後続気筒間のトルク差を完全に解消することが困難なためである。
【００６８】
なお、点火モードの切換については、上述のように特殊運転モード領域Ａの全体でエンジンの温度状態の判別に基づいて高温時に圧縮自己着火モード、低温時に強制点火モードとするように制御する以外に、例えば、エンジンの低速域（図５の領域Ａのうちの低速域）でのみエンジンの温度状態に応じて圧縮自己着火モードと強制点火モードとを切換え、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。
【００６９】
また、上記実施形態では、先行・後続気筒の燃料噴射量を決定する基礎となるマップ（空燃比の比率を設定するトルクバランス空燃比設定部５７に記憶されたマップ）の値は、先行気筒２Ａ，２Ｄと後続気筒２Ｂ，２Ｃの生成トルク差を実験的に求め、このトルク差が「０」となり得るように設計上の空燃比（理論空燃比よりも大きいリーン空燃比、好ましくは理論空燃比の略２倍もしくはそれ以上となる空燃比であって設計上求められる値）を補正したものであるが、勿論、先行・後続気筒のポンピングロス、熱効率に関するパラメータから設計値を用いて理論的（演算）に求めた値を用いてもよい。
【００７０】
【発明の効果】
以上のように本発明の制御装置は、各気筒においてそれぞれ独立して燃焼を行わせる通常運転モードと、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をそのまま吸気行程にある後続気筒に導入して燃焼を行わせる特殊運転モードとに切換え可能に構成され、例えば低負荷低回転の運転領域では特殊運転モードに設定されることにより、先行気筒ではリーン空燃比での燃焼が行われて、熱効率が高められるとともにポンピングロスが低減されることにより大幅な燃費改善効果が得られ、かつ上記後続気筒では先行気筒から導入されたリーン空燃比の既燃ガスに燃料が供給されて理論空燃比とされた状態で燃焼が行われることにより、少なくともポンピングロス低減による燃費効果が得られる。また、後続気筒から排出される理論空燃比の既燃ガスのみが排気通路に導かれるため、三元触媒だけで充分に排気浄化性能が確保される。一方、高負荷・高回転の運転領域では、通常運転モードに設定されることにより出力性能が確保される。そして、特殊運転モードでは、エンジンの運転状態に応じて、先行気筒と後続気筒との生成トルクが均等となるように各気筒に対する燃料噴射量が求められて噴射されることにより、ポンピングロスや熱効率の差異に起因する先後気筒間における生成トルク差の発生が解消、あるいは抑制され、その結果、ＮＶＨ性能（騒音振動防止性能）が高められることになる。
【図面の簡単な説明】
【図１】本発明に係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】燃料制御手段の機能構成を示すブロック図である。
【図５】運転領域を示す説明図である。
【図６】各気筒の排気行程、吸気行程、燃料噴射時期およ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およびガスの流通経路を示す説明図である。
【図８】高負荷、高低回転側の運転領域にある時の実質的な新気およびガスの流通経路を示す説明図である。
【符号の説明】
１エンジン本体
２Ａ〜２Ｄ気筒
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
２４三元触媒
３１吸気弁
３１ａ第１吸気弁
３１ｂ第２吸気弁
３２排気弁
３２ａ第１排気弁
３２ｂ第２排気弁
３５弁停止機構
４０ＥＣＵ
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点火制御手段
５５総燃料噴射量演算部（総燃料噴射量演算手段）
５６最終燃料噴射量演算部（最終燃料噴射量演算手段）
５７トルクバランス空燃比設定部（比率設定手段）
５８可燃判定部（可燃判定手段）
５９分割比演算部（最終燃料噴射量演算手段）[0001]
BACKGROUND OF THE INVENTION
The present invention relates to an engine control apparatus, and more particularly to an apparatus for controlling the combustion state of each cylinder in order to improve fuel consumption and emissions in a multi-cylinder engine.
[0002]
[Prior art]
Conventionally, in a spark ignition engine, a technique 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 is known. As disclosed in Japanese Patent Application Laid-Open No. 10-274085, a stratified combustion is provided by injecting fuel directly into a combustion chamber and injecting fuel from the fuel injection valve in a compression stroke in a low rotation and low load region. It is known that the super lean combustion is realized by this.
[0003]
In such an engine, an ordinary three-way catalyst (a catalyst having a high purification performance in the vicinity of the theoretical air-fuel ratio with respect to HC, CO, and NOx) alone as an exhaust gas purification catalyst is sufficient for NOx during lean operation. Since purification performance cannot be obtained, a lean NOx catalyst is provided that adsorbs NOx in an oxygen-excessive atmosphere and removes and reduces NOx in an oxygen-concentrated atmosphere as shown in the above publication. When such a lean NOx catalyst is used, if the NOx adsorption amount of the lean NOx catalyst increases during the lean operation, for example, as shown in the above publication, additional fuel is injected during the expansion stroke in addition to the main combustion. As a result, the air-fuel ratio of the exhaust gas is enriched and CO is generated, thereby promoting NOx separation and reduction.
[0004]
[Problems to be solved by the invention]
The engine that performs the conventional lean operation as described above requires the lean NOx catalyst in order to ensure the NOx purification performance during the lean operation. Further, a three-way catalyst is also required for exhaust purification in a region operated at a stoichiometric air-fuel ratio such as a high load region, the lean NOx catalyst is provided in addition to the three-way catalyst, and the lean A NOx catalyst requires a relatively large capacity in order to secure a certain amount of NOx adsorption, and is expensive compared to a three-way catalyst, which is disadvantageous in terms of cost.
[0005]
Moreover, in order to maintain the purification performance of the lean NOx catalyst, the NOx is temporarily removed by the removal of NOx, the supply of additional fuel for reduction, etc. at predetermined intervals such that the NOx adsorption amount increases as described above. It is necessary to enrich the air-fuel ratio, which reduces the fuel efficiency improvement effect due to lean combustion.
