JP2005016361A - Controller of spark ignition engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a controller of a spark ignition engine in which fuel consumption can be effectively improved with a simple construction while maintaining exhaust gas purification performance. <P>SOLUTION: In an operating range, a special operating mode for burning burned gas is performed from preceding cylinders into following cylinders 2B and 2C. The controller is provided with a valve opening and closing control means 43 for performing compression self ignition of the preceding cylinders 2A and 2D together with the following cylinders 2B and 2D by increasing the temperature in the preceding cylinders 2A and 2D, an ignition control means 46 for controlling to forcedly ignite the mixture gas in the preceding cylinders in the operating range in which it is hard to perform compression self ignition of the preceding cylinders 2a and 2D, and a fuel injection control means 45 for controlling the air-fuel ratio of the preceding cylinders 2A and 2d to rapidly change thereof into the lean side during shifting the operating range to perform compression self ignition of the preceding cylinders 2A and 2D into the operating range to perform forced ignition. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

上記吸入空気量制御手段４４は、アクチュエータ１８を制御することによってスロットル弁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、上記特殊運転モードとされる運転領域（部分負荷運転領域）Ａでは、先行気筒２Ａ，２Ｄの空燃比をリーン空燃比とするのに必要な量の空気が先行気筒２Ａ，２Ｄに供給されるとともに、後続気筒２Ｂ，２Ｃにおいて、分岐吸気通路１６から導入される新気と、先行気筒２Ａ，２Ｄから導入されるガス中の過剰空気と、燃料噴射弁９から新たに供給される燃料との比が理論空燃比となるようにスロットル開度が調節される。
【００５２】
上記燃料噴射制御手段４５は、各気筒２Ａ〜２Ｄに設けられた燃料噴射弁９からの燃料噴射量およ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ように構成されている。また、上記点火制御手段４６は、運転状態に応じた点火時期の制御および点火停止等の制御を行うように構成されている。そして、特に運転状態が図６中の部分負荷運転領域Ａにある場合と全運転領域Ｂにある場合とで燃焼状態の制御、つまり燃料噴射の制御および点火の制御状態が変更されるようになっている。
【００５３】
すなわち、特殊運転モードの制御が実行される上記部分負荷運転領域Ａ内において、その低負荷側領域Ａ１にある場合には、先行気筒２Ａ，２Ｄ内の空気過剰率λを３以上の大幅なリーン状態、例えば空燃比Ａ／Ｆを５０〜６０の範囲内とする制御が上記吸入空気量制御手段４４および燃料噴射制御手段４５からなる空燃比制御手段により実行されるとともに、図３に示すように、圧縮行程で燃料噴射弁９から先行気筒Ａ，２Ｄ内に対して燃料噴射Ｆを行う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Ｓを行って成層燃焼させる制御が、上記燃料噴射制御手段４５および点火制御手段４６からなる燃焼制御手段により実行されるように構成されている。
【００５４】
上記部分負荷運転領域Ａの中負荷側領域Ａ２では、先行気筒２Ａ，２Ｄ内の空気過剰率λを２〜３程度、例えば空燃比Ａ／Ｆを３５〜４０の範囲内とする制御が上記空燃比制御手段、つまり吸入空気量制御手段４４および燃料噴射制御手段４５のうち主として燃料噴射制御手段４５により実行されるとともに、図７に示すように、吸気行程で燃料噴射弁９から先行気筒２Ａ，２Ｄ内に燃料噴射Ｆを行うように噴射タイミングを設定し、かつ点火プラグ７に点火指令信号を出力して圧縮上死点前の上死点付近で混合気にアシスト点火Ｓを行うことにより、点火プラグ７周りの圧力を上昇させて先行気筒２Ａ，２Ｄの圧縮自己着火を促進する着火アシスト制御が上記点火制御手段４６からなる着火アシスト手段において実行されるようになっている。
【００５５】
また、上記着火アシスト手段により先行気筒２Ａ，２Ｄの圧縮自己着火を促進する部分負荷領域Ａの中負荷側領域Ａ２では、上記先行気筒２Ａ，２Ｄに設けられた第２排気弁３２ｂの閉弁時期ｔ１を、図７の破線で示す通常時の閉弁時期ｔ２に比べて進角させるように第２切換機構３５ｂを作動させることにより、先行気筒２Ａ，２Ｄの内部ＥＧＲ量を増大させて筒内温度を上昇させ、これによって先行気筒２Ａ，２Ｄを圧縮自己着火し易い状態とする制御が上記弁開閉制御手段４３からなる筒内温度上昇手段において実行されるように構成されている。
【００５６】
さらに、上記部分負荷運転領域Ａの高負荷側領域Ａ３では、先行気筒２Ａ，２Ｄ内の空気過剰率λを１〜３程度、例えば空燃比Ａ／Ｆを２０〜４０の範囲内とする制御が上記燃料噴射制御手段４５等からなる空燃比制御手段により実行されるとともに、上記点火制御手段４６からなる着火アシスト手段により先行気筒２Ａ，２Ｄの圧縮自己着火を促進する制御が実行され、かつ上記弁開閉制御手段４３からなる筒内温度上昇手段により先行気筒２Ａ，２Ｄ内の温度を上昇させる制御が必要に応じて実行されるようになっている。
【００５７】
先行気筒２Ａ，２Ｄの空燃比が運転領域に応じて上記のように設定されることにより、エンジンの運転領域が先行気筒２Ａ，２Ｄを圧縮自己着火させる上記中負荷側領域Ａ２から、先行気筒２Ａ，２Ｄを強制点火させる低負荷側領域Ａ１への移行時に、先行気筒２Ａ，２Ｄの空燃比をリーン側に急変させる制御が実行されることになる。
【００５８】
また、後続気筒２Ｂ，２Ｃに対しては、先行気筒２Ａ，２Ｄから導入された既燃ガス中の酸素濃度を考慮しつつ、分岐吸気通路１６から導入される新気に対して燃料を供給することにより、後続気筒２Ｂ，２Ｃの空燃比が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、吸気行程の前期またはそれ以前に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制御が上記燃焼状態制御手段により実行されるようになっている。
【００５９】
そして、エンジンの運転領域が先行気筒２Ａ，２Ｄを圧縮自己着火させる上記中負荷側領域Ａ２から、先行気筒２Ａ，２Ｄを強制点火させる低負荷側領域Ａ１への移行時には、先行気筒２Ａ，２Ｄの空燃比をＡ／Ｆ３５〜４０程度から５０〜６０程度に急変させて顕著なリーン状態とするのに伴い、後続気筒２Ｂ，２Ｃに対する燃料噴射量を増大させる制御が実行されるように構成されている。
【００６０】
一方、エンジンの運転状態が高負荷側ないし高回転側の全負荷運転領域Ｂにある場合には、通常運転モードの制御として、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、全開負荷およ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、原則として各気筒２Ａ〜２Ｄとも強制点火を行わせるように制御する。
【００６１】
上記構成において低負荷低回転側の部分運転領域Ａでは、上記特殊運転モードの制御が実行され、前述のように第１排気弁３２ａおよび第１吸気弁３１ａが停止状態、第２排気弁３２ｂおよび第２吸気弁３１ｂが作動状態とされることにより、実質的な新気およびガスの流通経路は図８に示すようになり、先行気筒（１番，４番気筒）２Ａ，２Ｄから排出される既燃ガスがそのまま気筒間ガス通路２２を介して後続気筒（２番，３番気筒）２Ｂ，２Ｃに導入されるとともに、この後続気筒２Ｂ，２Ｃから排出されるガスのみが排気通路２０に導かれるような２気筒接続状態とされる。
【００６２】
上記部分負荷運転領域Ａ１の低負荷側領域Ａ１では、図３中の矢印ａに示すように、先行気筒２Ａ，２Ｄの吸気行程で吸気通路１５から新気が導入されることにより、先行気筒２Ａ，２Ｄの空燃比Ａ／Ｆが、例えば５０〜６０程度の顕著なリーン状態となるように燃料噴射量が制御されつつ、吸気行程の前期またはそれ以前に燃料噴射Ｆが行われ、かつ、所定時期に強制点火Ｓが行われてリーン空燃比の燃焼が行われる。
【００６３】
また、上記部分負荷運転領域Ａ１の中負荷側領域Ａ２では、先行気筒２Ａ，２Ｄの空燃比Ａ／Ｆが、例えば３５〜４０程度のリーンとなるように燃料噴射量が制御されるとともに、図７に示すように、上記吸気行程の前期またはそれ以前に燃料噴射Ｆが行われる。そして、上記先行気筒２Ａ，２Ｄに設けられた第２排気弁３２ｂの閉弁時期ｔ１を、図７の破線で示す通常時の閉弁時期ｔ２に比べて進角させることにより、先行気筒２Ａ，２Ｄ内の温度を上昇させて圧縮自己着火が生じ易い状態とする制御が、上記弁開閉制御手段４３からなる筒内温度上昇手段において実行されるとともに、圧縮上死点前の上死点近傍でアシスト点火Ｓを行って圧縮自己着火を促進する制御が、上記点火制御手段４６からなる着火アシスト制御手段において実行されることにより、先行気筒２Ａ，２Ｄの圧縮自己着火が行われる。
【００６４】
さらに、上記部分負荷運転領域Ａ１の高負荷側領域Ａ３では、先行気筒２Ａ，２Ｄの空燃比Ａ／Ｆが、例えば２０〜４０程度のリーンとなるように燃料噴射量が制御されるとともに、必要に応じて上記筒内温度上昇手段による温度上昇制御および上記着火アシスト制御手段による圧縮自己着火の促進制御が実行されることにより、先行気筒２Ａ，２Ｄの圧縮自己着火が行われる。
【００６５】
そして、先行気筒２Ａ，２Ｄの吸気行程と後続気筒２Ｂ，２Ｃの排気行程とが重なる期間に、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に導入されるとともに（図３中の白抜き矢印および図８中の矢印ｂ）、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、Ｏ_２センサ２３の出力に基づいて燃料噴射量が制御されつつ、適当なタイミングで燃料が噴射されて燃焼が行われる。
【００６６】
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気筒間ガス通路２２を通って後続気筒２Ｂ，２Ｃ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、さらに吸気行程で燃料噴射弁９から吸気ポート１１ａに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の分布状態が得られることになる。
【００６７】
一方、高負荷側ないし高回転側の全負荷領域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および第１吸気弁３１ａが作動状態、第２排気弁３２ｂおよび第２吸気弁３１ｂが停止状態とされることにより、実質的な新気およびガスの流通経路は図９に示すような各気筒独立状態とされ、実質的に各気筒２Ａ〜２Ｄの吸気ポート３１，３１ａおよび排気ポート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３２，３２ａから排気通路２０に既燃ガスが排出される。そして、この場合には各気筒の空燃比が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および燃料噴射量が制御されることにより、エンジンの出力性能が確保されることになる。
