JP3606112B2 - Diesel engine control device - Google Patents

Description

図３１は可変ノズル２ｄの有効面積相当値Ａｖｎｔの演算フローである。ステップ１では指令開度ＶＮＴｓｔｅｐ、総排気重量Ｑｔｏｔａｌ（＝Ｑａｓ０＋Ｑｆ）、排気温度Ｔｅｘｈを読み込む。
【０１０９】
このうち総排気重量Ｑｔｏｔａｌと排気温度Ｔｅｘｈからステップ２で
【０１１０】
【数１０】
Ｗｅｘｈ＝Ｑｔｏｔａｌ×Ｔｅｘｈ／Ｔｓｔｄ ［ｍ^３／ｓｅｃ］
ただし、Ｔｓｔｄ：標準大気温度、
の式により排気流速相当値Ｗｅｘｈを算出する。
【０１１１】
ステップ３では、この排気流速相当値Ｗｅｘｈの平方根をとった値から図３２を内容とするテーブルを検索して摩擦損失ξｆｒｉｃを演算する。ステップ４では指令開度ＶＮＴｓｔｅｐと総ガス重量Ｑｔｏｔａｌから図３３を内容とするマップを検索してノズル損失ξｃｏｎｖを演算する。そして、これら２つの損失ξｆｒｉｃ、ξｃｏｎｖをステップ５において指令開度ＶＮＴｓｔｅｐに乗算して、つまり
【０１１２】
【数１１】
Ａｖｎｔ＝ ＶＮＴｓｔｅｐ×ξｆｒｉｃ×ξｃｏｎｖ
の式により可変ノズルの有効面積相当値Ａｖｎｔを演算する。
【０１１３】
図３４は排気圧（タービン入口圧）Ｐｅｘｈの演算のフローである。
【０１１４】
ステップ１では吸気量の加重平均値Ｑａｓ０、燃料噴射量Ｑｆ、有効面積相当値Ａｖｎｔ、排気温度Ｔｅｘｈ、大気圧（コンプレッサ入口圧）Ｐａを読み込み、これらのパラメータを用い、ステップ２において
【０１１５】
【数１２】
Ｐｅｘｈ０＝Ｋｐｅｘｈ×｛（Ｑａｓ０＋Ｑｆｕｅｌ）／Ａｖｎｔ｝^２×Ｔｅｘｈ＋Ｐａ
ただし、Ｋｐｅｘｈ：定数、
の式により排気圧Ｐｅｘｈ０を演算し、この排気圧に対してステップ３で加重平均処理を行い、その加重平均値を排気圧Ｐｅｘｈとして求める。
【０１１６】
ここで、上記の有効面積相当値Ａｖｎｔと排気圧Ｐｅｘｈ０の各演算方法は、次のようにして得たものである。
【０１１７】
〈１〉流路面積が縮小する場合の流れの基礎式
図４４のように緩やかに断面積が縮小する管内を流れる理想流体を考える。
【０１１８】
流体の圧力、流速、面積、比重をそれぞれＰ、ｗ、Ａ、ρとし、入口を添字１、出口を添字２とし、入口と出口の断面についてベルヌイ（Ｂｅｒｎｏｕｌｌｉ）の定理を適用すると、
ｗ_１ ^２／２＋Ｐ_１／ρ＝ｗ_２ ^２／２＋Ｐ_２／ρ ・・・（１ａ）
また、連続の式より
Ａ_１×ｗ_１＝Ａ_２×ｗ_２ ・・・（１ｂ）
したがって、両式からｗ_１を消去すると、
ｗ_２＝１／｛１−（Ａ_２／Ａ_１）^２｝^１／２×｛２（Ｐ_１−Ｐ_２）／ρ｝^１／２［ｍ／ｓｅｃ］ ・・・（２）
単位時間に流れる流量Ｑは、連続の式より一定であるから、

の式より表すことができる。
【０１１９】
（３）式の右辺の１／｛１−（Ａ_２／Ａ_１）^２｝^１／２を効率ηｎとおくと、次の流れの基礎式を得る。
【０１２０】
Ｑ＝ηｎ×Ａ_２×｛２ρ×（Ｐ_１−Ｐ_２）｝^１／２ ・・・（４）
〈２〉ターボチャージャの状態方程式
次に、ターボチャージャ２でのコンプレッサ２ｂと仕事の釣合いの関係を調べる。なお、以下で使用する記号は図４５の通りである。
【０１２１】
コンプレッサ２ｂの実効仕事率Ｌｃは、
Ｌｃ＝Ｑａｓ０×Ｗｃ／ηｃ ［Ｗ］ ・・・（５）
ただし、Ｑａｓ０：吸入新気重量流量［ｋｇ／ｓｅｃ］、
Ｗｃ：コンプレッサ理論仕事［Ｊ／ｋｇ］、
ηｃ：コンプレッサ効率相当値。
【０１２２】
また、タービン２ａの実効仕事率Ｌｔは、
Ｌｔ＝ηｔ×Ｑｔｏｔａｌ×Ｗｔ ［Ｗ］ ・・・（６）
ただし、Ｑｔｏｔａｌ：総排気重量流量［ｋｇ／ｓｅｃ］、
Ｗｔ：タービン理論仕事［Ｊ／ｋｇ］、
ηｔ：タービン効率相当値。
【０１２３】
タービン２ａとコンプレッサ２ｂは軸を介して直結されているので、コンプレッサ２ｂとタービン２ａの実仕事率Ｌｃ、Ｌｔが等しいとおけば（軸受けのフリクションは効率に含まれる）、ターボチャージャ２の状態方程式として次式を得る。
【０１２４】

〈３〉流路面積が縮小する場合の排気圧予測式の検討
（７）式の左辺に上記の（４）式を適用して、

ただし、Ａｖｎｔ：可変ノズルの有効面積相当値、
Ｐｅｘｈ：排気圧、
Ｐａ：大気圧相当値、
ρｅ：排気の密度、
ＶＮＴｓｔｅｐ：指令開度、
ηｎ：効率（損失分）、
の式を得る。
【０１２５】
（８ａ）式を排気圧Ｐｅｘｈについて整理すると、

ここで、排気密度ρｅは理論式によれば
ρｅ＝ρｓｔｄ×（Ｔａ／Ｔｅｘｈ）×（Ｐｅｘｈ／Ｐａ） ・・・（１０）
ただし、ρｓｔｄ：標準大気の密度（≒１．１６７９ｇ／ｃｍ^３）、
Ｔａ：コンプレッサ入口温度、
Ｔｅｘｈ：排気温度、
Ｐｅｘｈ：排気圧、
Ｐａ：大気圧、
であるが、この理論式では排気密度ρｅを求めるのに排気圧Ｐｅｘｈを用いることになって具合が悪いので、
ρｅ≒ρｓｔｄ×（Ｔａ／Ｔｅｘｈ）＝Ｔｓｔｄ／Ｔｅｘｈ ・・・（１１）
ただし、Ｔｓｔｄ：標準大気の温度（≒２９８．１５Ｋ）、
の近似式を用いる。近似できる理由は、排気圧Ｐｅｘｈが高くなれば、排気温度Ｔｅｘｈも高くなるので、排気圧Ｐｅｘｈの変化分を排気温度Ｔｅｘｈに含めて考えることができるからである。
【０１２６】
したがって、（１１）式を（９）式に代入することにより、次の式を得る。
【０１２７】

ただし、Ｋｐｅｘｈ：定数。
【０１２８】
ここで、（１２ａ）式右辺のコンプレッサ理論仕事Ｗｃとタービン理論仕事Ｗｔは次式で与えられる。
【０１２９】
【数１３】

さて、（１２ａ）式より、排気圧Ｐｅｘｈの演算式が求められたが、（１２ａ）式中のη
ｃ、ηｔ、Ｗｃ、Ｗｔの演算は複雑であり（ＥＣＵの能力が要る）、また、（１４）式ではこれから求めようとする排気圧Ｐｅｘｈを知る必要があるので、さらに考える。
【０１３０】
いま、総排気重量Ｑｔｏｔａｌと吸入新気量Ｑａｓ０および燃料噴射量Ｑｆ（単位はすべて［ｋｇ／ｓｅｃ］とする）の間には次の関係がある。
【０１３１】
Ｑｔｏｔａｌ＝Ｑａｓ０＋Ｑｆ ・・・（１５）
（１５）式の左辺に上記の（４）式を適用して、
Ａｖｎｔ×｛２×ρｅ×（Ｐｅｘｈ−Ｐａ）｝^１／２＝Ｑａｓ０＋Ｑｆ ・・・（１６ａ）
Ａｖｎｔ＝ηｎ×ＶＮＴｓｔｅｐ ・・・（１６ｂ）
（１６ａ）式の両辺を２乗して排気圧Ｐｅｘｈについて整理すると、次式が得られる。
【０１３２】
Ｐｅｘｈ＝｛（Ｑａｓ０＋Ｑｆ）／Ａｖｎｔ｝^２×（１／ρｅ）＋Ｐａ ・・・（１７）
ここでも、上記の排気密度ρｅの近似式である（１１）式を（１７）式に代入することにより、次の最終式を得る。
【０１３３】