[0006]
Therefore, in view of such a problem, the applicant of the present application, in a multi-cylinder engine that performs a cycle including intake, compression, expansion, and exhaust strokes, between a pair of cylinders in which an exhaust stroke and an intake stroke overlap in a low load low rotation range. In this case, the burned gas discharged from the preceding cylinder, which is the cylinder on the exhaust stroke side, is directly introduced into the subsequent cylinder, which is the cylinder on the intake stroke side, and the gas discharged from the subsequent cylinder is introduced into the exhaust passage provided with the three-way catalyst. In addition, when the two cylinders are connected, combustion is performed in a state where the preceding cylinder has a lean air-fuel ratio that is a predetermined amount larger than the stoichiometric air-fuel ratio, and in the succeeding cylinder, the lean introduced from the preceding cylinder is performed. While controlling the combustion state etc. so that combustion is performed in a state where the stoichiometric air-fuel ratio is supplied by supplying fuel to the burned gas of the air-fuel ratio (referred to as a special operation mode), in the high load high rotation range Usual, I thought to control the combustion conditions or the like so as to perform the combusting respective cylinders at the stoichiometric air-fuel ratio (referred to the normal operation mode) (Japanese Patent Application No. 2002-024548).
[0007]
According to this, by setting the special operation mode in the low load and low rotation range, the preceding cylinder is burned at a lean air-fuel ratio, greatly increasing the heat efficiency and reducing the pumping loss (pump loss). In the subsequent cylinder, fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder to achieve the stoichiometric air-fuel ratio, thereby reducing the pumping loss. A fuel efficiency effect is obtained. Moreover, since only the stoichiometric burned gas discharged from the subsequent cylinders is guided to the exhaust passage provided with the three-way catalyst, sufficient exhaust purification performance is ensured with only the three-way catalyst, and no lean NOx catalyst is required. Become.
[0008]
In addition, in the control of the combustion state and the like as described above, the applicant uses the fact that high-temperature burned gas is introduced from the preceding cylinder to the succeeding cylinder, particularly in the special operation mode. It is also considered that fuel efficiency is improved by supplying fuel to the fuel and causing combustion by compression self-ignition, thereby improving fuel efficiency (Japanese Patent Application No. 2002-29836).
[0009]
By the way, when the combustion state of each cylinder is controlled as described above, it is conceivable that a difference occurs in the generated torque between the cylinders particularly in the special operation mode. That is, for the subsequent cylinder, the burned gas discharged (pushed out) from the preceding cylinder is sent as it is, so that the effect of reducing the pumping loss is greater than that of the preceding cylinder. Further, as described above, when combustion by compression self-ignition is performed in the subsequent cylinder, the subsequent cylinder has less heat loss than the preceding cylinder. Therefore, it is conceivable that a torque difference occurs between the preceding cylinder and the succeeding cylinder due to the difference in thermal efficiency between the cylinders including the pumping loss. However, such a torque difference between the cylinders is one of the factors that deteriorate the NVH performance (noise vibration prevention performance), and therefore it is necessary to take some measures.
[0010]
The present invention has been made in consideration of the above-mentioned problems, and improves exhaust gas purification performance by using a three-way catalyst without the need for a lean NOx catalyst while providing fuel efficiency improvement effect by lean combustion. In addition, the present invention provides an engine control device that can maintain the balance of the generated torque among the cylinders.
[0011]
[Means for Solving the Problems]
In order to solve the above-described problems, the present invention provides an inter-cylinder gas passage for the burned gas of a preceding cylinder between a pair of cylinders in which an exhaust stroke and an intake stroke overlap each other and a cylinder independent state in which fresh air is introduced into each cylinder. The intake and exhaust flow paths are configured to be switchable to the two-cylinder connection state introduced to the succeeding cylinders via the cylinder, and each of the cylinders performs combustion independently with this flow path set as the cylinder independent state. The operation mode can be switched between a normal operation mode and a special operation mode in which the burned gas discharged from the preceding cylinder in the two-cylinder connected state is directly introduced into the succeeding cylinder in the intake stroke to perform combustion. A control device for a multi-cylinder engine, which is larger than the stoichiometric air-fuel ratio in the preceding cylinder based on the intake air amount to the preceding cylinder when in the special operation mode. Combustion is performed in a state where the lean air-fuel ratio is set, and in the subsequent cylinder, fuel is supplied to burned gas having a lean air-fuel ratio derived from the preceding cylinder so that combustion can be performed in the state where the stoichiometric air-fuel ratio is set. The total fuel injection amount calculation means for obtaining the sum of the fuel injection amounts for the preceding cylinder and the succeeding cylinder, and the empty space of the preceding cylinder for the succeeding cylinder according to the operating state of the engine so that the generated torque of the preceding cylinder and the succeeding cylinder becomes equal. The final fuel for the preceding cylinder and the succeeding cylinder based on the ratio setting means for setting the ratio of the fuel ratio, the ratio set by the ratio setting means and the sum of the fuel injection amounts obtained by the total fuel injection amount calculation means And a final fuel injection amount calculating means for obtaining an injection amount.
[0012]
According to the present invention, for example, in the low load and low rotation region of the engine, the combustion control in the special operation mode is executed, so that combustion at the lean air-fuel ratio is performed in the preceding cylinder, and the thermal efficiency is increased and the pumping loss is increased. (Pump loss) is reduced, and a significant fuel efficiency improvement effect is obtained. In the succeeding cylinder, fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder so that the stoichiometric air-fuel ratio is obtained. By performing the combustion, at least the fuel efficiency effect by reducing the pumping loss can be obtained. Further, since only the stoichiometric burned gas discharged from the succeeding cylinder is guided to the exhaust passage, the exhaust purification performance is sufficiently ensured only by the three-way catalyst. On the other hand, in the high load / high rotation operation region, the output performance is ensured by setting the normal operation mode.
[0013]
Further, in the special operation mode, the pumping loss is reduced by determining and injecting the fuel injection amount for each cylinder so that the generation torque of the preceding cylinder and the succeeding cylinder is equal according to the operating state of the engine. The generation torque difference between the front and rear cylinders due to the difference in the thermal efficiency is eliminated or suppressed, and the NVH performance (noise vibration prevention performance) is enhanced.