【００６８】
上記のように先行気筒２Ａ，２Ｄでは、リーン空燃比での燃焼により熱効率が高められるとともに、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。
【００６９】
そして、上記特殊運転モードの制御が実行される部分負荷領域Ａの中負荷領域Ａ２では、先行気筒２Ａ，２Ｄ内の温度を上昇させることにより、先行気筒２Ａ，２Ｄを圧縮自己着火させるように構成したため、先行気筒２Ａ，２Ｄ内における燃焼を急速に行わせて熱効率を向上させることにより燃費を大幅に向上させることができるとともに、先行気筒２Ａ，２Ｄ内における酸素と窒素との反応を可急的に回避してＮＯｘの発生を充分に抑制することができる。
【００７０】
また、先行気筒２Ａ，２Ｄを圧縮自己着火させることが困難な部分負荷領域Ａの低負荷領域Ａ１では、先行気筒２Ａ，２Ｄ内の混合気を強制着火させるとともに、上記先行気筒２Ａ，２Ｄを圧縮自己着火させる運転領域（Ａ２）から先行気筒２Ａ，２Ｄを強制点火する運転領域（Ａ１）への移行時に、先行気筒２Ａ，２Ｄの空燃比をリーン側に急変させるように構成したため、圧縮自己着火を行うことなく、先行気筒２Ａ，２ＤにおけるＮＯｘの発生量を効果的に低減することができる。
【００７１】
すなわち、混合気の空燃比Ａ／Ｆに対応したＮＯｘの発生量は、図１０に示すように、混合気の空燃比Ａ／Ｆが理論空燃比１４．７（λ＝１）となる点よりもややリーンなときに最大となり、この最大点を超えて上記空燃比Ａ／Ｆがリーン側となるのに応じて顕著に減少する傾向があるため、先行気筒２Ａ，２Ｄの空燃比を顕著なリーン状態に設定することにより、ＮＯｘの発生を効果的に抑制することができる。しかも、上記運転領域の移行時に、先行気筒２Ａ，２Ｄ内の混合気を強制点火するように構成したため、制御の応答遅れ等により先行気筒２Ａ，２Ｄ内に所定量の既燃ガス（内部ＥＧＲによる既燃ガス）が存在する状況下においても失火を生じることなく、先行気筒２Ａ，２Ｃの混合気を確実に燃焼させることができる。
【００７２】
さらに、上記後続気筒２Ｂ，２Ｃでは、先行気筒２Ａ，２Ｄからの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、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等の手段を講じることなく、同時圧縮自己着火による急速燃焼が行われるため、可及的に酸素と窒素との反応を避けられることにより、ＮＯｘの発生を充分に抑制することができ、このような点からもエミッションの向上に有利となる。
【００７３】
したがって、上記実施形態に示すように、先行気筒２Ａ，２Ｄを圧縮自己着火させる運転領域Ａ２では、先行気筒２Ａ，２Ｄの空気過剰率を２〜３の範囲内とすることにより、先行気筒２Ａ，２Ｄに噴射される燃料の量を所定値に確保して、先行気筒２Ａ，２Ｄの圧縮自己着火を適正に行わせることができるとともに、先行気筒２Ａ，２Ｄを強制点火させる上記運転領域Ａ１では、先行気筒２Ａ，２Ｄの空気過剰率を３以上とするように空燃比を制御することにより、上記運転領域Ａ１におけるＮＯｘの発生を効果的に抑制しつつ、顕著な燃費の改善効果が得られるという利点がある。
【００７４】
また、上記実施形態では、先行気筒の温度を上昇させることにより先行気筒２Ａ，２Ｄを圧縮自己着火させる運転領域Ａ２よりも高負荷側領域Ａ３にある場合に、先行気筒２Ａ，２Ｄの空気過剰率を２〜３の範囲内とするように空燃比を制御するように構成したため、後続気筒２Ｂ，２Ｃの気筒内温度が高くなる傾向がある上記高負荷側領域Ａ３で、先行気筒２Ａ，２Ｃから排出された多量の既燃ガスを後続気筒２Ｂ，２Ｃに導入させることにより、この後続気筒２Ａ，２ＣにおけるＮＯｘの発生を効果的に抑制することができる。
【００７５】
また、上記実施形態に示すように、圧縮上死点前の上死点近傍で先行気筒２Ａ，２Ｃ内の混合気を点火して先行気筒２Ａ，２Ｄの圧縮自己着火を促進する点火制御手段４６からなる着火アシスト手段を設け場合には、上記特殊運転モードの制御が実行される部分負荷領域Ａの中負荷領域Ａ２等において、上記弁開閉制御手段４３からなる筒内温度上昇手段により先行気筒２Ａ，２Ｄの筒内温度を上昇させた状態で、この先行気筒２Ａ，２Ｄ内の混合気を点火して先行気筒２Ａ，２Ｄ内の圧力を瞬時に高めることにより、先行気筒２Ａ，２上記を適正時期に確実に圧縮自己着火させることできるという利点がある。
【００７６】
なお、上記特殊運転モードの制御が実行される上記部分負荷領域Ａにおいて、後続気筒２Ｂ，２Ｃを圧縮自己着火させることが困難な場合、例えば先行気筒２Ａ，２Ｄから導出される既燃ガスの温度が低い低負荷側領域Ａ１の運転状態にある場合等に、上記点火制御手段４６からなる着火アシスト手段２より後続気筒２Ｂ，２Ｃの混合気を、圧縮上死点前の上死点近傍で点火して後続気筒２Ｂ，２Ｃ内の圧力を瞬時に高めることにより、後続気筒２Ｂ，２Ｃを適正時期に確実に圧縮自己着火させるようにしてもよい。
【００７７】
上記実施形態では、先行気筒２Ａ，２Ｄを圧縮自己着火させる運転領域Ａ２から先行気筒２Ａ，２Ｄを強制点火させる運転領域Ａ１への移行時に、後続気筒２Ｂ，２Ｃに対する燃料噴射量を増大させる制御を燃料噴射制御手段４５により実行するように構成したため、上記運転領域Ａ１への移行時に、後続気筒２Ｂ，２Ｃに対する燃料噴射量を充分な値に設定することにより、上記のように先行気筒２Ａ，２Ｄの空燃比が顕著なリーン状態に設定されることによるエンジンの出力低下を後続気筒２Ｂ，２Ｃの燃焼エネルギーにより補うことができるという利点がある。
【００７８】
また、上記実施形態に示すように、弁開閉制御手段４３により先行気筒２Ａ，２Ｄの第２排気弁３２ｂを早閉じして先行気筒２Ａ，２Ｄ内の温度を上昇させるように構成した場合には、特殊運転モードの燃焼が行われる部分負荷領域Ａの特定運転領域Ａ２で、先行気筒２Ｂ，２Ｄの内部ＥＧＲ量を増大させることにより筒内温度を上昇させることができるため、先行気筒２Ａ，２Ｄを簡単かつ効果的に圧縮自己着火させることができる。
【００７９】
また、上記実施形態では、特殊運転モードにおいて後続気筒２Ｂ，２Ｃの空燃比を略理論空燃比とし、この理論空燃比で燃焼した排気ガスのみを後続気筒２Ｂ，２Ｃから排気通路２０に排出させるように構成したため、リーンＮＯｘ触媒を排気通路２０に設けることなく、三元触媒２４だけで充分に排気ガスの浄化性能を確保することができる。しかも、上記リーン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リッチ化を行う必要がなく、燃費改善効果の目減りを防止できるとともに、リーンＮＯｘ触媒が硫黄被毒するという問題が生じるのを防止できるという利点がある。
【００８０】
なお、特殊運転モードにおいて後続気筒２Ｂ，２Ｃの空燃比を略理論空燃比とし、この理論空燃比で燃焼した排気ガスのみを後続気筒２Ｂ，２Ｃから排気通路２０に排出させるように構成してなる上記実施形態に替え、図１１に示すように、排気通路２０に三元触媒２４と、排ガス中のＮＯｘを浄化するＮＯｘ浄化触媒２６とを設けるとともに、先行気筒２Ａ，２Ｄおよび後続気筒２Ｂ，２Ｃを圧縮自己着火させる運転領域Ａ２では、後続気筒２Ｂ，２Ｃの空燃比をややリーンに設定するように構成してもよい。
【００８１】
上記ＮＯｘ浄化触媒２６は、空燃比が理論空燃比よりも大きいリーン状態で排気ガス中のＮＯｘを吸着し、かつリッチ状態で吸着したＮＯｘを放出して還元浄化するものであり、例えば担体の壁面にアルミナやセリアがサポート材として担持された触媒層を有し、このサポート材に白金Ｐｔ、ロジウムＲｈまたはパラジウムＰｄ等の貴金属と、カリウムＫ等のアルカリ金属やバリウムＢａ等のアルカリ土類金属とが担持された１コートタイプのものが用いられる。上記リーンＮＯｘ触媒２６として、担体の壁面に白金Ｐｔと、バリウムＢａ等のアルカリ土類金属とが担持されたアルミナや、セリアを有する内側触媒と、白金Ｐｔ等の貴金属が担持されたゼオライトを有する外側触媒とからなる２コートタイプのものを用いてもよい。
【００８２】
上記のように構成した場合には、先行気筒２Ａ，２Ｄおよび後続気筒２Ｂ，２Ｃを圧縮自己着火させることによりＮＯｘの発生量を効果的に低減することができるため、排気通路２０に設けられたＮＯｘ吸着触媒２６の容量を大きな値に設定することなく、排気浄化性能を維持することができるとともに、後続気筒２Ｂ，２Ｃの空燃比をややリーンに設定することにより、エンジンの燃費をさらに効果的に改善できるという利点がある。しかも、上記ＮＯｘ吸着触媒２６に吸着されたＮＯｘの放出、還元を行う際には、後続気筒２Ｂ，２Ｃの空燃比をややリッチに設定することにより、ＮＯｘ吸着触媒２５に吸着されたＮＯｘを放出させるとともに、排気ガス中のＣＯまたはＣＨ等からなる還元剤によりＮＯｘを還元することができるため、燃費を顕著に低下させることなく、排気ガスの浄化性能を維持できるという利点がある。
【００８３】
また、上記実施形態では先行気筒２Ａ，２Ｄ、後続気筒２Ｂ，２Ｃのいずれに対しても燃料噴射弁９は燃焼室に直接燃料を噴射する直噴タイプとしているが、後続気筒２Ｂ，２Ｃに対する燃料噴射弁は必ずしも直噴タイプに限定されず、例えば吸気ポートおよび気筒間ガス通路に燃料噴射弁を設け、通常運転モードでは吸気ポートの燃料噴射弁を駆動し、特殊運転モードでは気筒間ガス通路の燃料噴射弁を駆動するようにしてもよい。
【００８４】
さらに、上記吸・排気気弁を開閉駆動するソレノイドアクチュエータを備えた電磁動弁機構を設け、この電磁動弁機構の作動状態を制御することにより、上記２気筒接続状態と各気筒独立状態とに切り換えるように構成してもよく、あるいは先行気筒２Ａ，２Ｄに設けられた第２排気弁３２Ｂの開弁期間を進角させることにより、先行気筒２Ａ，２Ｄの内部ＥＧＲ量を増大させて筒内温度を上昇させるようにしてもよい。
【００８５】
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。
【００８６】
【発明の効果】
以上のように本発明は、各気筒内に燃料を供給する燃料噴射弁が吸気導入経路に設けられるとともに、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運転領域で、排気行程と吸気行程と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既燃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、上記特殊運転モードの制御が実行される運転領域の少なくとも一部で先行気筒内の温度を上昇させることにより後続気筒とともに先行気筒を圧縮自己着火させる筒内温度上昇手段と、この先行気筒を圧縮自己着火させることが困難な運転領域では先行気筒内の混合気を強制着火させるように制御する点火制御手段と、上記先行気筒を圧縮自己着火させる運転領域から先行気筒を強制点火させる運転領域への移行時に、先行気筒の空燃比をリーン側に急変させるように制御する空燃比制御手段とを設けたため、上記特殊運転モードとして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およびポンピングロス低減による燃費改善効果が得られるとともに、後続気筒ではポンピングロス低減による燃費改善効果が得られるという利点がある。
【００８７】
また、特殊運転モードの燃焼が行われる特定運転領域の少なくとも一部で先行気筒の筒内温度を上昇させて先行気筒および後続気筒の両方を圧縮自己着火させる制御を上記筒内温度上昇手段において実行することにより、ＮＯｘの発生を効果的に抑制することができるとともに、顕著な燃費の改善効果を得ることができる。そして、先行気筒を圧縮自己着火させる運転領域から先行気筒を強制点火させる運転領域への移行時には、先行気筒の空燃比をリーン側に急変させる制御を実行することにより、先行気筒の燃焼時に発生するＮＯｘ量を顕著に低減できるという利点がある。
【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】先行気筒および後続気筒の燃焼サイクルおよび開弁タイミング等を示す説明図である。
【図４】切換手段の具体的構成を示す斜視図である。
【図５】制御系統のブロック図である。
【図６】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７】各気筒の排気行程、吸気行程、燃料噴射時期および点火時期等を示す図である。