ただし、Ｋｐｅｘｈ：定数。
【０１３４】
（１８ａ）式は上記の（１２ａ）式と等価であり、（１８ａ）式による排気圧Ｐｅｘｈの演算式には、コンプレッサ２ｂ、タービン２ａの理論仕事の比（Ｗｃ／Ｗｔ）と各々の効率の積（ηｃ×ηｔ）が含まれており、（１８ａ）式を用いれば、ターボチャージャ２の理論仕事Ｗｃ、Ｗｔと効率ηｃ、ηｔが未知であっても考慮したことになる。ゆえにあとは、可変ノズル２ｄを流れるガスの効率ηｎを求めればよい。
【０１３５】
〈４〉ノズルを流れるガスの効率ηｎ
効率ηｎを含んだ可変ノズル２ｄの有効面積相当値Ａｖｎｔは上記の（８ｂ）式、（１６ｂ）式で与えられるが、さらに効率ηｎは次式で表すことができる。
【０１３６】

ただし、ξｃｏｎｖ：ノズル損失、
ξｆｒｉｃ：摩擦損失。
【０１３７】
（１９）式においてノズル損失ξｃｏｎｖは、ノズル開度毎に決まる損失であり、縮まり管の場合、（３）式からわかるように１／｛１−（Ａ_２／Ａ_１）^２｝^１／２が効率になる。
【０１３８】
しかしながら、流速の変化が大きい場合、１／｛１−（Ａ_２／Ａ_１）^２｝^１／２の値をそのままノズル損失ξｃｏｎｖとみなすと、実際のノズル損失と合わないことが多いので、ノズル開度に対する効率のテーブルを持たせることで記述している（図３３参照）。
【０１３９】
また、（１９）式の摩擦損失ξｆｒｉｃは、ノズル内部の流れを層流とみなすとハーゲンポアズイユ（Ｈａｇｅｎ−Ｐｏｉｓｅｕｉｌｌｅ）の式が成り立ち、流速の平方根に摩擦損失ξｆｒｉｃが比例する。そこで、
Ｗｅｘｈ＝Ｑｔｏｔａｌ／ρｅ ・・・（２０）
の式により体積流量相当値Ｗｅｘｈを算出し、これの平方根を排気流速として、これにより摩擦損失ξｆｒｉｃを検索する（図３２参照）。
【０１４０】
ここでも、排気密度ρｅの近似式である（１１）式を（２０）式に代入して、

このようにして、（１９）式によりノズル有効面積相当値Ａｖｎｔを演算し、このＡｖｎｔのほか、Ｑａｓ０、Ｑｆ、Ｔｅｘｈ、Ｐａ を用いて、（１８ａ）、（１８ｂ）式により排気圧Ｐｅｘｈを予測するようにしたわけである。排気圧の実測値と予測値の相関を調べた実験結果を図３５に示す。同図より、予測値でも十分な精度があることがわかる。
【０１４１】
次に、図３６はＥＧＲ（流）量Ｑｅを演算するフローである。ステップ１では上記した吸気圧Ｐｍ、排気圧Ｐｅｘｈ、ＥＧＲ弁実開度としてのＥＧＲ弁実リフト量Ｌｉｆｔｓを読み込む。あるいは、ステップモータのように目標値を与えれば実際のＥＧＲ弁リフト量が一義に決まる場合は、目標ＥＧＲ弁リフト量でもよい。
【０１４２】
ステップ２では、このＥＧＲ弁実リフト量Ｌｉｆｔｓから図３７を内容とするテーブルを検索して、ＥＧＲ弁５７の開口面積相当値Ａｖｅを求める。
【０１４３】
そして、ステップ３において、ＥＧＲ流量Ｑｅを、これら吸気圧Ｐｍと排気圧Ｐｅｘｈ、ＥＧＲ弁５７の開口面積相当値Ａｖｅとから、
【０１４４】
【数１４】
Ｑｅ＝Ａｖｅ×｛（Ｐｅｘｈ−Ｐｍ）×ＫＲ＃｝^１／２
ただし、ＫＲ＃：補正係数（定数）、
の式により計算する。
【０１４５】
図３８は目標ＥＧＲ率Ｍｅｇｒを演算するフローである。ステップ１でエンジン回転数Ｎｅ、燃料噴射量Ｑｆ、シリンダ吸入ガス温度Ｔｎを読み込み、このうちＮｅとＱｆとから図３９を内容とするマップを検索して、目標ＥＧＲ率基本値Ｍｅｇｒ０を求める。ステップ３ではシリンダ吸入ガス温度Ｔｎから図４０を内容とするテーブルを検索して目標ＥＧＲ率補正値Ｈｅｇｒを求め、この目標ＥＧＲ率補正値Ｈｅｇｒを目標ＥＧＲ率基本値Ｍｅｇｒ０に乗ずることによって目標ＥＧＲ率Ｍｅｇｒを計算する。
【０１４６】
図４１は要求ＥＧＲ（流）量Ｔｑｅの演算フローである。ステップ１でエンジン回転数Ｎｅ、目標ＥＧＲ率Ｍｅｇｒ、シリンダ吸入新気量Ｑａｃ、燃料噴射量のサイクル処理値Ｑｆ０を読み込み、このうちシリンダ吸入新気量Ｑａｃに目標ＥＧＲ率Ｍｅｇｒをステップ２において乗ずることで目標吸入ＥＧＲ量Ｍｑｅｃを計算する。
【０１４７】
ステップ３ではこの目標吸入ＥＧＲ量Ｍｑｅｃに対して、Ｋｉｎ×Ｋｖｏｌを加重平均係数として
【０１４８】
【数１５】
Ｒｑｅｃ＝Ｒｑｅｃ_ｎ−１×（１−Ｋｉｎ×Ｋｖｏｌ）＋Ｍｑｅｃ×Ｋｉｎ×Ｋｖｏｌ
ただし、Ｒｑｅｃ_ｎ−１：Ｒｑｅｃの前回値、
の式により中間処理値（加重平均値）Ｒｑｅｃを演算し、この中間処理値Ｒｑｅｃと上記の目標吸入ＥＧＲ量Ｍｑｅｃを用いてステップ４で
【０１４９】
【数１６】
Ｔｑｅｃ＝Ｍｑｅｃ×ＧＫＱＥＣ＋Ｒｑｅｃ_ｎ−１×（１−ＧＫＱＥＣ）
ただし、Ｒｑｅｃ_ｎ−１：Ｒｑｅｃの前回値、
ＧＫＱＥＣ：進み補償ゲイン、
の式により進み処理を行って目標シリンダ吸入ＥＧＲ量Ｔｑｅｃを求める。要求値に対して吸気系の遅れ（すなわちＥＧＲ弁５７→コレクタ５２ａ→吸気マニホールド→吸気弁の容量分の遅れ）があるので、ステップ３、４ではこの遅れ分の進み処理を行うものである。
【０１５０】
ステップ５ではこの目標シリンダ吸入ＥＧＲ量Ｔｑｅｃから、
【０１５１】
【数１７】
Ｔｑｅ＝（Ｔｑｅｃ／Ｎｅ）×ＫＣＯＮ＃
ただし、ＫＣＯＮ＃：定数、
の式により単位変換（１シリンダ当たり→単位時間当たり）を行って、要求ＥＧＲ量Ｔｑｅを計算する。
【０１５２】
図４２は指令ＥＧＲ弁開度としての指令ＥＧＲ弁リフト量Ｌｉｆｔｔを演算するフローである。ステップ１では吸気圧Ｐｍ、排気圧Ｐｅｘｈ、要求ＥＧＲ量Ｔｑｅを読み込む。ステップ２ではＥＧＲ弁５７の要求開口面積Ｔａｖを、
【０１５３】
【数１８】
Ｔａｖ＝Ｔｑｅ／｛（Ｐｅｘｈ−Ｐｍ）×ＫＲ＃｝^１／２
ただし、ＫＲ＃：補正係数（定数）、
の式（流体力学の法則）で計算する。
【０１５４】
ステップ３ではこのＥＧＲ弁５７の要求開口面積Ｔａｖより図４３を内容とするテーブルを検索して目標ＥＧＲ弁開度としてのＥＧＲ弁目標リフト量Ｍｌｉｆｔを求め、この目標リフト量Ｍｌｉｆｔに対して、ステップ４において、ＥＧＲ弁５７の作動遅れ分の進み処理を行い、その進み処理後の値を指令ＥＧＲ弁リフト量Ｌｉｆｔｔとして求める。
【０１５５】
このようにして求められた指令ＥＧＲ弁リフト量Ｌｉｆｔｔが図示しないフローによりステップモータ５７ａへと出力され、ＥＧＲ弁５７が駆動される。
【０１５６】
これでＥＧＲ制御の説明を終了する。
【０１５７】
このように、本発明の実施形態では、吸気量（の加重平均値）Ｑａｓ０、燃料噴射量Ｑｆ、可変ノズルの有効面積相当値Ａｖｎｔ、排気温度Ｔｅｘｈの４つの要素からダイレクトにかつ簡単な上記の数１２式を用いて排気圧Ｐｅｘｈを演算できることになったので、可変容量ターボチャージャを備える場合においても、過渡時に応答遅れなく排気圧を推定できる。
【０１５８】
また、有効面積相当値Ａｖｎｔを、可変ノズル２ｄを流れるガスの効率ηｎと可変ノズル２ｄを駆動するステップモータ２ｃに与える指令開度ＶＮＴｓｔｅｐとの積で与えるようにしたので、可変ノズル２ｄを流れるガスの効率ηｎを考慮できる。
【０１５９】
また、可変ノズル２ｄを流れるガスの効率ηｎは摩擦損失ξｆｒｉｃとノズル損失ξｃｏｎｖの積としたので、摩擦損失とノズル損失を別個に考慮できる。
【０１６０】
また、摩擦損失ξｆｒｉｃを、排気流速相当値Ｗｅｘｈの平方根に比例する値で与えるようにしたので、排気流速が相違しても、摩擦損失ξｆｒｉｃを精度よく与えることができる。
【０１６１】
また、流速の変化が大きい場合、縮まり管に対する損失（上記（３）式の１／｛１−（Ａ_２／Ａ_１）^２｝^１／２の値）をそのままノズル損失とみなすと、実際のノズル損失と合わないことが多いのであるが、本実施形態ではノズル損失ξｃｏｎｖを、指令開度ＶＮＴｓｔｅｐと総排気重量Ｑｔｏｔａｌに応じた値としたので、流速の変化が大きい場合にも実際のノズル損失とよく合致させることができる。
【０１６２】
また、指令開度ＶＮＴｓｔｅｐと排気量Ｑｅｘｈに応じて排気温度のノズル開度補正係数Ｋｔｅｘｈ４を演算し、この補正係数Ｋｔｅｘｈ４で排気温度基本値Ｔｅｘｈｂを補正するようにしたので、排気温度Ｔｅｘｈの演算精度が向上し、この向上分だけ排気圧Ｐｅｘｈの演算精度が向上する。同様にして、吸気ポートにスワール弁を備える場合には、このスワール弁の開度位置とエンジン回転数Ｎｅに応じて排気温度のスワール補正係数Ｋｔｅｘｈ３を演算し、この補正係数Ｋｔｅｘｈ３で排気温度基本値Ｔｅｘｈｂを補正するようにしたので、吸気ポートにスワール弁を備える場合にも排気温度Ｔｅｘｈの演算精度が向上し、この向上分だけ排気圧Ｐｅｘｈの演算精度が向上する。
【０１６３】
なお、排気圧（タービン入口圧）の演算式は上記の数１２式に限られるものでない。これを図４６の演算フローで説明すると、このフローは図３４に置き換わるものである。
【０１６４】
図３４では、ノズル２ｄを通過するガスの流れを、流路面積が縮小する場合の流れである（図４４参照）と仮定して排気圧を演算したのに対して、図４６は、ノズルを通過するガスの流れを、理想気体が断熱変化して流動する場合の流れ（図４８参照）と仮定して求めるものである。図４６において具体的には、ステップ１１で
【０１６５】
【数１９】
Ｐｅｘｈｒ＝Ｋｐｅｘｈｎ×｛（Ｑａｓ０＋Ｑｆｕｅｌ）／Ａｖｎｔ｝^２×Ｔｅｘｈ
ただし、Ｋｐｅｘｈｎ：定数、
の式によりタービン入口排気圧相当値Ｐｅｘｈｒを演算し、このＰｅｘｈｒと大気圧Ｐａからステップ１２において図４７を内容とするマップを検索することにより排気圧Ｐｅｘｈ０を求める。後は、図３４と同じであり、このＰｅｘｈ０に対してステップ３で加重平均処理を行い、その加重平均値を排気圧Ｐｅｘｈとして求める。
【０１６６】
ここで、どのようにして数１９式の排気圧の演算方法を得たかを次に説明する。
【０１６７】
〈５〉先細ノズルの場合の流れの基礎式
タービンノズルを通過する通過する流れを考察すると、外部との熱の出入りや仕事がほとんどないため、流体の持つエネルギは、内部エネルギの減少分が運動エネルギと押し出し仕事に変化すると考えられる。また、エンジンの排気は、低圧・高温なので理想気体とみなせる。したがって、タービンノズルを通過する排気の流れは、「理想気体が断熱変化をして流動する」と考えることができる。
【０１６８】
さて、タービンノズルのような先細ノズルにおいて、図４８に示したように、圧力、比容積、流速、面積、温度、比熱比、気体定数をそれぞれ、Ｐ、ｖ、ｗ、Ａ、Ｔ、κ、Ｒとし、入口を添字１、出口を添字２とすると、
【０１６９】
【数２０】