[0014]
Note that the ratio of the air-fuel ratio set by the ratio setting means may be determined from the data obtained by experimentally obtaining the torque difference between the preceding cylinder and the succeeding cylinder, and may be determined by the pump loss or the thermal efficiency in the preceding cylinder and the succeeding cylinder. It may be determined based on related parameters.
[0015]
In the present invention, it further comprises flammability determining means for determining in advance whether or not the combustion state of the preceding cylinder and the succeeding cylinder based on the ratio of the air-fuel ratio determined by the ratio setting means is within a normal combustible range, It is preferable that the final fuel injection amount calculating means is configured to calculate the fuel injection amount based on the ratio set in the ratio setting means only when the determination result in the combustible determination means is affirmative.
[0016]
That is, even when an air-fuel ratio that can theoretically maintain the torque balance between the front and rear cylinders is obtained, it is difficult to perform normal combustion in the preceding cylinder or the succeeding cylinder according to the ratio. In some cases, according to the above configuration, when such a ratio is obtained, calculation of the fuel injection amount based on the ratio is prohibited. As a result, it is possible to prevent the adverse effects of setting an unlimited amount of fuel injection based on the air-fuel ratio that can maintain the torque balance between the front and rear cylinders, that is, the occurrence of misfires and abnormal combustion (knocking). Is possible.
[0017]
When the determination result by the flammability determination means is negative, the final fuel injection amount for each cylinder is calculated based on a preset ratio within a range where normal combustion is performed in each cylinder. What is necessary is just to comprise a fuel injection amount calculation means.
[0018]
In this way, it is possible to prevent misfire and abnormal combustion (knocking) from occurring, and to perform combustion in the preceding cylinder and the succeeding cylinder normally.
[0019]
In the invention, the ignition control means for selecting whether the combustion in the subsequent cylinder is performed by compression self-ignition or forced ignition when in the special operation mode, according to the operating state of the engine, Preferably, the ratio setting means is configured to change the ratio of the air-fuel ratio to be set depending on whether combustion is performed by forced ignition or combustion by compression self-ignition.
[0020]
In this way, in the special operation mode, when combustion is performed by compression self-ignition in the subsequent cylinder, the combustion efficiency in the subsequent cylinder is further improved, and a further fuel efficiency improvement effect is obtained. Further, when the compression auto-ignition is performed in the subsequent cylinder in this way, the thermal efficiency of the subsequent cylinder is improved, and thus the torque difference between the preceding cylinder and the subsequent cylinder becomes large, but is different from the case where the compression self-ignition is not performed. By setting the ratio of the air-fuel ratio, that is, the ratio taking into account the improvement of the thermal efficiency of the subsequent cylinder, the generation of the generated torque difference between the front and rear cylinders can be satisfactorily eliminated as in the case where the compression self-ignition is not performed. Or it is suppressed.
[0021]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
[0022]
FIG. 1 shows a schematic configuration of an engine according to an embodiment of the present invention, and FIG. 2 schematically shows a structure of one cylinder of an engine main body and intake / exhaust valves provided for the cylinder. In these drawings, the engine body 1 has a plurality of cylinders, and in the illustrated embodiment, has four cylinders 2A to 2D. A piston 3 is fitted into each of the cylinders 2 </ b> A to 2 </ b> D, and a combustion chamber 4 is formed above the piston 3.
[0023]
A spark plug 7 is provided at the top of the combustion chamber 4 of each cylinder 2 </ b> A to 2 </ b> D, and the tip of the plug faces the combustion chamber 4. An ignition circuit 8 capable of controlling the ignition timing by electronic control is connected to the spark plug 7.
[0024]
A fuel injection valve 9 that directly injects fuel into the combustion chamber 4 is provided at a side portion of the combustion chamber 4. This fuel injection valve 9 incorporates a 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 timing to open the valve. An amount of fuel corresponding to the valve time is injected. The fuel injection valve 9 is supplied with fuel by a fuel pump (not shown) through a fuel supply passage and the like, and is supplied with fuel higher than the pressure in the combustion chamber during the compression stroke. A system is configured.
[0025]
Further, intake ports 11, 11a, 11b and exhaust ports 12, 12a, 12b are opened to the combustion chambers 4 of the respective cylinders 2A to 2D, and an intake passage 15 and an exhaust passage 20 are connected to these ports. Each port is opened and closed by intake valves 31, 31a, 31b and exhaust valves 32, 32a, 32b.
[0026]
The cylinders 2A to 2D perform a combustion cycle including intake, compression, expansion, and exhaust strokes with a predetermined phase difference. In the case of a four-cylinder engine, the first cylinder from one end in the cylinder row direction 2A, No. 2 cylinder 2B, No. 3 cylinder 2C, No. 4 cylinder 2D, as shown in FIG. 6, the cycle is in the order of No. 1 cylinder 2A, No. 3 cylinder 2C, No. 4 cylinder 2D, No. 2 cylinder 2B. The combustion cycle is performed with a phase difference of 180 ° in crank angle. In FIG. 6, EX represents an exhaust stroke, IN represents an intake stroke, F represents fuel injection, and S represents ignition.
[0027]
Between a pair of cylinders in which the exhaust stroke and the intake stroke overlap, a cylinder on the intake stroke side (this specification is referred to as a preceding cylinder) from the cylinder on the exhaust stroke side when the exhaust stroke and the intake stroke overlap (this specification is referred to as a preceding cylinder) The inter-cylinder gas passage 22 is provided so that the burned gas can be directly introduced to the subsequent cylinder). In 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 third Since the intake stroke (IN) of the cylinder 2C overlaps, the first cylinder 2A and the second cylinder 2B, and the fourth cylinder 2D and the third cylinder 2C form a pair, respectively, and the first cylinder 2A and the fourth cylinder 2D are the preceding cylinder. The second cylinder 2B and the third cylinder 2C are the subsequent cylinders.
[0028]
The intake / exhaust port of each cylinder and the intake passage, exhaust passage, and inter-cylinder gas passage connected to the cylinder are specifically configured as follows.