【図８】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９】高負荷、高低回転側の運転領域にある時の実質的な新気およびガスの流通経路を示す説明図である。
【図１０】空燃比とＮＯｘ発生量との対応関係を示すグラフである。
【図１１】本発明の別の実施形態に係るエンジン全体を示す概略平面図である。
【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
４３ 弁開閉制御手段（運転モード制御手段）
４４ 吸入空気量制御手段（空燃比制御手段）
４５ 燃料噴射制御手段（空燃比制御手段）
５５ 点火制御手段[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a control device for a spark ignition engine, and more particularly to a control device for controlling the combustion state of each cylinder in order to improve fuel consumption and emissions in a multi-cylinder engine.
[0002]
[Prior art]
2. Description of the Related Art Conventionally, in a spark ignition engine, a technique for improving fuel consumption 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. A fuel injection valve that directly injects fuel into the room, and in the low-revolution and low-load operation region, etc., fuel is injected from the fuel injection valve in the compression stroke, and stratified combustion is performed. The thing which implement | achieved combustion is known (for example, refer patent document 1).
[0003]
In such an engine, an ordinary three-way catalyst (a catalyst having a high purification performance near the theoretical air-fuel ratio with respect to HC, CO, and NOx) as an exhaust gas purification catalyst is sufficient for NOx during lean operation. Since the purification performance cannot be obtained, a lean NOx catalyst is provided that adsorbs NOx in an oxygen-excessive atmosphere and releases and reduces NOx in an oxygen-concentrated atmosphere as shown in Patent Document 1 below. 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 Patent Document 1, additional fuel is burned during the expansion stroke in addition to the main combustion. It is injected into the room to enrich the air-fuel ratio of the exhaust gas and generate CO or the like, thereby promoting NOx removal and reduction.
[0004]
[Patent Document 1]
Japanese Patent Laid-Open No. 10-29836
[0005]
[Problems to be solved by the invention]
In an engine that performs the conventional lean operation as described above, it is necessary to provide a large and expensive lean NOx catalyst in the exhaust passage in order to ensure NOx purification performance during the lean operation, which is disadvantageous in terms of cost. Further, in order to maintain the purification performance of the lean NOx catalyst, as described above, when NOx adsorption amount increases, NOx is separated from the lean NOx catalyst and effectively reduced. It is necessary to frequently enrich the air-fuel ratio. Furthermore, when the fuel used contains a large amount of sulfur, in order to eliminate sulfur poisoning of the lean NOx catalyst, the regeneration process such as the heating process of the catalyst and the supply of the reducing material must be performed over a predetermined period. Therefore, the fuel efficiency improvement effect is reduced by these. Moreover, if the air-fuel ratio of the air-fuel mixture becomes leaner than a certain level, the combustion speed becomes too slow, and combustion close to the end does not contribute to work, so there is a limit to improving fuel efficiency by leaning in stratified combustion. It was.
[0006]
As another method for improving fuel efficiency, compression self-ignition has been studied, and this compression self-ignition, like a diesel engine, causes the combustion chamber to self-ignite at a high temperature and high pressure at the end of the compression stroke. Even in a lean state where the air-fuel ratio is remarkable or a large amount of EGR is introduced, if such compression self-ignition is performed, the entire combustion chamber burns at once, so slow combustion that does not contribute to work Is avoided, which is advantageous for improving fuel economy.
[0007]
However, in a normal spark-ignition 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 such an extent that compression self-ignition occurs. However, in order to perform compression self-ignition, special measures are required to significantly increase the temperature or pressure in the combustion chamber. Conventionally, knocking in a high load region, that is, before the flame propagates in the combustion chamber, 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 operation region where fuel consumption improvement is required, while avoiding abnormal combustion due to spontaneous ignition.
[0008]
Therefore, the present applicant introduces the burned gas of the preceding cylinder in the exhaust stroke between the pair of cylinders in which the exhaust stroke and the intake stroke overlap in the low and high rotation region of the engine as it is to the subsequent cylinder in the intake stroke. The cylinder is connected and burned with the air-fuel ratio of the preceding cylinder set to be leaner than the stoichiometric air-fuel ratio. In the succeeding cylinder, fuel is supplied to the burned gas of the lean air-fuel ratio discharged from the preceding cylinder to generate the stoichiometric air-fuel ratio. Control of a spark ignition engine that improves the exhaust gas purification performance without the need for a lean NOx catalyst, while improving the fuel efficiency by allowing combustion by compression self-ignition in a state of A technology related to the device has been filed (Japanese Patent Application No. 2002-024548).
[0009]
The present invention provides a control device for a spark ignition engine that can improve fuel efficiency more effectively with a simple configuration while ensuring exhaust purification performance based on such a technique.