である。また、定常流動のエネルギ基本式から、次式が成り立つ。
【０１７０】
【数２１】
数１９式に（３１）式を代入して、
【０１７１】
【数２２】
あるいはＰ_１Ｖ_１＝ＲＴ_１から、
【０１７２】
【数２３】
先細ノズルでは、入口流速ｗ_１は出口流速ｗ_２に比べてきわめて小さいので省略すると、ノズル出口端の速度ｗ_２は次式で与えられる。
【０１７３】
【数２４】

ノズルの各断面を単位時間に流れる流量Ｑは、連続の式より一定であるから、
Ｑ＝Ａ_２×ｗ_２／ｖ_２＝ρｅ×Ａ_２×ｗ_２［ｋｇ／ｓｅｃ］ ・・・（３３）
である。また、ノズル内を流れる流体は理想気体で断熱変化するものとみなしているから、上記の（３１）式より、
【０１７４】
【数２５】

である。
【０１７５】
（３３）式に（３２）式と（３４）式を代入すると、
【０１７６】
【数２６】

（３５）式が先細ノズルの場合の流れの基礎式である。
【０１７７】
〈６〉先細ノズルの場合の排気圧予測式の検討
図４５を参照する。（１５）式から
Ｑａｓ０＋Ｑｆ＝Ｑｔｏｔａｌ［ｋｇ／ｓｅｃ］ ・・・（３６）
である。この（３６）式の右辺に、面積が縮小するノズルの流れの式である上記の（３５）式を適用して、
【０１７８】
【数２７】