[0029]
The first cylinder 2A and the fourth cylinder 2D, which are the preceding cylinders, respectively include an intake port 11 for introducing fresh air, and a first exhaust port 12a for sending burned gas (exhaust gas) to the exhaust passage. A second exhaust port 12b for leading the burned gas to the subsequent cylinder is provided. The second cylinder 2B and the third cylinder 2C, which are the subsequent cylinders, respectively, have a first intake port 11a for introducing fresh air and a second intake port for introducing burned gas from the preceding cylinder. 11b and an exhaust port 12 for sending burned gas to the exhaust passage.
[0030]
In the example shown in FIG. 1, the intake ports 11 in the first and fourth cylinders 2A and 2D and the first intake ports 11a in the second and third cylinders 2B and 2C are two per cylinder, the left half of the combustion chamber. The first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders 2A and 2D and the second intake port 11b and the exhaust port in the second and third cylinders 2B and 2C are provided in parallel on the part side. 12 are provided in parallel on the right half side of the combustion chamber.
[0031]
The intake port 11 in the first and fourth cylinders 2A and 2D and the first intake port 11a in the second and third cylinders 2B and 2C are connected to the downstream ends of the branch intake passages 16 for each cylinder in the intake passage 15. Yes. In the vicinity of the downstream end of each branch intake passage 16, a multiple throttle valve 17 that is linked to each other via a common shaft is provided. This multiple throttle valve 17 is driven by an actuator 18 in accordance with a control signal, The intake air amount is adjusted. Note that an air flow sensor 19 that detects an intake air flow rate is provided in a common intake passage upstream of the collecting portion in the intake passage 15.
[0032]
An upstream end of a branch exhaust passage 21 for each cylinder in the exhaust passage 20 is connected to the first exhaust port 12a in the first and fourth cylinders 2A and 2D and the exhaust port 12 in the second and third cylinders 2B and 2C. Yes. Further, an inter-cylinder gas passage 22 is provided between the first cylinder 2A and the second cylinder 2B and between the third cylinder 2C and the fourth cylinder 2D, respectively, and the first and fourth cylinders which are the preceding cylinders. The upstream end of the inter-cylinder gas passage 22 is connected to the second exhaust ports 12b of 2A and 2D, and the inter-cylinder gas passage 22 is connected to the second intake ports 11b of the second and third cylinders 2B and 2C as the subsequent cylinders. The downstream end is connected.
[0033]
An exhaust gas concentration detection means for detecting the stoichiometric air-fuel ratio is provided at the downstream of the branch exhaust passage 21 in the exhaust passage 20. ₂ A sensor 23 is provided, and a three-way catalyst 24 for exhaust purification is provided in the exhaust passage 20 downstream thereof. As is generally known, the three-way catalyst 24 is highly purified against HC, CO, and NOx when the air-fuel ratio of the exhaust gas is close to the stoichiometric air-fuel ratio (that is, the excess air ratio λ is λ = 1). It is a catalyst showing performance. O ₂ The sensor 23 detects the air-fuel ratio by detecting the oxygen concentration in the exhaust gas, and particularly the λO whose output changes suddenly near the stoichiometric air-fuel ratio. ₂ It is composed of sensors.
[0034]
The inter-cylinder gas passage 22 has a linear O output whose output changes linearly with respect to a change in oxygen concentration in the exhaust gas (change in air-fuel ratio). ₂ A sensor 25 is provided.
[0035]
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. That is, the intake port 11, the first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders 2A and 2D are provided with the intake valve 31, the first exhaust valve 32a and the second exhaust valve 32b, respectively. A first intake valve 31a, a second intake valve 31b, and an exhaust valve 32 are provided in the first intake port 11a, the second intake port 11b, and the exhaust port 12 in the second and third cylinders 2B, 2C, respectively. These intake / exhaust valves are opened and closed at predetermined timings by the valve mechanisms comprising the camshafts 33, 34, etc. so that the intake stroke and exhaust stroke of each cylinder are performed with the predetermined phase difference as described above. To be driven.
[0036]
Further, among these intake / exhaust valves, the first exhaust valve 32a, the second exhaust valve 32b, the first intake valve 31a, and the second intake valve 31b are switched between an operating state and a stopped state. A valve stop mechanism 35 is provided. The valve stop mechanism 35 has been known in the art and will not be shown in detail. For example, hydraulic oil can be supplied to and discharged from a tappet interposed between the cams of the camshafts 33 and 34 and the valve shaft. When a hydraulic oil is supplied to the hydraulic chamber, the operation of the cam is transmitted to the valve and the valve is opened and closed. When the hydraulic oil is discharged from the hydraulic chamber, the cam operation is not performed. The valve is stopped by not being able to be transmitted to.
[0037]
The first control valve 37 and the second exhaust valve 32b are stopped in the hydraulic oil supply / discharge passage 36 to the valve stop mechanism 35 of the first exhaust valve 32a and the valve stop mechanism 35 of the first intake valve 31a. A second control valve 39 is provided in each of the hydraulic oil supply / discharge passages 38 to the mechanism 35 and the valve stop mechanism 35 of the second intake valve 31b (see FIG. 3).
[0038]
FIG. 3 shows the configuration of the engine drive and control system. In this figure, an ECU (control unit) 40 for engine control composed of a microcomputer or the like is provided with an air flow sensor 19, O ₂ Sensor 23 and linear O ₂ A signal from the sensor 25 is input, a signal from a water temperature sensor 51 for detecting the coolant temperature of the engine is input, and further, a rotation speed sensor 52 for detecting the engine rotation speed and an accelerator opening to determine the operating state. A signal from an accelerator opening sensor 53 or the like that detects the degree (depressed amount of the accelerator pedal) is also input. The ECU 40 outputs control signals to the fuel injection valves 9, the actuator 18 of the multiple throttle valve 17, and the first and second control valves 37 and 39.
[0039]
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, an ignition control unit 47, and the like as its functional configuration. I have.
[0040]
The operating state discriminating means 41 has a control map in which the engine operating region is divided into a region A on the low speed and low load side and a region B on the high speed side or the high load side as shown in FIG. The region A on the side is the special operation mode region, and the region B on the high speed side or high load side is the normal operation mode region. Then, it is determined whether the operating state of the engine (engine speed and engine load), which is examined from signals from the rotational speed sensor 52 and the accelerator opening sensor 53, is in the above regions A and B.