[0010]
[Means for Solving the Problems]
The invention according to claim 1 is a multi-cylinder spark ignition in which a fuel injection valve for supplying fuel into each cylinder is provided in the intake air introduction path, and the combustion cycle of each cylinder is set to have a predetermined phase difference. In the engine partial load operation region, the burned gas discharged from the preceding cylinder in the exhaust stroke is directly in the intake stroke between a pair of cylinders in which the exhaust stroke and the intake stroke overlap. A lean state in which the air-fuel ratio of the preceding cylinder is larger than the stoichiometric air-fuel ratio while the two-cylinder connection state is established in which the burned gas discharged from the succeeding cylinder is introduced into the cylinder through the inter-cylinder gas passage and led to the exhaust passage. Combustion is performed as an air-fuel ratio, and control of a special operation mode is performed in which fuel is supplied to the burned gas having a lean air-fuel ratio introduced from the preceding cylinder to the succeeding cylinder and combustion of the succeeding cylinder is performed. Rotation mode control means, and in-cylinder temperature increasing means for compressing and igniting the preceding cylinder together with the succeeding cylinder by increasing the temperature in the preceding cylinder in at least a part of the operation region in which the control of the special operation mode is executed. In the operation region where it is difficult to perform compression self-ignition of the preceding cylinder, ignition control means for controlling the air-fuel mixture in the preceding cylinder to be forcedly ignited, and forcibly igniting the preceding cylinder from the operation region in which the preceding cylinder is subjected to compression self-ignition And an air-fuel ratio control means for controlling the air-fuel ratio of the preceding cylinder to suddenly change to the lean side at the time of transition to the operating range.
[0011]
According to the present invention, when combustion is performed in the special operation mode, the preceding cylinder can achieve the fuel efficiency improvement effect by the lean combustion and the fuel efficiency improvement by the pumping loss reduction, and the subsequent cylinder can obtain the fuel efficiency improvement effect by the pumping loss reduction. It is done. Further, in the in-cylinder temperature increasing means, the in-cylinder temperature increasing means performs compression self-ignition of both the preceding cylinder and the succeeding cylinder by increasing the in-cylinder temperature of the preceding cylinder in at least a part of the specific operation region in which the combustion in the special operation mode is performed. As a result, the generation of NOx is effectively suppressed, and a remarkable fuel economy improvement effect is obtained.
Then, at the time of transition from the operation region in which the preceding cylinder is subjected to compression self-ignition to the operation region in which the preceding cylinder is forcibly ignited, the control is executed to suddenly change the air-fuel ratio of the preceding cylinder to the lean side, thereby generating when the preceding cylinder burns The amount of NOx to be effectively reduced.
[0012]
Further, the invention according to claim 2 is the spark ignition engine control device according to claim 1, wherein the excess air ratio of the preceding cylinder is in a range of 2 to 3 in the operation region in which the preceding cylinder is compressed and self-ignited, In the operation region in which the preceding cylinder is forcibly ignited, the air-fuel ratio is controlled so that the excess air ratio of the preceding cylinder is 3 or more.
[0013]
According to the above configuration, in the operation region in which the in-cylinder temperature of the preceding cylinder is raised and both the preceding cylinder and the succeeding cylinder are compressed and self-ignited, the fuel injection amount to the preceding cylinder is secured at a predetermined value, In the region where the compression self-ignition of the cylinder is properly performed and the preceding cylinder is forcibly ignited, the air-fuel ratio of the preceding cylinder is set to a remarkable lean state, thereby effectively suppressing the generation of NOx, A remarkable fuel economy improvement effect can be obtained.
[0014]
Further, the invention according to claim 3 is a control device for the spark ignition engine according to claim 1 or 2, wherein the load is higher than that in an operating region in which the preceding cylinder is compressed and self-ignited by increasing the temperature of the preceding cylinder. In the side region, the air-fuel ratio is controlled so that the excess air ratio of the preceding cylinder is in the range of 2-3.
[0015]
According to the above configuration, the excess air ratio of the preceding cylinder is set in the range of 2 to 3 in the region of higher load than the operation region in which the preceding cylinder is compressed and self-ignited, and a large amount of burned gas is introduced into the succeeding cylinder. As a result, the generation of NOx in the succeeding cylinder is effectively suppressed.
[0016]
According to a fourth aspect of the present invention, in the spark ignition engine control device according to any one of the first to third aspects, the air-fuel mixture in the preceding cylinder is near top dead center before compression top dead center. Is provided with ignition assist means for accelerating compression self-ignition of the preceding cylinder.
[0017]
According to the above configuration, in a state in which the in-cylinder temperature increasing means increases the in-cylinder temperature of the preceding cylinder, the air-fuel mixture in the preceding cylinder is ignited and the pressure in the preceding cylinder is instantaneously increased, so that the appropriate time is reached. In addition, it is possible to reliably perform compression self-ignition in the preceding cylinder.
[0018]
Further, the invention according to claim 5 is the operation for forcibly igniting the preceding cylinder from the operation region in which the preceding cylinder is compressed and self-ignited in the control apparatus for the spark ignition engine according to any one of claims 1 to 4. Fuel injection control means is provided for increasing the fuel injection amount for the subsequent cylinders when shifting to the region.
[0019]
According to the above configuration, the fuel injection amount for the subsequent cylinder is set to a sufficient value at the time of transition from the operation region in which the preceding cylinder performs compression self-ignition to the operation region in which the preceding cylinder is forcibly ignited. The reduction in engine output due to the lean state of the fuel ratio being compensated by the combustion energy of the subsequent cylinder.
[0020]
Further, the invention according to claim 6 is the spark ignition engine control device according to any one of claims 1 to 5, wherein the exhaust valve of the preceding cylinder is quickly closed by the valve opening / closing control means. It is configured to increase the temperature inside.
[0021]
According to the above configuration, in the specific operation area where the combustion in the special operation mode is performed, the exhaust valve of the preceding cylinder is quickly closed to raise the in-cylinder temperature, thereby effectively compressing and auto-igniting the preceding cylinder. It becomes possible.
[0022]
Further, the invention according to claim 7 is the spark ignition engine control device according to any one of claims 1 to 6, wherein in the operation region in which the preceding cylinder and the succeeding cylinder are subjected to compression self-ignition, The air-fuel ratio is set slightly lean.
[0023]
According to the above configuration, the amount of NOx generated can be effectively reduced by compressing and auto-igniting the preceding cylinder and the succeeding cylinder. Therefore, even when the air-fuel ratio of the succeeding cylinder is set to be slightly lean, Thus, it is possible to maintain the exhaust purification performance without providing the NOx adsorption catalyst or frequently separating and reducing the NOx adsorbed on the NOx adsorption catalyst.
[0024]
DETAILED DESCRIPTION OF THE INVENTION
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 body 1 and intake / exhaust valves provided for the cylinder. In these drawings, the engine body 1 has a plurality of cylinders, and in the illustrated embodiment, has four cylinders 2A to 2D. A piston 3 is fitted into each of the cylinders 2 </ b> A to 2 </ b> D, and a combustion chamber 4 is formed above the piston 3.
[0025]
A spark plug 7 is provided at the top of the combustion chamber 4 provided in each of the cylinders 2 </ b> A to 2 </ b> D, and the plug tip 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.
[0026]
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. The ports are opened and closed by intake valves 31, 31a, 31b and exhaust valves 32, 32a, 32b.
[0027]
A combustion cycle including 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, from one end side in the cylinder row direction When referred to as the first cylinder 2A, the second cylinder 2B, the third cylinder 2C and the fourth cylinder 2D, as shown in FIG. 3, the combustion cycle is the first cylinder 2A, the third cylinder 2C, the fourth cylinder 2D, 2 The operation is performed with a phase difference of 180 ° in crank angle in the order of the numbered cylinder 2B. In FIG. 3, EX is an exhaust stroke, IN is an intake stroke, F is fuel injection, S is forced ignition, and a star mark indicates that compression self-ignition is performed.
[0028]
Between a pair of cylinders in which the exhaust stroke and the intake stroke overlap, a cylinder on the intake stroke side (referred to as a preceding cylinder in this specification) from the cylinder on the exhaust stroke side when the exhaust stroke and the intake stroke overlap (this cylinder is referred to as the preceding cylinder). The inter-cylinder gas passage 22 is provided so that the burned gas can be guided as it is to the subsequent cylinder in the specification). In the four-cylinder engine of this embodiment, as shown in FIG. 3, the exhaust stroke (EX) of the first cylinder 2A and the intake stroke (IN) of the second cylinder 2B overlap, and the exhaust stroke (EX) of the fourth cylinder 2D. ) And the intake stroke (IN) of the third cylinder 2C overlap, so that the first cylinder 2A and the second cylinder 2B and the fourth cylinder 2D and the third cylinder 2C form a pair, respectively. The second cylinder 2D becomes the preceding cylinder, and the second cylinder 2B and the third cylinder 2C become the subsequent cylinders.
[0029]
The intake / exhaust ports of each of the cylinders 2A to 2D, the intake passage 15, the exhaust passage 20, and the inter-cylinder gas passage 22 connected thereto are specifically configured as follows. That is, the first cylinder 2A and the fourth cylinder 2D, which are the preceding cylinders, respectively, have an intake port 11 for introducing fresh air and a first exhaust port for sending burned gas (exhaust gas) to the exhaust passage 20. 12a and a second exhaust port 12b for leading the burned gas to the succeeding cylinder are 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 11b for introducing burned gas from the preceding cylinder. And an exhaust port 12 for sending burned gas to the exhaust passage 20.
[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 are 2 per cylinder. The first exhaust port 12a and the second exhaust port 12b and the second and third cylinders in the first and fourth cylinders 2A and 2D (preceding cylinders) are provided in parallel on the left half side of the combustion chamber. (Subsequent 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 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, and this multiple throttle valve 17 is driven by an actuator 18 in accordance with a control signal. Thus, the intake air amount is adjusted. Note that an air flow sensor 19 for detecting the intake air flow rate is provided in the common intake passage upstream of the collecting portion in the intake passage 15.