の式を得る。
【０１７９】
ここで、タービン入口排気圧相当値Ｐｅｘｈｒを、
【０１８０】
【数２８】

とおくと、（３７ａ）式は

となるので、（３９）式をタービン入口排気圧相当値Ｐｅｘｈｒについて整理すると、次式が得られる。
【０１８１】
Ｐｅｘｈｒ＝（１／（２×ρｅ））×（κｅ／（κｅ−１））×｛（Ｑａｓ０＋Ｑｆ）／Ａｖｎｔ｝^２［Ｐａ］ ・・・（４０）
ここでも、排気密度ρｅの近似式である（１１）式を（４０）式に代入することにより、次の最終式を得る。
【０１８２】
Ｐｅｘｈｒ＝Ｋｐｅｘｈｎ×｛（Ｑａｓ０＋Ｑｆ）／Ａｖｎｔ｝^２×Ｔｅｘｈ［Ｐａ］・・・（４１ａ）
Ｋｐｅｘｈｎ＝（１／（２×ρｅ））×（κｅ／（κｅ−１）） ・・・（４１ｂ）
ただし、Ｋｐｅｘｈｎ：定数。
【０１８３】
さて、上記の数１２式のように、ノズルを通過するガスの流れを、流路面積が縮小する場合の流れであると仮定して排気圧を演算するものでは、標準状態（２９８Ｋ、０．１ＭＰａ）において排気圧の高い演算精度が得られるのであるが、実験によると、標準状態と異なる場合（たとえば高地、標準温度より温度が高い場合、湿度が標準状態と異なる場合など）に、排気圧の演算精度が低下することがわかっている。これは、数１２式が比重の変化を考慮してはいるが、まだ正確でないためと思われる。
【０１８４】
これに対して、ノズルを通過するガスの流れを、理想気体が断熱変化して流動する場合の流れであると仮定して求めた第４実施形態によれば、演算式により単位時間当たりの流量と圧力（つまり比重の変化）が正確に記述できているので、標準状態と異なる気圧や温度の状態においても、排気圧の高い演算精度が得られることになった。
【０１８５】
しかもマッチングしなければならない図４７の特性は、図示のように単純なものであるため、ほとんど計算だけで足り（マッチングの必要なし）、机上のみの計算でも排気圧の演算精度は高いのである（実験により確認している）。
【０１８６】
さて、ＥＧＲ装置と可変容量ターボチャージャをエンジンに備え、前述したモデル規範制御を用いたＥＧＲ制御と上記の過給圧制御とを行うとともに、手動変速機を備える車両を用いて発進加速を行ったとき、この車両においては発進加速直後に燃焼騒音とスモークが一時的に増えること、また加速途中の変速中には燃焼騒音が一時的に増えることを見い出した。そこで、これを解析するため、エンジン回転数を一定に保ったまま、燃料噴射量をステップ的に増大し、一定期間のあとにステップ的に燃料噴射量を減少させる実験を行った。このときの実験結果をモデル的に示したのが図４９である。同図において、燃料噴射量Ｑｆがステップ的に増加するｔ１のタイミングが発進加速時に、燃料噴射量Ｑｆがステップ的に減少するｔ２のタイミングが加速途中の変速中に対応する。
【０１８７】
なお、ｔ２のタイミングが加速途中の変速中に対応するのは次の理由からである。手動変速機を備える車両でたとえばギア位置を１速にした状態でアクセルペダルを踏み込んで加速（発進加速）を行い、そのあとギア位置を２速に入れ換えるにはそれまで踏み込んでいたアクセルペダルから足を離してクラッチペダルに移し、クラッチペダルを踏み込んでクラッチを切らなけれならない。このクラッチ切断の過程でアクセルペダル開度が減少し、このアクセルペダル開度の減少に対応して燃料供給量が減少するわけである。ただし、本発明は手動変速機を備える車両に限定されるものでなく、自動変速機を備える車両においてもｔ２のタイミングが加速途中の変速中に対応する。自動変速機を備える車両では、変速時に変速ショックを和らげるため、燃料噴射量を減量する指令が出るからである。
【０１８８】
以下、（１）発進加速直後と（２）加速途中の変速中に分けて解析する。
（１）発進加速直後：
負荷増大によって排気圧Ｐｅｘｈが応答よく上昇するのに対してターボチャージャ２の作動遅れによって吸気圧Ｐｍのほうは遅れて上昇するため、吸気圧Ｐｍと排気圧Ｐｅｘｈの差圧ΔＰ（＝Ｐｅｘｈ−Ｐｍ）が第５段目のように増大し、これによって実ＥＧＲ流量（数１３式により演算されるＱｅ）が、第３段目一点鎖線のように一時的に増加し、その後に増加後の燃料噴射量Ｑｆに対応するＥＧＲ流量へと一次遅れで収束する。このターボチャージャ２の作動遅れに伴う一時的なＥＧＲ流量の増加で低温予混合燃焼を維持できなくなる。燃焼状態が拡散燃焼を主体とする通常のディーゼル燃焼に移行し、燃焼騒音（１ｋＨｚバンドのＣＰＬ）とスモーク（ＩＳＦ）が増大するのである。
（２）加速途中の変速中：
負荷減少によって排気圧Ｐｅｘｈが応答よく減少するのに対してターボチャージャ２の作動遅れで吸気圧Ｐｍが遅れて減少するため、差圧ΔＰが第５段目のように減少し、これによって実ＥＧＲ流量が一時的に減少し、そのあとは減少後の燃料噴射量Ｑｆに対応するＥＧＲ流量へと一次遅れで収束する。このターボチャージャ２の作動遅れに伴う一時的なＥＧＲ流量の減少によっても燃焼状態が低温予混合燃焼から拡散燃焼を主体とする通常のディーゼル燃焼に移行してしまい、燃焼騒音が増大するのである。
【０１８９】
そこでコントロールユニット４１では、上記（１）の場合、つまり実過給圧（＝吸気圧）が目標過給圧より小さくかつ実ＥＧＲ流量が目標ＥＧＲ流量より大きい場合に、コモンレール圧力（燃料噴射圧力）を増大補正し、かつ噴射時期を遅角補正する。コモンレール圧力を増大させることで、燃料噴霧の微粒化を促進させてスモークを低減し、かつ噴射時期を遅角して着火遅れ期間を増大させることで、燃料と空気の混合を促進させた後に燃焼させ、これによって低温予混合燃焼を維持させ、燃焼騒音とスモークをともに抑制するのである。
【０１９０】
また、（２）の場合、つまり実過給圧が目標過給圧より大きくかつ実ＥＧＲ流量が目標ＥＧＲ流量より小さい場合に、コモンレール圧力を減少補正し、かつ噴射時期を遅角補正する。コモンレール圧力を減少させることで、燃料噴霧の微粒化を鈍化させて燃焼を緩慢にする（酸素量が十分なのでスモークは悪化しない）とともに、燃料噴射時期を遅角して着火遅れ期間を増大させることで、燃料と空気の混合を促進させた後に燃焼させることにより、この場合も低温予混合燃焼を維持させ、これによって燃焼騒音を抑制するのである。
【０１９１】
これをまとめたのが図５０の表図で、同図において右上欄が上記（１）の場合に、左下欄が上記（２）の場合に対応する。
【０１９２】
この場合、実過給圧と目標過給圧の比較結果および実ＥＧＲ流量と目標ＥＧＲ流量の比較結果の組み合わせによれば、次の（３）、（４）の場合が考えられるので、この（３）、（４）の場合に対しては、次のようにコモンレール圧力および噴射時期を制御する。
（３）実過給圧が目標過給圧より大きくかつ実ＥＧＲ流量が目標ＥＧＲ流量より大きい場合：
上記（２）の場合よりコモンレール圧力の減少補正量および噴射時期の遅角補正量を小さくする（図５０の右下欄に対応）。
（４）実過給圧が目標過給圧より小さくかつ実ＥＧＲ流量が目標ＥＧＲ流量より小さい場合：
上記（１）の場合よりコモンレール圧力の増大補正量および噴射時期の遅角補正量を小さくする（図５０の左上欄に対応）。
【０１９３】
ここで、（３）や（４）の場合は、実ＥＧＲ流量がオーバーシュートする応答の場合に対処するためのものである。たとえば、加速途中の変速中のＥＧＲ流量の変化を図５１にモデル的に示すと、実ＥＧＲ流量の通常の応答の場合には、上記（２）の場合のコモンレール圧力および噴射時期の制御で対処した。しかしながら、実ＥＧＲ流量が目標値を超えてオーバーシュートする応答（図ではゲイン過多の応答）の場合に、実ＥＧＲ流量が目標ＥＧＲ流量を上回るｔ３からｔ４までの区間においても（２）の場合の制御を行ったのでは、コモンレール圧力の減少補正量および燃料噴射時期の遅角補正量が大き過ぎることになるので、各補正量を小さくするため、（３）の場合の制御を行うのである。
【０１９４】
コントロールユニット４１で行われる上記の制御を次に詳述する。
【０１９５】
図５２はコモンレール圧力補正量Ｋ Ｐｒａｉｌの演算フローで、１０ｍｓｅｃ毎に実行する。
【０１９６】
ステップ１でエンジン回転数Ｎｅ、燃料噴射量Ｑｆ、シリンダ吸入ＥＧＲ量Ｑｅｃ（図１３により演算）、目標シリンダ吸入ＥＧＲ量Ｔｑｅｃ（図４１ステップ４で演算）、実過給圧Ｐｍ、目標過給圧ＴＰｍを読み込む。
【０１９７】
ここでは、実ＥＧＲ流量としてシリンダ吸入ＥＧＲ量Ｑｅｃを、その目標値として目標シリンダ吸入ＥＧＲ量Ｔｑｅｃを用いたが、実ＥＧＲ流量Ｑｅ（図３６により演算）とこれに対応する目標値としての要求ＥＧＲ流量Ｔｑｅ（図４１参照）を用いてもかまわない。なお、実過給圧Ｐｍ＿ｉｓｔ（センサ７２により検出）は吸気圧Ｐｍに等しいので、符号としてはＰｍを用いている。また、目標過給圧Ｐｍ＿ｓｏｌは図４のステップ４で演算されているが、Ｐｍという符号に対応してＴＰｍという符号を用いる。
【０１９８】
ステップ２ではシリンダ吸入ＥＧＲ量Ｑｅｃと目標シリンダ吸入ＥＧＲ量Ｔｑｅｃとの差ｄＱｅｃ、実過給圧Ｐｍと目標過給圧ＴＰｍとの差ｄＰｍを演算し、これらの差ｄＱｅｃ、ｄＰｍからステップ３で図５３を内容とする２つのテーブルのいずれかを検索することにより、コモンレール圧力補正量の基本値ＫＱＢ Ｐを演算する。
【０１９９】
また、エンジン回転数Ｎｅと燃料噴射量Ｑｆからステップ４で図５４を内容とするマップを検索することにより補正ゲインとしての運転条件反映係数ＫＱＣ Ｐを演算し、この係数ＫＱＣ Ｐと上記の基本値ＫＱＢ Ｐとを用いステップ５において、
【０２００】
【数１８】
Ｋ Ｐｒａｉｌ＝ＫＱＢ Ｐ×ＫＱＣ Ｐ
の式によりコモンレール圧力補正量Ｋ Ｐｒａｉｌを演算する。
【０２０１】
ここで、基本値ＫＱＢ Ｐは図５０で前述したコモンレール圧力補正量を与えるものである。すなわちＫＱＢ Ｐの値は、ＴＰｍ＞ＰｍかつＱｅｃ＞Ｔｑｅｃの場合（上記（１）の場合）に、ｄＱｅｃに比例して大きくなり（図５３右図参照）、ＴＰｍ＜ＰｍかつＱｅｃ＜Ｔｑｅｃの場合（上記（２）の場合）に、ｄＱｅｃに比例して負で大きくなる値である（図５３左図参照）。また、ＴＰｍ＞ＰｍかつＱｅｃ＜Ｔｑｅｃの場合（上記（３）の場合）に上記（２）の場合より直線の傾きを緩やかに（図５３右図参照）、同様にしてＴＰｍ＜ＰｍかつＱｅｃ＞Ｔｑｅｃの場合（上記（４）の場合）に上記（１）の場合より直線の傾きを緩やかにしている（図５３左図参照）。
【０２０２】
図５４においては、Ａが低温予混合燃焼域、Ｃが拡散燃焼を主体とする通常のディーゼル燃焼域、Ｂが両者の中間の領域で、燃焼騒音は特にＢの領域においいて顕著であることから、Ｂの領域でＫＱＣ Ｐを最大の１．０としている。
【０２０３】
Ａの領域でＫＱＣ Ｐ＝０．８としているのは、この領域ではＥＧＲ流量、過給圧とも応答が遅いため、補正ゲインが大きいとハンチングの可能性があるので、このハンチング防止のため補正ゲインを小さくしているものである。
【０２０４】
またＣの領域でＫＱＣ Ｐ＝０．６としているのは、次の理由からである。この領域はＥＧＲを中止しかつ高過給圧の領域であるため、原則的には補正する必要のない領域である。しかしながら、ＫＱＣ Ｐの値は、最終的には適合によって定めることになるので、Ｃでの値が必要な場合も生じる。そこで、Ｃの領域でも値を入れることができるようにしている（０．６は仮の値）。
【０２０５】
図５５は噴射時期補正量Ｋ ＩＴの演算フローである。
【０２０６】
ステップ１、２は図５２のステップ１、２と同じである（シリンダ吸入ＥＧＲ量Ｑｅｃと目標シリンダ吸入ＥＧＲ量Ｔｑｅｃとの差ｄＱｅｃ、実過給圧Ｐｍと目標過給圧ＴＰｍとの差ｄＰｍを演算する）。
【０２０７】
ステップ３ではこれらの差ｄＱｅｃ、ｄＰｍから図５６を内容とする２つのテーブルのいずれかを検索することにより、噴射時期補正量の基本値ＫＱＢ ＩＴを、またエンジン回転数Ｎｅと燃料噴射量Ｑｆからステップ４で図５７を内容とするマップを検索することにより、補正ゲインとしての運転条件反映係数ＫＱＣ ＩＴを演算し、これら運転条件反映係数ＫＱＣ ＩＴ、基本値ＫＱＢ ＩＴを用いステップ５において
【０２０８】
【数１９】
Ｋ ＩＴ＝ＫＱＢ ＩＴ×ＫＱＣ ＩＴ
の式により噴射時期補正量Ｋ＿ＩＴを演算する。
【０２０９】
ここで、基本値ＫＱＢ ＩＴは図５０で前述した料噴射時期の遅角補正量を与えるものである。すなわち、基本値ＫＱＢ ＩＴはＴＰｍ＞ＰｍかつＱｅｃ＞Ｔｑｅｃの場合（上記（１）の場合）に、ｄＱｅｃに比例してＫＱＢ ＩＴの値が負で大きくなり（図５６右図参照）、ＴＰｍ＜ＰｍかつＱｅｃ＜Ｔｑｅｃの場合（上記（２）の場合）に、ｄＱｅｃに比例してＫＱＢ ＩＴの値が負で大きくなる（図５６左図参照）値である。また、ＴＰｍ＞ＰｍかつＱｅｃ＜Ｔｑｅｃの場合（上記（３）の場合）に上記（２）の場合より直線の傾きを緩やかに（図５６右図参照）、同様にしてＴＰｍ＜ＰｍかつＱｅｃ＞Ｔｑｅｃの場合（上記（４）の場合）に上記（１）の場合より直線の傾きを緩やかにしている（図５６左図参照）。
【０２１０】
なお、基本値ＫＱＢ ＩＴを負の値で与えるのは、図６５のステップ７で後述するように補正量Ｋ ＩＴを目標噴射時期ＴＩＴ０に加算した値を目標噴射時期ＴＩＴ１として求めており、この場合の目標噴射時期は所定のクランク角位置から進角側に測った値（進角量）であるため、遅角補正するには補正量Ｋ ＩＴを負の値で与える必要があるからである。
【０２１１】
図５７においてＡの領域でＫＱＣ Ｐ＝０．８とし、またＣの領域でＫＱＣ Ｐ＝０．６としているのは図５４と同様の理由によるものである。
【０２１２】
図５８は目標コモンレール圧力ＴＰｒａｉｌの演算フローである。
【０２１３】
ここで、ステップ７において上記のコモンレール圧力補正量Ｋ Ｐｒａｉｌを用いてコモンレール圧力を補正している点が従来と異なる点で、残りは従来と同様である。
【０２１４】
詳細にはステップ１でエンジン回転数Ｎｅ、燃料噴射量Ｑｆ、大気圧（コンプレッサ入口圧）Ｐａ、冷却水温Ｔｗ、吸入新気温度Ｔａ（図１８により演算）、実コモンレール圧力Ｐｒａｉｌ（センサ３２により検出）に加えてコモンレール圧力補正量Ｋ Ｐｒａｉｌを読み込み、このうちエンジン回転数Ｎｅと燃料噴射量Ｑｆからステップ２において図５９を内容とするマップを検索することにより目標コモンレール圧力の基本値ＴＰｒａｉｌＢを、また冷却水温Ｔｗ、吸入新気温度Ｔａ、大気圧Ｐａからステップ３、４、５で図６０、図６１、図６２を内容とするテーブルを検索することにより、目標コモンレール圧力の水温補正係数ＫＰＴｗ、吸気温度補正係数ＫＰＴａ、大気圧補正係数ＫＰＰａを求め、ステップ６において
【０２１５】
【数２０】
ＴＰｒａｉｌ０＝ＴＰｒａｉｌＢ×ＫＰＴｗ×ＫＰＴａ×ＫＰＰａ
の式により目標コモンレール圧力ＴＰｒａｉｌ０を計算する。
【０２１６】