[0041]
The temperature state discriminating means 42 discriminates the temperature state of the engine based on a signal from the water temperature sensor 51, and discriminates whether the water temperature (engine temperature) is a low temperature below a predetermined value or a high temperature higher than a predetermined temperature. It has become.
[0042]
In the special operation mode region A, the mode setting means 43 introduces burnt gas discharged from the preceding cylinder in the exhaust stroke into the subsequent cylinder in the intake stroke as it is based on the determination by the operating state determination means 41 and burns it. In the normal operation mode region B, the normal operation mode in which each cylinder is burned independently is selected.
[0043]
In response to the mode setting by the mode setting means 43, the valve stop mechanism control means 44 follows the burned gas of the preceding cylinders (first and fourth cylinders) 2A and 2D via the inter-cylinder gas passage 22 in the special operation mode. The valve stop mechanism 35 is controlled to change the intake / exhaust flow state so that the cylinders (second and third cylinders) 2B and 2C are introduced into the cylinders independently. Each valve stop mechanism 35 is controlled as follows by controlling the control valves 37 and 39 depending on which of the regions A and B is present.
Area A: (Special operation mode)
Stop the first exhaust valve 32a and the first intake valve 31a
The second exhaust valve 32b and the second intake valve 31b are activated.
Area B: (Normal operation mode)
The first exhaust valve 32a and the first intake valve 31a are activated.
Stop the second exhaust valve 32b and the second intake valve 31b
[0044]
The intake air amount control means 45 controls the opening degree (throttle opening degree) of the throttle valve 17 by controlling the actuator 18, and obtains the target intake air amount from a map or the like according to the operating state, and the target The throttle opening is controlled according to the intake air amount. In this case, in the operation region A in the special operation mode, the excess air in the gas introduced from the preceding cylinders 2A and 2D and the intake air from the branch intake passage 16 to the succeeding cylinders 2B and 2C are blocked. Since the combustion is performed while the ratio of the newly supplied fuel to the lean air-fuel ratio is made, the amount of air necessary for the combustion of the fuel corresponding to the required torque for the preceding and subsequent two cylinders is reduced to the preceding cylinders 2A and 2D. The throttle opening is adjusted so that it is supplied to the engine.
[0045]
The fuel control means 46 controls the fuel injection amount and 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, and the ignition control means 47 corresponds to the operating state. The ignition timing is controlled. The combustion state control (fuel control and ignition control) is changed according to the mode set by the mode setting means 43.
[0046]
That is, when the special operation mode is set, the lean air-fuel ratio for the preceding cylinders 2A and 2D is larger than the stoichiometric air-fuel ratio, preferably approximately twice or more than the stoichiometric air-fuel ratio. In this way, the fuel injection amount is controlled, and the injection timing is set so that stratified combustion is performed by injecting fuel in the compression stroke, and the ignition timing is set so that forced ignition is performed near the compression top dead center. To do. On the other hand, for the succeeding cylinders 2B and 2C, the fuel injection amount is controlled so that the fuel is supplied to the burned gas having the lean air-fuel ratio introduced from the preceding cylinders 2A and 2D to obtain the stoichiometric air-fuel ratio. In addition, the injection timing is set so that ignition and combustion are possible in a situation where there is a lot of burned gas, for example, the injection timing is set so that fuel is injected in the compression stroke to ensure ignitability, and the compression top dead center The ignition timing is set so that forced ignition is performed at a predetermined timing in the vicinity.
[0047]
Further, 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 theoretical air-fuel ratio. The stoichiometric air-fuel ratio is set in the region, and the stoichiometric air-fuel ratio is made richer in the fully-open load and the operation region in the vicinity thereof. In this case, the injection timing is set so that fuel is injected into the cylinders 2A to 2D in the intake stroke to make the air-fuel mixture uniform, and each cylinder 2A to 2D is forcedly ignited. To.
[0048]
When the special operation mode is selected, the sum of the fuel injection amounts for both the pair of cylinders is an amount that becomes the stoichiometric air-fuel ratio with respect to the air amount introduced into the preceding cylinders 2A and 2D, and The ratio between the fuel injection amount for the preceding cylinders 2A and 2D and the fuel injection amount for the subsequent cylinders 2A and 2D is adjusted so that the generation torque of the preceding cylinders 2A and 2D and the generation torque of the subsequent cylinders 2B and 2C are equal. .
[0049]
In this regard, a more detailed configuration of the fuel control means 46 will be described with reference to FIG. That is, the fuel control means 46 has, as its functional configuration, a total fuel injection amount calculation unit 55, a final fuel injection amount calculation unit 56, a torque balance air / fuel ratio setting unit 57, a combustibility determination unit 58, and a division ratio as shown in FIG. A calculation unit 59 is provided.
[0050]
The total fuel injection amount calculation unit 55 obtains the fuel injection amount to be injected from the fuel injection valve 9 based on the intake air amount detected by the air flow sensor 19, and particularly when the special operation mode is selected. Then, the sum of the fuel injection amount for the preceding cylinder and the fuel injection amount for the subsequent cylinder (total fuel injection amount) is obtained. In this case, as described above, the fuel injection amount is determined so that the sum of the fuel injection amounts becomes the stoichiometric air-fuel ratio with respect to the air amount introduced into the preceding cylinders 2A and 2D.
[0051]
The final fuel injection amount calculation unit 56 determines the fuel injection amount to be finally controlled. When the normal operation mode is selected, the total fuel injection amount calculated by the total fuel injection amount calculation unit 55 is used as it is. Final fuel injection amount. On the other hand, when the special operation mode is selected, the fuel injection amounts for the preceding cylinder and the subsequent cylinder are calculated from the total fuel injection amount and a distribution ratio described later, and this is set as the final fuel injection amount.