[0032]
In addition, a fuel injection valve 9 for injecting fuel to the merging portion of each port is provided in the intake air introduction path including the intake port 11 and the first intake port 11a. This fuel injection valve 9 incorporates a needle valve and a solenoid (not shown). When a pulse signal is input from a fuel injection control means, which will be described later, the fuel injection valve 9 is driven and opened for a time corresponding to the pulse width at the pulse input timing. The fuel is injected and an amount of fuel corresponding to the valve opening time is injected. The fuel injection valve 9 is configured to be supplied with fuel at a predetermined pressure via a fuel pump and a fuel supply passage (not shown).
[0033]
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. 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 as the preceding cylinders, and the second and third cylinders 2B as the subsequent cylinders. The downstream end of the inter-cylinder gas passage 22 is connected to the 2C second intake port 11b.
[0034]
The inter-cylinder gas passage 22 is a relatively short passage that connects adjacent cylinders so that the amount of heat released while the gas discharged from the preceding cylinders 2A and 2D passes through the passage 22 can be kept relatively small. It has become. The inter-cylinder gas passage 22 has a linear O output whose output changes linearly in response to a change in oxygen concentration (change in air-fuel ratio) in the exhaust gas. ₂ A sensor 25 is provided.
[0035]
In the exhaust passage 20, downstream of the branch exhaust passage 21, an air-fuel ratio is detected by detecting the oxygen concentration in the exhaust gas. ₂ A sensor 23 is provided.
In addition, 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 portion 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 the catalyst shown.
[0036]
The intake / exhaust valves for opening and closing the intake / exhaust ports of the cylinders 2A to 2D and the valve operating mechanisms for these valves are as follows.
[0037]
The intake port 11, the first exhaust port 12a and the second exhaust port 12b in the first and fourth cylinders (preceding cylinders) 2A and 2D are respectively provided with an intake valve 31, a first exhaust valve 32a and a second exhaust valve 32b. The first intake port 11a, the second intake port 11b and the exhaust port 12 in the second and third cylinders (subsequent cylinders) 2B and 2C are respectively connected to the first intake valve 31a, the second intake valve 31b and the exhaust valve 32. Is provided. These intake / exhaust valves are each provided at a predetermined timing by a valve operating mechanism having a camshaft 34 or the like so that the intake stroke and the exhaust stroke of each cylinder 2A to 2D are performed with the predetermined phase difference as described above. Driven to open and close.
[0038]
Further, among the intake / exhaust valves, the first intake valve 31a, the second intake valve 31b, and the first exhaust valve 32a are operated by a first switching mechanism 35a for switching each valve between an operating state and a stopped state. Is provided. In the valve mechanism of the second exhaust valve 32b among the intake / exhaust valves, the valve is switched between an operating state and a stopped state, and the valve opening period is switched according to the operating state of the engine. A switching mechanism 35b is provided.
[0039]
As shown in FIG. 4, the first switching mechanism 35a includes a camshaft 34 disposed above the first intake valve 31a, the second intake valve 31b, and the first exhaust valve 32a, the camshaft 34, and the above-mentioned It has a rocker shaft 55 disposed between the valves and first to third rocker arms 56 to 58 supported by the rocker shaft 55. The camshaft 34 is integrally formed with a first cam 52 for stopping a valve having a circular outer peripheral surface and second and third cams 53 and 54 having projecting portions (cam noses) for driving the valve. Has been. The second and third cams 53 and 54 have the same shape, and are disposed on the left and right sides of the first cam 52 so as to sandwich the first cam 52.
[0040]
The first rocker arm 56 is disposed at a position corresponding to the first cam 52, and at the tip of the first rocker arm 56 is located at the upper end of the valve shaft of the first intake valve 31a, the second intake valve 31b, and the first exhaust valve 32a. An abutting portion 60 that abuts is provided. On the other hand, the second and third rocker arms 57 and 58 are arranged on both sides of the first rocker arm 56 so as to sandwich the first rocker arm 56, and are separated from the first rocker arm 56 and are not shown in the drawing. Thus, they are urged to be brought into pressure contact with the second and third cams 53 and 54, respectively.
[0041]
The second and third rocker arms 57 and 58 are configured to be connectable to the first rocker arm 56. Specifically, plungers (not shown) provided on the second and third rocker arms 57 and 58 are operated by hydraulic oil supplied from first and second hydraulic oil supply / discharge passages 36 and 38 described later. The first rocker arm 56 and the second and third rocker arms 57 and 58 are integrated with each other by being driven and having its distal end inserted into a connection hole (not shown) formed in the first rocker arm 56. It swings and displaces in a connected state.
[0042]
That is, the hydraulic oil from the first and second hydraulic oil supply / discharge passages 36, 38 is supplied by the first and second control valves 37, 39 provided in the first and second hydraulic oil supply / discharge passages 36, 38. The first and second rocker arms 57 driven by the second and third cams 53 and 54 are connected to the first rocker arm 56 and the second and third rocker arms 57 and 58 by controlling supply and discharge. , 58 is transmitted to the first rocker arm 56, and the first intake valve 31a, the second intake valve 31b, and the first exhaust valve 32a are opened and closed.
[0043]
On the other hand, when the connection state between the first rocker arm 56 and the second and third rocker arms 57 and 58 is released, transmission of the driving force from the second and third rocker arms 57 and 58 to the first rocker arm 56 is interrupted, Even if the camshaft 34 rotates, the first intake valve 31a, the second intake valve 31b, and the first exhaust valve 32a are maintained in the closed state without the first rocker arm 56 swinging and displaced. Yes.
[0044]
The second switching mechanism 35b provided in the valve operating mechanism of the second exhaust valve 32b is formed such that the cam nose of the second cam 53 and the cam nose of the third cam 54 are formed in different shapes. By selectively switching between a state where the first rocker arm 56 is connected to the second rocker arm 57 and a state where the first rocker arm 56 is connected to the third rocker arm 58, a second rocker arm 56 is provided on the camshaft 34. The second switching mechanism 35a is configured in the same manner as the first switching mechanism 35a except that the opening period of the second exhaust valve 32b driven by the third cams 53 and 54 is switched.
[0045]
FIG. 5 shows the configuration of the 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 and O ₂ A signal from the sensor 23 is input, and a signal from an engine speed sensor 61 for detecting the engine speed and an accelerator position sensor 62 for detecting an accelerator position (accelerator pedal depression amount) to determine the driving state are also included. Have been entered. Control signals are output from the ECU 40 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. .
[0046]
The ECU 40 includes an operation state determination unit 41 that determines the operation state of the engine, a mode setting unit 42 that sets an operation mode according to the determination result of the operation state, and the first and second switching mechanisms 35a, The valve opening / closing control means 43 for controlling the opening / closing of the intake / exhaust valves by switching the operation state of the cam provided in 35b, the intake air amount control means 44 for controlling the inflow amount of intake air to each of the cylinders 2A to 2D, and fuel injection Fuel injection control means 45 for controlling the operating state of the valve 9 and ignition control means 46 for controlling the operating state of the spark plug 7 are provided.
[0047]
As shown in FIG. 6, the operating state discriminating means 41 includes an operating region (partial load operating region) A on the low load and low rotation side and an operating region (full load operating region) on the high load side or high rotation side. ) The control state divided into B and the engine operating state (engine speed and engine load) examined by signals from the rotational speed sensor 61 and the accelerator opening sensor 62 and the like are the operating regions A, It is configured to determine which region of B is present.
[0048]
Further, the partial load operation region A that becomes the special operation mode includes a low load side region A1 in which the engine speed and load are low, a medium load side region A2 in which the engine load is higher than the low load side region A1, and The vehicle is partitioned into a high load side region A3 where the engine load is higher than the middle load side region A2.
[0049]
Based on the determination result of the operation mode determination means 41, the mode setting means 42, in the partial load operation region A on the low load and low rotation side, burns the burned gas discharged from the preceding cylinders 2A and 2D in the exhaust stroke. A special operation mode is selected in which it is introduced into the subsequent cylinders 2B and 2C in the intake stroke and burned, and in the high load side or high rotation side operation region B, each cylinder 2A to 2D is burned independently. The operation mode is selected.
[0050]
The valve opening / closing control means 43 is in a two-cylinder connection state in which the burnt gas of the preceding cylinder is introduced into the succeeding cylinder via the inter-cylinder gas passage 22 in the special operation mode, and fresh air is introduced into each cylinder in the normal operation mode. The first and second switching mechanisms 35a and 35b are controlled so as to change the intake / exhaust flow state so that each cylinder is in an independent state. Specifically, the engine is in any one of the operating regions A and B. Accordingly, the control valves 37 and 39 are controlled to operate the first and second switching mechanisms 35a and 35b, whereby the intake / exhaust valves are controlled as follows in principle.
[0051]

The intake air amount control means 44 controls the opening degree of the throttle valve 17 (throttle opening degree) by controlling the actuator 18, and obtains a target intake air amount from a map or the like according to the operating state. The throttle opening is controlled according to the target intake air amount. In this case, in the operation region (partial load operation region) A in the special operation mode, an amount of air necessary to make the air-fuel ratio of the preceding cylinders 2A, 2D the lean air-fuel ratio is supplied to the preceding cylinders 2A, 2D. In addition, in the succeeding cylinders 2B and 2C, fresh air introduced from the branch intake passage 16, excess air in the gas introduced from the preceding cylinders 2A and 2D, and fuel newly supplied from the fuel injection valve 9 The throttle opening is adjusted so that the ratio to the stoichiometric air-fuel ratio becomes the stoichiometric air fuel ratio.