ここで、図６０に示したように低水温時に補正係数ＫＰＴｗの値を１．０より大きな値としているのは、低水温時に燃料温度が低くて燃料噴霧の状態が悪くなるので、これを防止するためである。図６１のように吸入新気温度Ｔａが低い場合に補正係数ＫＰＴａを１．０より大きな値としているのは、吸入新気温度Ｔａが低い場合に燃料噴霧が気化しにくくなるので、噴霧粒径を小さくするためである。図６２のように大気圧Ｐａが低い場合に補正係数ＫＰＰａを１．０より大きな値としているのは、大気圧Ｐａが低いと実圧縮比が低くなり着火しにくくなるので、噴霧粒径を小さくして着火しやすくするためである。
【０２１７】
ステップ７ではコモンレール圧力補正量Ｋ Ｐｒａｉｌを用いて
【０２１８】
【数２１】
ＴＰｒａｉｌ１＝ＴＰｒａｉｌ０×Ｋ Ｐｒａｉｌ
の式により上記の目標コモンレール圧力ＴＰｒａｉｌ０を補正し、補正後の目標コモンレール圧力を目標コモンレール圧力ＴＰｒａｉｌ１とおく。
【０２１９】
ステップ８では実コモンレール圧力Ｐｒａｉｌがこの目標コモンレール圧力ＴＰｒａｉｌ１と一致するようにＰＩ制御によりＰＩ補正量を演算し、ステップ９においてこのＰＩ補正量を目標コモンレール圧力ＴＰｒａｉｌ１に加算し、加算後の目標コモンレール圧力を目標コモンレール圧力ＴＰｒａｉｌ２とおく。
【０２２０】
ステップ１０ではエンジン回転数Ｎｅと燃料噴射量Ｑｆとから図６３、図６４のマップを検索して最大コモンレール圧力ＰｒａｉｌＭＡＸ、最小コモンレール圧力ＰｒａｉｌＭＩＮを求め、ＴＰｒａｉｌ２がこの最大値と最小値の間にあればＴＰｒａｉｌ２の値を、またＴＰｒａｉｌ２が最大コモンレール圧力ＰｒａｉｌＭＡＸを超える場合は最大コモンレール圧力ＰｒａｉｌＭＡＸを、ＴＰｒａｉｌ２が最小コモンレール圧力ＰｒａｉｌＭＩＮを下回る場合は、最小コモンレール圧力ＰｒａｉｌＭＩＮを目標コモンレール圧ＴＰｒａｉｌとして演算する。
【０２２１】
図６５は目標噴射時期ＴＩＴの演算フローである。ここで、ステップ７において噴射時期補正量Ｋ ＩＴを用いて目標噴射時期を補正している点が従来と異なる点で、残りは従来と同様である。
【０２２２】
詳細にはステップ１でエンジン回転数Ｎｅ、燃料噴射量Ｑｆ、大気圧Ｐａ、冷却水温Ｔｗ、吸入新気温度Ｔａ、噴射時期補正量Ｋ ＩＴを読み込み、このうちエンジン回転数Ｎｅと燃料噴射量Ｑｆからステップ２において図６６を内容とするマップを検索することにより目標噴射時期の基本値ＴＩＴＢを、また冷却水温Ｔｗ、吸入新気温度Ｔａ、大気圧Ｐａからステップ３、４、５で図６７、図６８、図６９を内容とするテーブルを検索することにより、目標噴射時期の水温補正係数ＫＩＴＴｗ、吸気温度補正係数ＫＩＴＴａ、大気圧補正係数ＫＩＴＰａを求め、ステップ６において
【０２２３】
【数２２】
ＴＩＴ０＝ＴＩＴＢ×ＫＩＴＴｗ×ＫＩＴＴａ×ＫＩＴＰａ
の式により目標噴射時期ＴＩＴ０を計算する。
【０２２４】
ここで、基本値ＴＩＴＢは、クランク角で圧縮上死点後の所定の範囲内で燃料噴射が開始されるように設定される。
【０２２５】
また、基本値ＴＩＴＢは、所定のクランク角位置から進角側に測った値（進角量）である。したがって、補正係数ＫＩＴＴｗ、ＫＩＴＴａ、ＫＩＴＰａが１．０より大きな値のとき噴射時期が進角される。図６７に示したように低水温時に補正係数ＫＩＴＴｗの値を１．０より大きな値としているのは、低水温時に燃料温度が低くて燃焼が遅れがちになるので、燃焼の中心を進角側にもってくるためである。図６８のように吸入新気温度Ｔａが低い場合に補正係数ＫＩＴＴａを１．０より大きな値とし、図６９のように大気圧Ｐａが低い場合に補正係数ＫＩＴＰａを１．０より大きな値としているのも、同様の理由からである。
【０２２６】
ステップ７では噴射時期補正量Ｋ ＩＴを用いて
【０２２７】
【数２３】
ＴＩＴ１＝ＴＩＴ０＋Ｋ ＩＴ
の式により目標噴射時期ＴＩＴ０を補正し、補正後の値を目標噴射時期ＴＩＴ１とする。噴射時期補正量Ｋ ＩＴは負の値であるため、数２３式により目標噴射時期が遅角補正される。
【０２２８】
ステップ８ではエンジン回転数Ｎｅと燃料噴射量Ｑｆとから図７０、図７１のマップを検索して最大噴射時期ＩＴＭＡＸ、最小噴射時期ＩＴＭＩＮを求め、目標噴射時期ＴＩＴ１が最大値と最小値の間にあればＴＩＴ１の値を、また目標噴射時期ＴＩＴ１が最大噴射時期ＩＴＭＡＸを超える場合は最大噴射時期ＩＴＭＡＸを、目標噴射時期ＴＩＴ１が最小噴射時期ＩＴＭＩＮを下回る場合は最小噴射時期ＩＴＭＩＮを目標燃料噴射時期ＴＩＴとして演算する。
【０２２９】
ここで、本実施形態の作用を図４９を参照しながら説明する。
【０２３０】
本実施形態によれば。ｔ１のタイミングより、実過給圧Ｐｍが目標過給圧ＴＰｍより小さくかつ実ＥＧＲ流量が目標ＥＧＲ流量より大きくなると、目標コモンレール圧力ＴＰｒａｉｌが増大補正されることで（図４９第６段目破線参照）、燃料噴霧の微粒化が促進されてスモークが低減し、かつ図４９第７段目破線のように目標噴射時期ＴＩＴが遅角補正され、着火遅れ期間が増大されることで、燃料と空気の混合が促進された後に燃焼する。つまり、低温予混合燃焼が実現されるように目標コモンレール圧力および目標噴射時期の補正を行って、発進加速時においても低温予混合燃焼を維持させ、これによって燃焼騒音とスモークをともに抑制できるのである（図４９下から１段目と２段目の一点鎖線参照）。
【０２３１】
また、本実施形態によれば、ｔ２のタイミングより実過給圧Ｐｍが目標過給圧ＴＰｍより大きくかつ実ＥＧＲ流量が目標ＥＧＲ流量より小さくなると、目標コモンレール圧力ＴＰｒａｉｌが減少補正されることで（図４９第６段目破線参照）、燃料噴霧の微粒化が鈍化し燃焼が緩慢になり（酸素量が十分なのでスモークは悪化しない）、かつ図４９第７段目破線のように目標噴射時期ＴＩＴが遅角補正され、着火遅れ期間が増大されることで、燃料と空気の混合が促進された後に燃焼する。この場合においても、低温予混合燃焼が実現されるように目標コモンレール圧力および目標噴射時期の補正を行って、発進加速途中の変速中においても低温予混合燃焼を維持させ、これによって燃焼騒音を抑制できる（図４９下から２段目の一点鎖線参照）。
【０２３２】
また、図示しないがｔ１のタイミング直後の一時的増加のあとに実ＥＧＲ流量がオーバーシュートする応答の場合に、実ＥＧＲ流量が目標ＥＧＲ流量を下回る区間においても上記（１）の場合の制御を行ったのでは、目標コモンレール圧力の増大補正量および目標噴射時期の遅角補正量が大き過ぎることになり、あるいはｔ２のタイミング直後の一時的減少のあと実ＥＧＲ流量がオーバーシュートする応答の場合（図５１参照）に、実ＥＧＲ流量が目標ＥＧＲ流量を上回る区間においても上記（２）の場合の制御を行ったのでは、目標コモンレール圧力の減少補正量および目標噴射時期の遅角補正量が大き過ぎることになるのであるが、この場合に本実施形態によれば上記（１）や（２）の場合より各補正量を小さくするので、各補正量が大き過ぎることがなく、これによって発進加速時や発進加速途中の変速中に一時的増加や一時的減少のあと実ＥＧＲ流量がオーバーシュートする応答の場合においても適切な値の補正量を与えることができる。
【０２３３】
これに対して、コモンレール式燃料噴射装置とＥＧＲ装置とを備え、広域空燃比センサにより検出される排気中の酸素濃度から実ＥＧＲ率を求め、この実ＥＧＲ率が目標ＥＧＲ率より高くなる加速時にコモンレール圧力を増大補正してスモークを低減し、この逆に実ＥＧＲ率が目標ＥＧＲ率より低いときにはコモンレール圧力を減少補正してＮＯｘ発生量を低減するようにした従来装置があり、この従来装置によれば、広域空燃比センサに検出遅れが生じるため、目標ＥＧＲ率と実ＥＧＲ率が最もずれる加速初期や変速初期に燃料噴射圧力の補正が遅れ、これによってスモークがかえって悪化したり、ディーゼルノック音が出たりしてしまうことがある。
【０２３４】
この場合に、実ＥＧＲ率と目標ＥＧＲ率の比較は、実ＥＧＲ流量と目標ＥＧＲ流量の比較でもよいので、実ＥＧＲ流量をモデル規範制御により演算することにより、過渡運転時においても応答性のよい実ＥＧＲ流量を得るようにした装置（特開平１０−３１８０４７号公報参照）と上記の従来装置とを組み合わせれば、実ＥＧＲ流量の応答遅れのタイミングに合わせて燃料噴射圧力の補正量を与えることができる。
【０２３５】
しかしながら、大量ＥＧＲによって空気過剰率が低下するのを避けるため、タービン内に可変ノズルを有する可変容量ターボチャージャを備える場合には、加速時や変速中にいわゆるターボラグにより実過給圧の応答遅れが生じ、これが実ＥＧＲ流量に影響するので、加速時の実ＥＧＲ流量の一時的増加を回避するため、たとえば可変ノズルのノズル開度を減少させることにより吸気流量を小さくして、空気過剰率を改善することも考えられるが、この方法だとかえって加速時の実ＥＧＲ流量の一時的増加を大きくしてしまう。
【０２３６】
これに対して本実施形態では、大量ＥＧＲが可能なＥＧＲ装置と、タービン内に可変ノズルを有する可変容量ターボチャージャとを備えるエンジンを対象として、実ＥＧＲ流量をモデル規範制御により過渡運転時においても応答よく演算しつつ、このようにして得られる実ＥＧＲ流量が目標ＥＧＲ流量より大きくかつ実過給圧が目標過給圧より小さい場合に目標噴射圧力を増大補正し、また、このようにして得られる実ＥＧＲ流量が目標ＥＧＲ流量より小さくかつ実過給圧が目標過給圧より大きい場合に目標噴射圧力を減少補正するようにしたので、実用運転域（低速時や低負荷時）でのエンジンの排気組成と運転性の双方を改善しつつ、加速時や加速途中の変速中に、ＥＧＲ弁の作動遅れに伴う実ＥＧＲ流量の応答遅れがあり、この応答遅れにさらにターボラグが影響する場合においても、実ＥＧＲ流量の応答遅れの位相に合わせた燃料噴射圧力の補正が可能となり、噴射圧力の補正精度を高めることができる。
【０２３７】
実施形態では、目標コモンレール圧力と目標噴射時期の両方を補正する場合で説明したが、目標コモンレール圧力だけを補正するようにしてもかまわない。
【０２３８】
実施形態では、コモンレール式燃料噴射装置（燃料噴射圧力制御装置）、ＥＧＲ装置および可変容量ターボチャージャを備え、所定の運転域（たとえば低中負荷域）において低温予混合燃焼を行わせるエンジンで説明したが、これに限られるものでなく、全運転域で拡散燃焼を主体とした通常のディーゼル燃焼を行うエンジンに対しても適用できる。
【０２３９】
実施形態ではコンプレッサ入口圧Ｐａを検出するセンサ７３を設けた場合で説明したが、上記ＥＧＲ装置と可変容量ターボチャージャとを備えるエンジンを搭載する車両が標準大気（やこれに近い大気）のもとで運転される限りにおいては、コンプレッサ入口圧センサは不要である。このときは、標準大気に対するＰａの値を設定してやれば済むからである。
【図面の簡単な説明】
【図１】第１実施形態の制御システム図。
【図２】コモンレール式燃料噴射装置のシステム図。
【図３】ＥＧＲ制御システム図。
【図４】可変ノズルアクチュエータに与える指令開度の演算を説明するためのフローチャート。
【図５】基本過給圧の特性図。
【図６】大気圧補正値の特性図。
【図７】基本開度の特性図。
【図８】大気圧補正値の特性図。
【図９】ＥＧＲ制御システムのブロック図。
【図１０】モデル規範制御におけるパラメータの演算順を示すフローチャート。
【図１１】サイクル処理を説明するためのフローチャート。
【図１２】シリンダ吸入新気量の演算を説明するためのフローチャート。
【図１３】シリンダ吸入ＥＧＲ量の演算を説明するためのフローチャート。
【図１４】体積効率相当値の演算を説明するためのフローチャート。
【図１５】空気密度の特性図。
【図１６】吸気圧の演算を説明するためのフローチャート。
【図１７】センサ出力電圧に対する圧力の特性図。
【図１８】吸気温度の演算を説明するためのフローチャート。
【図１９】吸気温度の車速補正値の特性図。
【図２０】吸気温度の吸気量補正値の特性図。
【図２１】シリンダ吸入ガス温度の演算を説明するためのフローチャート。
【図２２】燃料噴射量の演算を説明するためのフローチャート。
【図２３】基本燃料噴射量の特性図。
【図２４】最大噴射量の特性図。
【図２５】排気温度の演算を説明するためのフローチャート。
【図２６】排気温度基本値の特性図。
【図２７】吸気温度補正係数の特性図。
【図２８】排気圧補正係数の特性図。
【図２９】スワール補正係数の特性図。
【図３０】ノズル開度補正係数の特性図。
【図３１】ノズル有効面積相当値の演算を説明するためのフローチャート。
【図３２】摩擦損失の特性図。
【図３３】ノズル損失の特性図。
【図３４】排気圧の演算を説明するためのフローチャート。
【図３５】排気圧の実測値と予測値の相関を調べた特性図。
【図３６】ＥＧＲ流量の演算を説明するためのフローチャート。
【図３７】ＥＧＲ弁開口面積相当値の特性図。
【図３８】目標ＥＧＲ率の演算を説明するためのフローチャート。
【図３９】目標ＥＧＲ率基本値の特性図。
【図４０】目標ＥＧＲ率補正値の特性図。
【図４１】要求ＥＧＲ量の演算を説明するためのフローチャート。
【図４２】指令ＥＧＲ弁リフト量の演算を説明するためのフローチャート。
【図４３】ＥＧＲ弁目標リフト量の特性図。
【図４４】流路面積が縮小する流れのモデル図。
【図４５】吸排気系の力学的釣合いの検討に使用したモデル図。
【図４６】他の排気圧の演算を説明するためのフローチャート。
【図４７】排気圧Ｐｅｘｈ０の特性図。
【図４８】先細ノズルのモデル図。
【図４９】本実施形態の作用を説明するための波形図。
【図５０】本実施形態の制御内容を示す表図。
【図５１】加速途中の変速中の実ＥＧＲ流量の変化を示すモデル図。
【図５２】コモンレール圧力補正量の演算を説明するためのフローチャート。
【図５３】コモンレール圧力補正量の基本値の特性図。
【図５４】運転条件反映係数の特性図。
【図５５】噴射時期補正量の演算を説明するためのフローチャート。
【図５６】噴射時期補正量の基本値の特性図。
【図５７】運転条件反映係数の特性図。
【図５８】目標コモンレール圧力の演算を説明するためのフローチャート。
【図５９】目標コモンレール圧力基本値の特性図。
【図６０】水温補正係数の特性図。
【図６１】吸気温度補正係数の特性図。
【図６２】大気圧補正係数の特性図。
【図６３】最大コモンレール圧力の特性図。
【図６４】最小コモンレール圧力の特性図。
【図６５】目標噴射時期の演算を説明するためのフローチャート。
【図６６】目標噴射時期基本値の特性図。
【図６７】水温補正係数の特性図。
【図６８】吸気温度補正係数の特性図。
【図６９】大気圧補正係数の特性図。
【図７０】最大噴射時期の特性図。
【図７１】最小噴射時期の特性図。
【図７２】加速時にノズル開度を小さくした場合の波形図。
【図７３】第１の発明のクレーム対応図。
【符号の説明】
２ 可変容量ターボチャージャ
２ｄ 可変ノズル
１０ コモンレール式燃料噴射装置
１７ 燃料噴射弁
４１ コントロールユニット