[0052]
The torque balance air-fuel ratio setting unit 57, the combustibility determination unit 58, and the division ratio calculation unit 59 function when the special operation mode is selected. The torque balance air-fuel ratio setting unit 57 includes the rotation speed sensor 52 and the accelerator opening. The air-fuel ratio of the preceding cylinder is obtained from a map stored in advance according to the operating state of the engine (engine speed and engine load) checked from the signal from the degree sensor 53 and the like. The map, for example, experimentally obtains a difference in generated torque generated between the preceding cylinder and the succeeding cylinder due to a difference in thermal efficiency including pumping loss, and the design sky so that this torque difference can be “0”. Correct the fuel ratio (that is, a lean air-fuel ratio larger than the stoichiometric air-fuel ratio, preferably an air-fuel ratio that is approximately twice or more than the stoichiometric air-fuel ratio and is a value that is required in design), and change this to the engine operating state. It is a correspondence.
[0053]
The combustibility determination unit 58 determines whether or not combustion can be normally performed based on the ratio of the air / fuel ratio obtained by the torque balance air / fuel ratio setting unit 57 based on a signal from the water temperature sensor 51, a map, and the like. The result is determined in advance, and the result is output to the division ratio calculation unit 59.
[0054]
The division ratio calculation unit 59 determines the distribution ratio of the fuel (the total fuel injection amount) to the preceding and succeeding cylinders from the ratio of the air / fuel ratio of the preceding cylinder and the air / fuel ratio (theoretical air / fuel ratio) of the succeeding cylinder obtained from the map. However, when the determination result in the combustible determination unit 58 is affirmative, that is, when it is determined that normal combustion is performed, based on the air-fuel ratio ratio obtained by the torque balance air-fuel ratio setting unit 57. The fuel division ratio is obtained, and the result is output to the final fuel injection amount calculation unit 56. On the other hand, when the determination result in the flammable determination unit 58 is negative, that is, when there is a possibility that misfire or knocking may occur, it is set in advance within a range where normal combustion can be performed in the preceding and subsequent cylinders. For example, the distribution ratio when the air-fuel ratio of the preceding cylinder is set to a designed value (value before correction based on the torque difference) is output to the final fuel injection amount calculation unit 56. In this embodiment, the division ratio calculation unit 59 and the final fuel injection amount calculation unit 56 constitute the final fuel injection amount means of the present invention.
[0055]
Next, the operation of the apparatus according to the above embodiment will be described with reference to FIGS.
[0056]
In the operation region A on the low load and low rotation side, the special operation mode is set, and as described above, 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 are operated. As a result, as shown in FIG. 7, the substantial fresh air and gas flow paths are such that the burned gas discharged from the preceding cylinders 2A, 2D is directly passed through the inter-cylinder gas passage 22 and the subsequent cylinders 2B, 2C. And the two-cylinder connection state is established in which only the burned gas discharged from the succeeding cylinders 2B and 2C is guided to the exhaust passage 20 provided with the three-way catalyst 24.
[0057]
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. 7), and the lean air-fuel ratio in the preceding cylinders 2A and 2D is larger than the stoichiometric air-fuel ratio. Thus, fuel is injected in the compression stroke while the fuel injection amount is controlled, and ignition is performed at a predetermined ignition timing to perform stratified combustion at a lean air-fuel ratio (see FIG. 6).
[0058]
Thereafter, during a period in which the exhaust strokes of the preceding cylinders 2A and 2D overlap with the intake strokes of the succeeding cylinders 2B and 2C, the burned gas discharged from the preceding cylinders 2A and 2D passes through the inter-cylinder gas passage 22 and the succeeding cylinders 2B and 2C. (The white arrow in FIG. 6 and the arrow b in FIG. 7). In the succeeding cylinders 2B and 2C, fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinders 2A and 2D, and the fuel injection amount is controlled so that the stoichiometric air-fuel ratio is obtained. For example, fuel is injected in the compression stroke), and ignition is performed at a predetermined ignition timing to perform combustion (see FIG. 6). Then, the burnt gas after combustion in the succeeding cylinders 2B and 2C is discharged to the exhaust passage 20 provided with the three-way catalyst 24 (arrow c in FIG. 7).
[0059]
Thus, in the preceding cylinders 2A and 2D, stratified combustion is performed at a lean air-fuel ratio, so that the thermal efficiency is increased and the pumping loss is reduced by reducing the intake negative pressure compared to a normal engine that does not perform stratified combustion. The fuel efficiency is greatly improved by these synergistic effects. On the other hand, in the subsequent cylinders 2B and 2C, the combustion is performed while the air-fuel ratio is substantially the stoichiometric air-fuel ratio, so that the thermal efficiency is slightly higher than that in which the stratified combustion is performed at the lean air-fuel ratio as in the preceding cylinders 2A and 2D. Although inferior, since the burned gas pushed out from the preceding cylinders 2A and 2D is sent, the pumping loss is further reduced as compared with the preceding cylinders 2A and 2D, and thereby the fuel consumption improvement effect can be sufficiently obtained.
[0060]
In addition, the burned gas discharged from the succeeding cylinders 2B and 2C to the exhaust passage 20 has a value corresponding to the stoichiometric air-fuel ratio, so there is no need to provide a lean NOx catalyst as in a conventional lean burn engine, and a three-way catalyst. Exhaust purification performance is sufficiently ensured with only 24. And since there is no need to provide a lean NOx catalyst in this way, there is no need to temporarily enrich the air-fuel ratio for NOx release and reduction when the NOx occlusion amount of the lean NOx catalyst increases, improving fuel economy Can be avoided. Furthermore, the problem of sulfur poisoning of the lean NOx catalyst does not occur.
[0061]
The preceding cylinders 2A and 2D have a lean air-fuel ratio that is approximately twice or more than the stoichiometric air-fuel ratio, so that the amount of NOx generated is relatively small. In the succeeding cylinders 2B and 2C, the preceding cylinders 2A and 2D Since the burned gas is introduced, the state is equivalent to that in which a large amount of EGR is performed, so that the generation of NOx is sufficiently suppressed. This is also advantageous for improving emissions.