[0052]
The fuel injection control means 45 is configured to control the fuel injection amount and the injection timing from the fuel injection valve 9 provided in each of the cylinders 2A to 2D according to the operating state of the engine. The ignition control means 46 is configured to perform control such as ignition timing control and ignition stop according to the operating state. In particular, the combustion state control, that is, the fuel injection control and the ignition control state, is changed depending on whether the operation state is in the partial load operation region A or the entire operation region B in FIG. ing.
[0053]
That is, in the partial load operation region A in which the control in the special operation mode is executed, when the vehicle is in the low load side region A1, the excess air ratio λ in the preceding cylinders 2A and 2D is significantly reduced to 3 or more. The state, for example, control for setting the air-fuel ratio A / F within the range of 50 to 60 is executed by the air-fuel ratio control means comprising the intake air amount control means 44 and the fuel injection control means 45, as shown in FIG. The control for setting the injection timing so as to perform the fuel injection F from the fuel injection valve 9 into the preceding cylinders A and 2D in the compression stroke and performing the stratified combustion by performing the forced ignition S near the compression top dead center, The fuel injection control means 45 and the ignition control means 46 are configured to be executed by a combustion control means.
[0054]
In the middle load side region A2 of the partial load operation region A, the control for setting the excess air ratio λ in the preceding cylinders 2A and 2D to about 2 to 3, for example, the air-fuel ratio A / F in the range of 35 to 40 is performed. The fuel ratio control means, that is, the intake air amount control means 44 and the fuel injection control means 45 are executed mainly by the fuel injection control means 45, and as shown in FIG. By setting the injection timing so as to perform the fuel injection F in 2D, and outputting the ignition command signal to the spark plug 7 and performing the assist ignition S on the air-fuel mixture near the top dead center before the compression top dead center, Ignition assist control that promotes compression self-ignition of the preceding cylinders 2A, 2D by increasing the pressure around the spark plug 7 is executed in the ignition assist means comprising the ignition control means 46. .
[0055]
In the middle load side region A2 of the partial load region A where the ignition assisting means promotes the compression self-ignition of the preceding cylinders 2A and 2D, the closing timing of the second exhaust valve 32b provided in the preceding cylinders 2A and 2D. By operating the second switching mechanism 35b so that t1 is advanced compared to the normal valve closing timing t2 indicated by the broken line in FIG. 7, the internal EGR amount of the preceding cylinders 2A and 2D is increased to increase the in-cylinder position. Control is performed in the in-cylinder temperature raising means comprising the valve opening / closing control means 43 to raise the temperature and thereby bring the preceding cylinders 2A and 2D into a state in which they are easily subjected to compression self-ignition.
[0056]
Further, in the high load side region A3 of the partial load operation region A, control is performed so that the excess air ratio λ in the preceding cylinders 2A and 2D is about 1 to 3, for example, the air-fuel ratio A / F is in the range of 20 to 40. The control is executed by the air-fuel ratio control means including the fuel injection control means 45 and the like, and the control for promoting the compression self-ignition of the preceding cylinders 2A and 2D is executed by the ignition assist means including the ignition control means 46, and the valve Control for increasing the temperature in the preceding cylinders 2A, 2D by the in-cylinder temperature increasing means comprising the opening / closing control means 43 is executed as necessary.
[0057]
The air-fuel ratio of the preceding cylinders 2A and 2D is set as described above according to the operating region, so that the operating region of the engine starts from the middle load side region A2 that compresses and auto-ignites the preceding cylinders 2A and 2D. , 2D is forcibly ignited, the control is executed to suddenly change the air-fuel ratio of the preceding cylinders 2A, 2D to the lean side when shifting to the low load side region A1.
[0058]
In addition, for the subsequent cylinders 2B and 2C, fuel is supplied to fresh air introduced from the branch intake passage 16 in consideration of the oxygen concentration in the burned gas introduced from the preceding cylinders 2A and 2D. Thus, the fuel injection amount is controlled so that the air-fuel ratio of the succeeding cylinders 2B and 2C becomes substantially the stoichiometric air-fuel ratio, and the injection timing is set so that the fuel is injected before or before the intake stroke, In addition, in order to perform compression self-ignition, control for stopping forced ignition is executed by the combustion state control means.
[0059]
When the engine operating region shifts from the medium load side region A2 in which the preceding cylinders 2A and 2D are compressed and self-ignited to the low load side region A1 in which the preceding cylinders 2A and 2D are forcibly ignited, the preceding cylinders 2A and 2D As the air-fuel ratio is suddenly changed from about A / F 35-40 to about 50-60 to achieve a remarkable lean state, control for increasing the fuel injection amount for the succeeding cylinders 2B, 2C is executed. Yes.
[0060]
On the other hand, when the operating state of the engine is in the full load operation region B on the high load side or the high rotation side, the air-fuel ratio of each cylinder 2A to 2D is set to the stoichiometric air-fuel ratio or lower as control in the normal operation mode. In this way, the fuel injection amount is controlled so that, for example, the stoichiometric air-fuel ratio is set in the most region in the operating region B, and the stoichiometric air-fuel ratio is made richer in the fully-open load and the operating region in the vicinity thereof. In this case, the injection timing is set so that the air-fuel mixture is made uniform by injecting fuel to each of the cylinders 2A to 2D, and in principle, the cylinders 2A to 2D are also ignited forcibly. To control.
[0061]
In the partial operation region A on the low load and low rotation side in the above configuration, the control in the special operation mode is executed, and the first exhaust valve 32a and the first intake valve 31a are stopped as described above, the second exhaust valve 32b and When the second intake valve 31b is activated, the actual fresh air and gas flow paths are as shown in FIG. 8 and are discharged from the preceding cylinders (first and fourth cylinders) 2A and 2D. Burned gas is introduced as it is into the subsequent cylinders (second and third cylinders) 2B and 2C through the inter-cylinder gas passage 22, and only the gas discharged from the subsequent cylinders 2B and 2C is introduced into the exhaust passage 20. Such a two-cylinder connection state is set.
[0062]
In the low load side region A1 of the partial load operation region A1, as shown by the arrow a in FIG. 3, the fresh air is introduced from the intake passage 15 during the intake stroke of the preceding cylinders 2A and 2D, thereby leading the preceding cylinder 2A. The fuel injection amount is controlled so that the 2D air-fuel ratio A / F becomes a remarkably lean state of, for example, about 50 to 60, and the fuel injection F is performed before or before the intake stroke, The forced ignition S is performed at the timing, and the lean air-fuel ratio combustion is performed.
[0063]
Further, in the middle load side region A2 of the partial load operation region A1, the fuel injection amount is controlled so that the air-fuel ratio A / F of the preceding cylinders 2A and 2D becomes lean, for example, about 35 to 40. As shown in FIG. 7, the fuel injection F is performed before or before the intake stroke. Then, the preceding cylinders 2A, 2D are advanced by advancing the closing timing t1 of the second exhaust valve 32b provided in the preceding cylinders 2A, 2D as compared with the normal closing timing t2 indicated by a broken line in FIG. Control for increasing the temperature in 2D so that compression self-ignition is likely to occur is performed in the in-cylinder temperature increasing means including the valve opening / closing control means 43 and near the top dead center before the compression top dead center. The control for accelerating the compression self-ignition by performing the assist ignition S is executed by the ignition assist control means including the ignition control means 46, whereby the compression self-ignition of the preceding cylinders 2A and 2D is performed.
[0064]
Further, in the high load side region A3 of the partial load operation region A1, the fuel injection amount is controlled so that the air-fuel ratio A / F of the preceding cylinders 2A, 2D becomes lean, for example, about 20 to 40, and is necessary. Accordingly, the temperature increase control by the in-cylinder temperature increasing means and the compression self-ignition promotion control by the ignition assist control means are executed, whereby the compression self-ignition of the preceding cylinders 2A and 2D is performed.
[0065]
The burned gas discharged from the preceding cylinders 2A and 2D passes through the inter-cylinder gas passage 22 during the period in which the intake strokes of the preceding cylinders 2A and 2D overlap with the exhaust strokes of the succeeding cylinders 2B and 2C. 2C (the white arrow in FIG. 3 and the arrow b in FIG. 8), 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. So that the stoichiometric air-fuel ratio is obtained. ₂ While the fuel injection amount is controlled based on the output of the sensor 23, fuel is injected at an appropriate timing and combustion is performed.
[0066]
In this case, since the high-temperature burned gas discharged from the preceding cylinders 2A and 2D is introduced into the succeeding cylinders 2B and 2C through the inter-cylinder gas passage 22, the temperature in the combustion chamber is increased in the intake stroke in the succeeding cylinders 2B and 2C. From this state, the pressure and temperature further increase in the compression stroke, so that 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. Moreover, the burned gas is sufficiently mixed and evenly distributed from the time when it is discharged from the preceding cylinders 2A and 2D to the time when it is introduced into the succeeding cylinders 2B and 2C, and further from the fuel injection valve 9 in the intake stroke. Since the fuel injected into 11a is also uniformly dispersed throughout the combustion chamber until the end of the compression stroke, a uniform air-fuel mixture distribution that satisfies the ideal simultaneous compression self-ignition condition can be obtained.