５７ ＥＧＲ弁[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a control device for a diesel engine.
[0002]
[Prior art]
A common rail fuel injection device and an EGR device are provided, and fuel and air are mixed by increasing and correcting the common rail pressure (fuel injection pressure) during acceleration when the actual EGR rate detected by the EGR rate detection means is higher than the target EGR rate. However, when the actual EGR rate is lower than the target EGR rate, the premixed combustion rate is reduced by reducing the common rail pressure to reduce the amount of NOx generated. (See JP-A-9-242617).
[0003]
[Problems to be solved by the invention]
By the way, in the above conventional apparatus, the actual EGR rate is obtained from the oxygen concentration in the exhaust gas detected by the wide-range air-fuel ratio sensor.
[0004]
However, since the wide-range air-fuel ratio sensor is expensive and a detection delay occurs (the detection response time constant of the sensor is on the order of 1/10 seconds to several seconds), the target EGR rate and the actual EGR rate are most shifted from the initial stage of acceleration. And the correction of fuel injection pressure is delayed at the beginning of shifting. That is, when the correction amount of the fuel injection pressure must be given in accordance with the response delay timing of the actual EGR rate, the fuel injection pressure is detected when the error from the target EGR rate of the actual EGR rate is the largest due to the detection delay of the sensor. If the fuel injection pressure changes when the error of the actual EGR rate is reduced without changing the smoke, the smoke will be worsened or the diesel knock noise will be produced.
[0005]
In this case, the comparison between the actual EGR rate and the target EGR rate may be a comparison between the actual EGR flow rate and the target EGR flow rate. Therefore, by calculating the actual EGR flow rate by the model reference control, the response can be obtained even during transient operation. If a device (see Japanese Patent Application Laid-Open No. 10-318047) capable of obtaining a good actual EGR flow rate and the above-mentioned conventional device are combined, the response delay of the actual EGR flow rate can be reduced as long as the engine is not supercharged The correction amount of the fuel injection pressure can be given in accordance with the timing.
[0006]
However, in order to avoid a decrease in the excess air ratio due to a large amount of EGR, a variable capacity turbocharger having a variable nozzle in the turbine is provided, and when supercharging is performed in a practical operation range (low speed or low load). Is a response delay of the actual supercharging pressure caused by so-called turbo lag during acceleration or shifting, and this greatly affects the actual EGR flow rate. For example, in order to avoid a temporary increase in the actual EGR flow rate during acceleration, Although it is conceivable to reduce the intake air flow rate by reducing the nozzle opening and improve the excess air ratio, this method is rather rather a temporary increase in the actual EGR flow rate during acceleration.
[0007]
This will be described with reference to FIG. 72. The actual EGR flow rate is proportional to the differential pressure ΔP (= Pexh−Pm) between the exhaust pressure Pexh and the intake pressure Pm. In the figure, at the time of acceleration, the exhaust pressure Pexh first rises with good response, the EGR valve closes with a delay, and then the actual supercharging pressure (= intake pressure Pm) rises. In this case, if the nozzle opening is decreased, the rise of the exhaust pressure Pexh increases and the rise of the intake pressure Pm is delayed by an amount corresponding to the decrease in the nozzle opening. As a result, the differential pressure ΔP is increased as compared with the case where the nozzle opening is not decreased, and the actual EGR flow rate is increased.
[0008]
Therefore, the present invention is intended for an engine equipped with an EGR device capable of mass EGR and a variable displacement turbocharger, and calculates the actual EGR flow rate with good response even during transient operation by model reference control. When the EGR flow rate is larger than the target EGR flow rate and the actual boost pressure is smaller than the target boost pressure, the target injection pressure is increased and corrected.TheaccelerationSometimesEven when there is a response delay of the actual EGR flow rate due to the operation delay of the EGR valve, and the turbo lag further affects this response delay, the object is to improve the correction accuracy of the injection pressure.
[0009]
[Means for Solving the Problems]
As shown in FIG. 73, the first invention is
An EGR valve 81; a variable displacement turbocharger 82 having a variable nozzle in the turbine; means 83 for calculating a target injection pressure according to the engine load; and means 84 for calculating a target EGR flow rate according to the engine load. , Means 85 for controlling the EGR valve 81 so that the target EGR flow rate flows, means 86 for calculating the actual EGR flow rate by model reference control, and means 87 for calculating the target boost pressure corresponding to the engine load, The means 88 for controlling the variable nozzle opening so as to obtain the target supercharging pressure, the means 89 for detecting the actual supercharging pressure, the actual supercharging pressure and the target supercharging pressure are compared, The means 90 for comparing the actual EGR flow rate with the target EGR flow rate, and based on the comparison result, the actual supercharging pressure is lower than the target supercharging pressure and the actual EGR flow rate is higher than the target EGR flow rate. Includes a means 91 for increasing correcting the target injection pressure in case, and means 92 for controlling the fuel injection pressure so that the corrected target injection pressureIn addition, when the actual boost pressure is lower than the target boost pressure and the actual EGR flow rate is greater than the target EGR flow rate, the target injection timing is retarded from the comparison result..
[0010]