[0062]
Further, for the preceding cylinders 2A, 2D and the succeeding cylinders 2B, 2C, the fuel injection amount is controlled so that the generated torque between the cylinders becomes equal according to the operating state of the engine (total fuel is distributed). Thereby, NVH performance (noise vibration prevention performance) is improved satisfactorily. That is, since there is a difference in thermal efficiency including the pumping loss between the preceding cylinders 2A, 2D and the succeeding cylinders 2B, 2C, fuel is injected into the preceding cylinders 2A, 2D and the succeeding cylinders 2B, 2C. If the amount is evenly controlled, vibration and noise may be generated due to the difference in generated torque between the cylinders. However, according to the apparatus of the above embodiment, the fuel injection means controls the combustion injection amount as described above. Therefore, there is almost no torque difference between the preceding cylinders 2A, 2D and the succeeding cylinders 2B, 2C. Therefore, the occurrence of vibration or the like due to the torque difference is effectively prevented.
[0063]
In addition, in the process of setting the fuel injection amount (distribution ratio) in the fuel control means 46, it is determined whether or not combustion is normally performed in the preceding and subsequent cylinders (determination by the flammability determining unit 58). When there is a possibility that normal combustion may not be performed, the fuel injection amount is determined within a range in which normal combustion is possible. Therefore, there is an adverse effect of controlling the fuel injection amount so that the generated torque is uniform. For example, occurrence of misfire or knocking can be effectively prevented. In other words, if priority is given to equalizing the generated torque, the fuel injection amount (distribution ratio) may be set in such a range that the air-fuel ratio of the preceding cylinder or the succeeding cylinder exceeds the normal combustion range. May cause misfire or knocking. However, according to the apparatus of this embodiment, since the fuel injection amount is determined after the above determination is made in advance, the fuel injection amounts for the preceding and subsequent cylinders are always within a range where normal combustion is possible. Will be controlled. Therefore, it is possible to prevent the occurrence of misfire and knocking and ensure a normal operation state.
[0064]
On the other hand, in the operation region B on the high load side or high rotation side, the normal operation mode is set, and the first exhaust valve 32a and the first intake valve 31a are in the operating state as described above, and the second exhaust valve 32b and the second intake valve 31b. Is brought into a stopped state, the substantial new air and gas flow paths are as shown in FIG. 8, and the intake ports 31 and 31a and the exhaust ports 12a and 12 of each cylinder 2A to 2D are substantially independent. Then, fresh air is introduced from the intake passage 15 to the intake ports 31 and 31a of the respective cylinders 2A to 2D, and burned gas is discharged from the exhaust ports 31 and 31a of the respective cylinders 2A to 2D to the exhaust passage 20. In this case, the output air performance is ensured by controlling the intake air amount and the fuel injection amount so that the stoichiometric air-fuel ratio or richer (λ ≦ 1).
[0065]
In addition, the specific structure of the apparatus of this invention is not limited to the said embodiment, A various change is possible.
[0066]
For example, in the special operation mode, based on the determination of the temperature state of the engine, a compression ignition mode in which combustion in the subsequent cylinder is performed by forced ignition at a low temperature and a combustion in the subsequent cylinder is performed by compression self-ignition at a high temperature is performed. In the ignition mode, and particularly in the compression self-ignition mode, the fuel may be controlled to be injected in the intake stroke. That is, in the special operation mode, 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 take in the intake air. The temperature in the fuel chamber rises in the stroke, and the pressure and temperature rise in the compression stroke from this state. As a result, the temperature in the combustion chamber rises to such an extent that the air-fuel mixture can self-ignite near the top dead center at the end of the compression stroke. To rise. In addition, the burned gas is discharged from the preceding cylinders 2A and 2D and distributed uniformly, and the fuel injected in the intake stroke is also uniformly distributed throughout the combustion chamber until the end of the compression stroke. A uniform mixture distribution state that satisfies the simultaneous compression self-ignition condition is obtained. Therefore, at a high temperature, the combustion in the succeeding cylinder is set to the compression self-ignition mode and the compression self-ignition is performed, so that the combustion is rapidly performed, and thereby the thermal efficiency can be greatly improved.
[0067]
Even in such a compression self-ignition mode, the NVH performance (noise level) is controlled by controlling the fuel injection amount so that the generated torque between the cylinders is equal to the preceding cylinders 2A, 2D and the succeeding cylinders 2B, 2C. As a result, it is possible to ensure a good operating state. In this case, forced ignition is performed as a map for obtaining the ratio of the air-fuel ratio in the torque balance air-fuel ratio setting unit 57. It is necessary to store the compressed self-ignition mode separately from the mode and set the air-fuel ratio based on the map corresponding to the selected mode. That is, in compression self-ignition, heat loss is small, and the temperature in the cylinder does not easily rise due to rapid combustion, so that thermal efficiency is improved compared to forced ignition. Therefore, in the compression self-ignition mode, it is considered that the generated torque difference between the preceding cylinders 2A, 2D and the succeeding cylinders 2B, 2C is larger than that in the forced ignition mode. This is because it is difficult to completely eliminate the torque difference between the preceding and succeeding cylinders only by obtaining the ratio of the air-fuel ratio.
[0068]
As described above, the ignition mode is switched other than controlling the compression ignition mode at the high temperature and the forced ignition mode at the low temperature based on the determination of the temperature state of the engine in the entire special operation mode region A as described above. For example, the compression self-ignition mode and the forced ignition mode are switched only in the engine low speed range (low speed range in the region A in FIG. 5) according to the engine temperature state, and self ignition is performed in the special operation mode region A. In the high-speed, high-load region where the above is easily performed, the compression self-ignition mode may be set regardless of the engine temperature state.
[0069]
Further, in the above embodiment, the value of the map (map stored in the torque balance air / fuel ratio setting unit 57 for setting the ratio of the air / fuel ratio) serving as the basis for determining the fuel injection amount of the preceding / following cylinder is the preceding cylinder 2A. , 2D and the subsequent cylinders 2B, 2C are experimentally obtained, and the designed air-fuel ratio (a lean air-fuel ratio larger than the stoichiometric air-fuel ratio, preferably about the stoichiometric air-fuel ratio is set so that this torque difference can be “0”). This is a correction of the air-fuel ratio that is twice or more, and is a value that is required in design). Of course, theoretical (calculation) using design values from parameters related to pumping loss and thermal efficiency of the preceding and succeeding cylinders. You may use the value calculated | required by.