[0067]
On the other hand, in the full load region B on the high load side or the high rotation side, the normal operation mode is set, and as described above, 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. When the valve 31b is brought into a stopped state, the substantial fresh air and gas flow paths are brought into an independent state of each cylinder as shown in FIG. 9, and substantially the intake ports 31, 31a of each of the cylinders 2A to 2D and The exhaust ports 12a and 12 are independent, and fresh air is introduced from the intake passage 15 to the intake ports 31 and 31a of the respective cylinders 2A to 2D, and the exhaust ports 32 and 32a of the respective cylinders 2A to 2D are already introduced into the exhaust passage 20. Fuel gas is discharged. In this case, the engine output performance is ensured by controlling the intake air amount and the fuel injection amount so that the air-fuel ratio of each cylinder becomes the stoichiometric air-fuel ratio or richer.
[0068]
As described above, in the preceding cylinders 2A and 2D, the thermal efficiency is increased by combustion at the lean air-fuel ratio, and the pumping loss is reduced by reducing the intake negative pressure as compared with the normal engine, while the succeeding cylinders 2B and 2D In 2C, while the air-fuel ratio is substantially the stoichiometric air-fuel ratio, compression self-ignition is performed in a uniform air-fuel mixture distribution state, so that thermal efficiency is improved and burned gas extruded from the preceding cylinders 2A and 2D is sent. Therefore, the pumping loss is further reduced as compared with the preceding cylinders 2A and 2D. These effects greatly improve fuel efficiency.
[0069]
In the middle load region A2 where the control in the special operation mode is executed, the preceding cylinders 2A and 2D are compressed and self-ignited by increasing the temperature in the preceding cylinders 2A and 2D. Therefore, the fuel efficiency can be greatly improved by rapidly performing combustion in the preceding cylinders 2A and 2D to improve the thermal efficiency, and the reaction between oxygen and nitrogen in the preceding cylinders 2A and 2D is urgent. Therefore, the generation of NOx can be sufficiently suppressed.
[0070]
Further, in the low load region A1 of the partial load region A where it is difficult to compress and ignite the preceding cylinders 2A and 2D, the air-fuel mixture in the preceding cylinders 2A and 2D is forcibly ignited and the preceding cylinders 2A and 2D are compressed. Since the air-fuel ratio of the preceding cylinders 2A, 2D is suddenly changed to the lean side at the time of transition from the operation area (A2) for self-ignition to the operation area (A1) forcibly igniting the preceding cylinders 2A, 2D, compression self-ignition The amount of NOx generated in the preceding cylinders 2A and 2D can be effectively reduced without performing the above.
[0071]
That is, the amount of NOx generated corresponding to the air-fuel ratio A / F of the air-fuel mixture is, as shown in FIG. 10, from the point that the air-fuel ratio A / F of the air-fuel mixture becomes the stoichiometric air-fuel ratio 14.7 (λ = 1). Since the air-fuel ratio A / F tends to decrease significantly as the air-fuel ratio A / F goes to the lean side beyond the maximum point, the air-fuel ratio of the preceding cylinders 2A and 2D becomes conspicuous. By setting the lean state, the generation of NOx can be effectively suppressed. In addition, since the air-fuel mixture in the preceding cylinders 2A and 2D is forcibly ignited at the time of transition to the operation region, a predetermined amount of burned gas (according to internal EGR) is generated in the preceding cylinders 2A and 2D due to a delay in control response. It is possible to reliably burn the air-fuel mixture of the preceding cylinders 2A and 2C without causing misfire even in a situation where there is (burned gas).
[0072]
Further, in the succeeding cylinders 2B and 2C, the burned gas from the preceding cylinders 2A and 2D is introduced, so that a state equivalent to a large amount of EGR is performed and a special heating means is used. Because rapid combustion by simultaneous compression self-ignition is performed without taking measures such as extremely increasing the compression ratio of the engine, the reaction between oxygen and nitrogen is avoided as much as possible. Generation | occurrence | production can fully be suppressed and it becomes advantageous to the improvement of an emission also from such a point.
[0073]
Therefore, as shown in the above embodiment, in the operation region A2 in which the preceding cylinders 2A and 2D are compressed and self-ignited, the excess air ratio of the preceding cylinders 2A and 2D is set within a range of 2 to 3, thereby In the operation region A1 in which the amount of fuel injected into 2D can be ensured to be a predetermined value, the compression self-ignition of the preceding cylinders 2A, 2D can be performed properly, and the preceding cylinders 2A, 2D are forcibly ignited. By controlling the air-fuel ratio so that the excess air ratio of the preceding cylinders 2A and 2D is 3 or more, it is possible to obtain a remarkable fuel efficiency improvement effect while effectively suppressing the generation of NOx in the operation region A1. There are advantages.
[0074]
Further, in the above-described embodiment, the excess air ratio of the preceding cylinders 2A and 2D is higher in the region A3 on the higher load side than the operation region A2 in which the preceding cylinders 2A and 2D are compressed and self-ignited by increasing the temperature of the preceding cylinders. Since the air-fuel ratio is controlled so as to be in the range of 2 to 3, in the high load side region A3 where the in-cylinder temperature of the succeeding cylinders 2B and 2C tends to increase, the preceding cylinders 2A and 2C By introducing a large amount of burned exhaust gas into the succeeding cylinders 2B and 2C, generation of NOx in the succeeding cylinders 2A and 2C can be effectively suppressed.
[0075]
Further, as shown in the above embodiment, the ignition control means 46 for accelerating the compression self-ignition of the preceding cylinders 2A and 2D by igniting the air-fuel mixture in the preceding cylinders 2A and 2C near the top dead center before the compression top dead center. When the ignition assist means is provided, the preceding cylinder 2A is controlled by the in-cylinder temperature increasing means including the valve opening / closing control means 43 in the middle load area A2 or the like of the partial load area A in which the control of the special operation mode is executed. In the state in which the in-cylinder temperature of 2D is raised, the air-fuel mixture in the preceding cylinders 2A, 2D is ignited to instantaneously increase the pressure in the preceding cylinders 2A, 2D, thereby appropriately adjusting the preceding cylinders 2A, 2 There is an advantage that compression self-ignition can be surely performed at the time.
[0076]
In the partial load region A in which the control in the special operation mode is performed, when it is difficult to cause the subsequent cylinders 2B and 2C to perform compression self-ignition, for example, the temperature of burned gas derived from the preceding cylinders 2A and 2D When the engine is in the low load side region A1 where the engine is low, the air-fuel mixture of the succeeding cylinders 2B and 2C is ignited near the top dead center before the compression top dead center by the ignition assist means 2 comprising the ignition control means 46. Then, the subsequent cylinders 2B and 2C may be compressed and ignited reliably at an appropriate time by instantaneously increasing the pressure in the subsequent cylinders 2B and 2C.
[0077]
In the above-described embodiment, the control for increasing the fuel injection amount for the subsequent cylinders 2B and 2C is performed at the time of transition from the operation region A2 in which the preceding cylinders 2A and 2D are compression self-ignited to the operation region A1 in which the preceding cylinders 2A and 2D are forcibly ignited. Since the fuel injection control means 45 is configured to execute the operation, the fuel injection amount for the subsequent cylinders 2B and 2C is set to a sufficient value at the time of transition to the operation region A1, so that the preceding cylinders 2A and 2D are set as described above. There is an advantage that the engine output decrease due to the lean air-fuel ratio being set to a remarkable lean state can be compensated by the combustion energy of the succeeding cylinders 2B and 2C.
[0078]
Further, as shown in the above embodiment, when the valve opening / closing control means 43 is configured to quickly close the second exhaust valves 32b of the preceding cylinders 2A, 2D to increase the temperature in the preceding cylinders 2A, 2D. Since the in-cylinder temperature can be increased by increasing the internal EGR amount of the preceding cylinders 2B and 2D in the specific operation region A2 of the partial load region A where combustion in the special operation mode is performed, the preceding cylinders 2A and 2D Can be easily and effectively compressed self-ignited.
[0079]
In the above embodiment, the air-fuel ratio of the succeeding cylinders 2B and 2C is set to the substantially stoichiometric air-fuel ratio in the special operation mode, and only the exhaust gas burned at the stoichiometric air-fuel ratio is discharged from the succeeding cylinders 2B and 2C to the exhaust passage 20. Therefore, the exhaust gas purification performance can be sufficiently ensured by the three-way catalyst 24 without providing the lean NOx catalyst in the exhaust passage 20. In addition, since it is not necessary to provide the lean NOx catalyst, it is not necessary to perform temporary enrichment for NOx release and reduction when the NOx occlusion amount of the lean NOx catalyst increases, thereby preventing a reduction in fuel consumption improvement effect. In addition, there is an advantage that the problem that the lean NOx catalyst is poisoned with sulfur can be prevented.
[0080]
In the special operation mode, the air-fuel ratio of the succeeding cylinders 2B and 2C is set to a substantially stoichiometric air-fuel ratio, and only exhaust gas burned at the stoichiometric air-fuel ratio is discharged from the succeeding cylinders 2B and 2C to the exhaust passage 20. Instead of the above embodiment, as shown in FIG. 11, a three-way catalyst 24 and a NOx purification catalyst 26 for purifying NOx in the exhaust gas are provided in the exhaust passage 20, and the preceding cylinders 2A and 2D and the succeeding cylinders 2B and 2C are provided. In the operation region A2 in which the compression self-ignition is performed, the air-fuel ratio of the subsequent cylinders 2B and 2C may be set slightly lean.
[0081]
The NOx purification catalyst 26 adsorbs NOx in the exhaust gas in a lean state in which the air-fuel ratio is larger than the stoichiometric air-fuel ratio, and releases and reduces NOx adsorbed in the rich state. Having a catalyst layer in which alumina or ceria is supported as a support material. The support material includes a noble metal such as platinum Pt, rhodium Rh, or palladium Pd, an alkali metal such as potassium K, and an alkaline earth metal such as barium Ba. A one-coat type in which is supported. The lean NOx catalyst 26 includes alumina having platinum Pt and an alkaline earth metal such as barium Ba supported on the wall surface of the carrier, an inner catalyst having ceria, and a zeolite having a noble metal such as platinum Pt supported thereon. You may use the 2 coat type thing which consists of an outer side catalyst.