The second invention is a figure.73As shown, the EGR valve 81, the variable displacement turbocharger 82 having a variable nozzle in the turbine, the means 83 for calculating the target injection pressure according to the engine load, and the target EGR flow rate according to the engine load. A means 84 for calculating, a means 85 for controlling the EGR valve 81 so that the target EGR flow rate flows, a means 86 for calculating the actual EGR flow rate by model reference control, and a target boost pressure corresponding to the engine load. A means 87 for calculating, a means 88 for controlling the variable nozzle opening so as to obtain the target supercharging pressure, a means 89 for detecting the actual supercharging pressure, and the actual supercharging pressure and the target supercharging pressure. And a means 90 for comparing the actual EGR flow rate with the target EGR flow rate,
From these comparison results, the actual boost pressure is higher than the target boost pressure.LowAnd actual EGR flow rate is more than target EGR flow ratelargeIf the target injection pressure isIncreaseMeans to correct91And means for controlling the fuel injection pressure so as to be the corrected target injection pressure92WithAt the same time, when the actual boost pressure is lower than the target boost pressure and the actual EGR flow rate is smaller than the target EGR flow rate, the target injection pressure is corrected to be increased by a correction amount smaller than the target injection pressure increase correction amount. Do.
[0015]
First3In the invention of the firstOr secondIn the present invention, the actual EGR flow rate is calculated based on the differential pressure between the exhaust pressure Pexh and the intake pressure Pm (= actual boost pressure).
[0016]
First4In the invention of the3In the present invention, an intake air amount Qas0, a fuel injection amount Qf, an effective area equivalent value Avnt of the variable nozzle and an exhaust temperature Texh are detected, and the exhaust pressure Pexh is determined using these four elements.
Pexh = Kpexh × {(Qas0 + Qf) / Avnt)}²× Texh + Pa
Where Pexh: exhaust pressure,
Qas0: intake air volume,
Qf: fuel injection amount,
Avnt: Effective area equivalent value of variable nozzle,
Texh: exhaust temperature at the turbine inlet,
Pa: compressor inlet pressure,
Kpexh: constant,
It calculates by the formula of
[0017]
First5In the invention of the3In this invention, the intake air amount Qas0, the fuel injection amount Qf, the effective area equivalent value Avnt of the variable nozzle and the exhaust temperature Texh are detected, and these four elements are used to obtain the turbine inlet exhaust pressure equivalent value Pexhr.
Pexhr = Kpexhn × {(Qas0 + Qf) / Avnt)}²× Texh
Where Pexhr: turbine inlet exhaust pressure equivalent value,
Qas0: intake air volume,
Qf: fuel injection amount,
Avnt: Effective area equivalent value of variable nozzle,
Texh: exhaust temperature at the turbine inlet,
Pa: compressor inlet pressure,
Kpexhn: constant,
The exhaust pressure Pexh is calculated from the turbine inlet exhaust pressure equivalent value Pexhr and the compressor inlet pressure Pa.
[0018]
First6In the invention of the1In this invention, the target injection pressure and the target injection timing are corrected so as to realize low-temperature premixed combustion.
[0019]
【The invention's effect】
In order to avoid a decrease in the excess air ratio due to a large amount of EGR, when a variable capacity turbocharger having a variable nozzle is provided in the turbine, a response delay of the actual supercharging pressure occurs due to a so-called turbo lag during acceleration or gear shifting. Since the actual EGR flow rate is affected, in order to avoid a temporary increase in the actual EGR flow rate during acceleration, for example, if the nozzle opening of the variable nozzle is decreased, the temporary increase in the actual EGR flow rate during acceleration is rather large. Will be the first, second, second3In the present invention, for an engine equipped with an EGR device and a variable capacity turbocharger, the actual EGR flow rate is calculated with good response even during transient operation by model reference control.3According to the present invention, when the actual EGR flow rate obtained in this way is larger than the target EGR flow rate and the actual boost pressure is smaller than the target boost pressure, the target injection pressure is increased and compensated.correctAs a result, both the exhaust composition and operability of the engine in the practical operating range can be improved, and there is a response delay of the actual EGR flow rate due to the delay in operation of the EGR valve during acceleration or during shifting during acceleration. Even when the turbo lag further affects the response delay, the fuel injection pressure can be corrected in accordance with the response delay phase of the actual EGR flow rate, and the injection pressure correction accuracy can be improved.
[0020]
When the fuel injection pressure is increased when the actual EGR flow rate is larger than the target EGR flow rate, the ignition delay period is shortened due to the excessive actual EGR flow rate, and the ignition period is shortened due to the increase in the fuel injection pressure. Therefore, although the ignition delay period is almost the same as the appropriate EGR flow rate, the amount of fuel injected before ignition is as much as the fuel injection pressure rises, and this increased fuel burns at the same time as ignition. Combustion suddenly produces diesel knock (combustion noise)JarThe second1According to this invention, since the ignition timing is corrected to be retarded so that the ignition delay period increases, combustion noise can be suppressed.
[0021]
In the case of a response in which the actual EGR flow rate overshoots after a temporary increase immediately after acceleration, even in a section where the actual EGR flow rate is lower than the target EGR flow rate, the target injection pressure is the same as in the normal response without overshoot. If the correction is performed, the increase correction amount of the target injection pressure is too large.BecomeBut in this case2According to the invention, since the correction amount is made smaller than in the case of a normal response without overshoot, the correction amount is not too large, thereby accelerating.SometimesTemporary increaseAdditionalEven in the case of a response in which the actual EGR flow rate overshoots, an appropriate correction amount can be provided.