[0070]
【The invention's effect】
As described above, the control device according to the present invention has exhausted from the preceding cylinder in the exhaust stroke between the normal operation mode in which each cylinder performs combustion independently and the pair of cylinders in which the exhaust stroke and the intake stroke overlap. It is configured to be able to switch to a special operation mode in which the fuel gas is introduced into the subsequent cylinder in the intake stroke as it is to perform combustion. For example, in the low load low rotation operation region, the special operation mode is set, so that the preceding cylinder In this case, combustion at a lean air-fuel ratio is performed to improve the thermal efficiency and reduce the pumping loss, thereby obtaining a significant fuel efficiency improvement effect. In the succeeding cylinder, the lean air-fuel ratio already introduced from the preceding cylinder is obtained. Combustion is performed in a state where the fuel is supplied to the fuel gas and the stoichiometric air-fuel ratio is set, so that at least a fuel efficiency effect by reducing pumping loss can be obtained. . Further, since only the stoichiometric burned gas discharged from the succeeding cylinder is guided to the exhaust passage, the exhaust purification performance is sufficiently ensured only by the three-way catalyst. On the other hand, in the high load / high rotation operation region, the output performance is ensured by setting the normal operation mode. In the special operation mode, the fuel injection amount for each cylinder is determined and injected so that the generation torque of the preceding cylinder and the succeeding cylinder is equal according to the operating state of the engine. The generation torque difference between the front and rear cylinders due to the difference is eliminated or suppressed, and as a result, the NVH performance (noise vibration prevention performance) is enhanced.
[Brief description of the drawings]
FIG. 1 is a schematic plan view of an entire engine including a control device according to the present invention.
FIG. 2 is a schematic cross-sectional view of an engine body and the like.
FIG. 3 is a block diagram of a control system.
FIG. 4 is a block diagram showing a functional configuration of fuel control means.
FIG. 5 is an explanatory diagram showing an operation region.
FIG. 6 is a diagram illustrating 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 substantial fresh air and gas flow paths during low load and low rotation.
FIG. 8 is an explanatory diagram showing substantial fresh air and gas flow paths when in an operation region on a high load, high and low rotation side.
[Explanation of symbols]
1 Engine body
2A to 2D cylinder
9 Fuel injection valve
11 Intake port
11a First 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
24 Three-way catalyst
31 Intake valve
31a First intake valve
31b Second intake valve
32 Exhaust valve
32a First exhaust valve
32b Second exhaust valve
35 Valve stop mechanism
40 ECU
41 Operating state discriminating 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
55 Total fuel injection amount calculation unit (total fuel injection amount calculation means)
56 Final fuel injection amount calculation unit (final fuel injection amount calculation means)
57 Torque balance air-fuel ratio setting section (ratio setting means)
58 Flammability determination part (flammability determination means)
59 Division ratio calculation unit (final fuel injection amount calculation means)

Claims

Cylinder independent state in which fresh air is introduced into each cylinder and a two-cylinder connected state in which the burned gas of the preceding cylinder is introduced into the succeeding cylinder via the inter-cylinder gas passage between a pair of cylinders in which the exhaust stroke and the intake stroke overlap. In addition, the intake and exhaust flow paths are configured to be switchable, and the normal operation mode in which combustion is performed independently in each cylinder with the flow paths set to the cylinder independent state and the two cylinder connected state preceded. A control device for a multi-cylinder engine configured to be able to switch an operation mode to a special operation mode in which burned gas discharged from a cylinder is directly introduced into a subsequent cylinder in an intake stroke to perform combustion,
When in the special operation mode, based on the intake air amount to the preceding cylinder, combustion is performed in a state where the preceding cylinder has a lean air-fuel ratio that is larger than the theoretical air-fuel ratio by a predetermined amount, and in the succeeding cylinder, the preceding cylinder Total fuel injection amount calculation means for obtaining the sum of the fuel injection amounts for the preceding cylinder and the succeeding cylinder so that the fuel is supplied to the burned gas having a lean air-fuel ratio derived from the above and the combustion is performed in the state of the stoichiometric air-fuel ratio And ratio setting means for setting the ratio of the air-fuel ratio of the preceding cylinder to the succeeding cylinder in accordance with the operating state of the engine so that the generated torques of the preceding cylinder and the succeeding cylinder are equalized, and the ratio setting means A final fuel injection amount calculation unit for determining a final fuel injection amount for the preceding cylinder and the subsequent cylinder based on the ratio and the sum of the fuel injection amounts obtained by the total fuel injection amount calculation means. Control apparatus for an engine, characterized in that it comprises and.

The engine control device according to claim 1,
The air-fuel ratio ratio is determined based on a parameter related to pump loss or thermal efficiency in the preceding cylinder and the succeeding cylinder.

The engine control apparatus according to claim 1 or 2,
Combustion determining means for determining in advance whether or not the combustion state of the preceding cylinder and the succeeding cylinder based on the ratio of the air-fuel ratio determined by the ratio setting means is within a normal combustible range, and further comprising the final fuel injection amount The engine control apparatus according to claim 1, wherein the calculation means calculates the fuel injection amount based on the ratio set by the ratio setting means only when the determination result by the combustibility determination means is affirmative.

The engine control apparatus according to claim 3, wherein
When the determination result in the combustible determination means is negative, the final fuel injection amount calculation means determines the final fuel injection for each cylinder based on a ratio set in advance within a range where normal combustion is performed in each cylinder. An engine control device characterized by calculating a quantity.

The engine control apparatus according to any one of claims 1 to 4,
When in the special operation mode, it comprises ignition control means for selecting whether to perform combustion in the subsequent cylinder by compression self-ignition or forced ignition according to the operating state of the engine, and the ratio setting means, An engine control device configured to change a ratio of an air-fuel ratio to be set according to combustion by forced ignition or combustion by compression self-ignition.