[0082]
When configured as described above, the amount of NOx generated can be effectively reduced by compressing and auto-igniting the preceding cylinders 2A and 2D and the succeeding cylinders 2B and 2C. The exhaust purification performance can be maintained without setting the capacity of the NOx adsorption catalyst 26 to a large value, and the fuel efficiency of the engine can be made more effective by setting the air-fuel ratio of the succeeding cylinders 2B and 2C to be slightly lean. There is an advantage that can be improved. Moreover, when the NOx adsorbed on the NOx adsorption catalyst 26 is released and reduced, the NOx adsorbed on the NOx adsorption catalyst 25 is released by setting the air-fuel ratio of the succeeding cylinders 2B and 2C to be slightly rich. In addition, since NOx can be reduced by a reducing agent made of CO or CH in the exhaust gas, there is an advantage that the exhaust gas purification performance can be maintained without significantly reducing the fuel consumption.
[0083]
In the above embodiment, the fuel injection valve 9 is a direct injection type that directly injects fuel into the combustion chamber for any of the preceding cylinders 2A, 2D and the succeeding cylinders 2B, 2C. The injection valve is not necessarily limited to the direct injection type. For example, a fuel injection valve is provided in the intake port and the inter-cylinder gas passage, the fuel injection valve of the intake port is driven in the normal operation mode, and the inter-cylinder gas passage is driven in the special operation mode. The fuel injection valve may be driven.
[0084]
Further, an electromagnetic valve mechanism having a solenoid actuator for opening and closing the intake / exhaust air valve is provided, and the operation state of the electromagnetic valve mechanism is controlled, so that the two-cylinder connected state and each cylinder independent state are provided. Alternatively, the internal EGR amount of the preceding cylinders 2A and 2D may be increased by advancing the valve opening period of the second exhaust valve 32B provided in the preceding cylinders 2A and 2D. You may make it raise temperature.
[0085]
The apparatus of the present invention can also be applied to multi-cylinder engines other than four cylinders. For example, in the case of six cylinders, the exhaust stroke of one cylinder and the intake stroke of another cylinder do not completely overlap. In such a case, the exhaust stroke of one cylinder precedes the intake stroke of the other cylinder. In addition, two cylinders in which both strokes partially overlap may be used as a pair of preceding and succeeding cylinders.
[0086]
【The invention's effect】
As described above, the present invention provides a multi-cylinder spark ignition in which a fuel injection valve for supplying fuel into each cylinder is provided in the intake air introduction path, and the combustion cycle of each cylinder is set to have a predetermined phase difference. In the engine, the burned gas discharged from the preceding cylinder in the exhaust stroke between the pair of cylinders in which the exhaust stroke and the intake stroke overlap in the partial load operation region of the engine is directly transferred to the subsequent cylinder in the intake stroke. Combusting with the air / fuel ratio of the preceding cylinder set to a lean air / fuel ratio greater than the stoichiometric air / fuel ratio while the burned gas introduced through the passage and exhausted from the succeeding cylinder is guided to the exhaust passage. An operation mode for performing control of a special operation mode in which fuel is supplied to burned gas having a lean air-fuel ratio introduced from the preceding cylinder to the succeeding cylinder to burn the succeeding cylinder. And an in-cylinder temperature increasing means for compressing and auto-igniting the preceding cylinder together with the succeeding cylinder by increasing the temperature in the preceding cylinder in at least a part of the operation region in which the control of the special operation mode is executed. Ignition control means for controlling the air-fuel mixture in the preceding cylinder to be ignited forcibly in the operation region where it is difficult to perform compression self-ignition of the cylinder, and operation for forcibly igniting the preceding cylinder from the operation region in which the preceding cylinder is compressed and self-ignited The air-fuel ratio control means for controlling the air-fuel ratio of the preceding cylinder to suddenly change to the lean side at the time of transition to the region is provided, so that when the combustion is performed in the special operation mode, the thermal efficiency of the preceding cylinder due to lean combustion is provided. Improved fuel efficiency by reducing pumping loss and improving fuel efficiency by reducing pumping loss in subsequent cylinders There is an advantage that is obtained.
[0087]
Further, in the in-cylinder temperature increasing means, the in-cylinder temperature increasing means performs compression self-ignition of both the preceding cylinder and the succeeding cylinder by increasing the in-cylinder temperature of the preceding cylinder in at least a part of the specific operation region where the combustion in the special operation mode is performed. By doing so, it is possible to effectively suppress the generation of NOx, and to obtain a remarkable fuel economy improvement effect. Then, at the time of transition from the operation region in which the preceding cylinder is subjected to compression self-ignition to the operation region in which the preceding cylinder is forcibly ignited, control is performed to suddenly change the air-fuel ratio of the preceding cylinder to the lean side, thereby occurring during combustion of the preceding cylinder. There is an advantage that the amount of NOx can be remarkably reduced.
[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 cross-sectional view of an engine body and the like.
FIG. 3 is an explanatory diagram showing a combustion cycle, valve opening timing, and the like of a preceding cylinder and a succeeding cylinder.
FIG. 4 is a perspective view showing a specific configuration of switching means.
FIG. 5 is a block diagram of a control system.
FIG. 6 is an explanatory diagram showing an example of operation region setting for performing control according to an operation state.
FIG. 7 is a diagram showing an exhaust stroke, an intake stroke, a fuel injection timing, an ignition timing, and the like of each cylinder.
FIG. 8 is an explanatory diagram showing substantial fresh air and gas flow paths during low load and low rotation.
FIG. 9 is an explanatory diagram showing substantial fresh air and gas flow paths when in an operating region on a high load, high and low rotation side.
FIG. 10 is a graph showing a correspondence relationship between the air-fuel ratio and the NOx generation amount.
FIG. 11 is a schematic plan view showing an entire engine according to another embodiment of the present invention.
[Explanation of symbols]
1 Engine body
2A to 2D cylinder
9 Fuel injection valve
15 Intake passage
20 Exhaust passage
22 Gas passage between cylinders
43 Valve opening / closing control means (operation mode control means)
44 Intake air amount control means (air-fuel ratio control means)
45 Fuel injection control means (air-fuel ratio control means)
55 Ignition control means

Claims (7)

A control device for a multi-cylinder spark ignition engine in which a fuel injection valve for supplying fuel into each cylinder is provided in an intake air introduction path, and a combustion cycle of each cylinder is set to have a predetermined phase difference. ,
In the partial load operation region of the engine, the burnt gas discharged from the preceding cylinder in the exhaust stroke between a pair of cylinders in which the exhaust stroke and the intake stroke overlap is directly passed to the subsequent cylinder in the intake stroke via the inter-cylinder gas passage. The combustion is performed with the air-fuel ratio of the preceding cylinder set to be a lean air-fuel ratio larger than the stoichiometric air-fuel ratio while the two-cylinder connection state is established in which the burned gas discharged from the subsequent cylinder is introduced into the exhaust passage. An operation mode control means for performing control in a special operation mode in which fuel is supplied to burned gas having a lean air-fuel ratio introduced from the preceding cylinder to the succeeding cylinder to burn the succeeding cylinder;
In-cylinder temperature increasing means for compressing and self-igniting the preceding cylinder together with the succeeding cylinder by increasing the temperature in the preceding cylinder in at least a part of the operation region in which the control of the special operation mode is performed,
An ignition control means for controlling the air-fuel mixture in the preceding cylinder to be forcedly ignited in an operation region where it is difficult to perform compression self-ignition of the preceding cylinder;
Air-fuel ratio control means for controlling the air-fuel ratio of the preceding cylinder to suddenly change to the lean side at the time of transition from the operation region in which the preceding cylinder is subjected to compression self-ignition to the operation region in which the preceding cylinder is forcibly ignited. A spark ignition engine control device. The air-fuel ratio is set so that the excess air ratio of the preceding cylinder is in the range of 2 to 3 in the operation region where the preceding cylinder is compressed and self-ignited, and the excess air ratio of the preceding cylinder is 3 or more in the operation region where the preceding cylinder is forcibly ignited. The control device for a spark ignition engine according to claim 1, wherein the control device is controlled. The air-fuel ratio is controlled so that the excess air ratio of the preceding cylinder is within a range of 2 to 3 in the high load side region from the operation region in which the preceding cylinder is compressed and self-ignited by increasing the temperature of the preceding cylinder. The control device for a spark ignition engine according to claim 1 or 2, The ignition assisting means for accelerating the compression self-ignition of the preceding cylinder by igniting the air-fuel mixture in the preceding cylinder near the top dead center before the compression top dead center is provided. The spark-ignition engine control device according to item. 5. The fuel injection control means for increasing the fuel injection amount for the succeeding cylinder at the time of transition from the operation region for compressing and auto-igniting the preceding cylinder to the operation region for forcibly igniting the preceding cylinder. The control device for a spark ignition engine according to any one of the preceding claims. 6. The spark according to claim 1, wherein the in-cylinder temperature raising means is constituted by a valve opening / closing control means for quickly closing the exhaust valve of the preceding cylinder to raise the temperature in the preceding cylinder. Control device for ignition engine. The spark ignition engine control device according to any one of claims 1 to 6, wherein an air-fuel ratio of the subsequent cylinder is set to be slightly lean in an operation region in which the preceding cylinder and the subsequent cylinder are subjected to compression self-ignition. .

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