[0022]
First4This invention calculates the exhaust pressure on the assumption that the flow of gas passing through the turbine nozzle is a flow when the flow path area is reduced.4According to this invention, high calculation accuracy of the exhaust pressure can be obtained in the standard state.
[0023]
If the exhaust pressure is calculated assuming that the flow of gas passing through the turbine nozzle is a flow when the flow path area is reduced, it is different from the standard state (for example, when the altitude is higher than the high altitude and standard temperature, the humidity is in the standard state) The exhaust pressure calculation accuracy will decrease,5According to the invention, the flow rate and pressure per unit time (that is, the change in specific gravity) are calculated based on the calculation formula that assumes that the flow of the gas passing through the turbine nozzle is a flow when the ideal gas is adiabatically changed. Can be accurately described, so that a high calculation accuracy of the exhaust pressure can be obtained even in a state of atmospheric pressure or temperature different from the standard state. In addition, since the characteristic that must be matched in this case is simple, only calculation is sufficient (no need for matching), and the calculation accuracy of the exhaust pressure can be increased by calculation only on the desk.
[0024]
First6According to the invention, low temperature premixed combustion can be maintained even in a transient operation state in which the response delay of the actual EGR rate and the response of the supercharging pressure due to the delay in the operation of the turbocharger occur, thereby enabling smoke and combustion noise during acceleration. Further, combustion noise during gear shifting during acceleration can be further suppressed.
[0025]
DETAILED DESCRIPTION OF THE INVENTION
FIG. 1 is a schematic configuration diagram of a diesel engine.
[0026]
In the combustion of a diesel engine, the amount of NOx produced greatly depends on the combustion temperature, and it is effective to lower the combustion temperature relatively for the reduction. In the low temperature premixed combustion system, the oxygen concentration is reduced by an exhaust gas recirculation system (EGR), thereby realizing low temperature combustion. For this reason, the exhaust passage 53 and the intake passage 52 are connected by the EGR passage 54, and an EGR valve 57 is provided in the middle of the EGR passage 54 so that a part of the exhaust gas is recirculated into the intake air.
[0027]
The EGR valve 57 is driven by a step motor 57a that receives a control signal from the control unit 41, and an appropriate EGR rate is obtained according to the operating conditions of the engine. For example, the EGR rate is set to a maximum of 100 percent (the intake air flow rate and the EGR gas flow rate are the same amount) in a low rotation and low load range, and the EGR rate is decreased as the rotation speed and load increase. Since the exhaust gas temperature rises on the high load side, when a large amount of EGR gas is recirculated, the intake air temperature rises, which in turn raises the combustion temperature relatively, reducing the effect of reducing NOx, and igniting the injected fuel. The delay period is shortened and premixed combustion cannot be realized. For this reason, the EGR rate is decreased as the load becomes higher.
[0028]
As shown in FIG. 3, an intake throttle valve 56 that restricts intake air in two stages may be provided downstream of the air flow meter 55 for measuring the intake air amount. At this time, the recirculation amount of the exhaust gas flowing from the exhaust passage 53 to the intake passage 52 depends on the differential pressure between the suction negative pressure generated according to the opening of the intake throttle valve 56 and the exhaust pressure of the exhaust passage 53, and It is determined corresponding to the opening degree of the EGR valve 57 at that time.
[0029]
Here, the opening degree of the intake throttle valve 56 is controlled in two stages by a negative pressure actuator 56a, and a first pressure is introduced to the negative pressure actuator 56a via a first electromagnetic valve 61 from a vacuum pump (not shown). A negative pressure passage 62 and a second negative pressure passage 64 that similarly guides the negative pressure through the second electromagnetic valve 63 are connected to each other, and the intake throttle valve 56 is controlled by the negative pressure adjusted by the electromagnetic valves 61 and 62. Is controlled in two stages, and the suction negative pressure generated downstream thereof is controlled.
[0030]
For example, when the first electromagnetic valve 61 stops introducing negative pressure, introduces atmospheric pressure, and the second electromagnetic valve 63 introduces negative pressure, the negative pressure of the negative pressure actuator 56a is weak and the intake throttle is reduced. On the other hand, the opening degree of the valve 56 becomes relatively large. On the other hand, when the first electromagnetic valve 61 also introduces a negative pressure, the negative pressure is strong and the opening degree of the intake throttle valve 56 becomes small. When both the first and second electromagnetic valves 61 and 63 are introducing atmospheric pressure, the intake throttle valve 56 is held in the fully open position by the return spring.
[0031]
Thus, when the intake throttle valve 56 is provided, the control unit 41 controls the operations of the first and second electromagnetic valves 61 and 63 and the step motor 57a to control the exhaust gas recirculation amount. .
[0032]
An EGR gas cooling device 3 is provided in the middle of the EGR passage 54. This is formed around the EGR passage 54 and has a water jacket (not shown), in which a part of the engine cooling water is circulated, and the circulation amount of this cooling water is provided at the inlet of the cooling water. It can be adjusted by a flow control valve (not shown). The degree of cooling of the EGR gas increases as the opening degree of the flow control valve increases in accordance with a command from the control unit 41.
[0033]
A swirl control valve 4 is provided in the intake passage near the intake port of the engine. When the opening degree of the swirl control valve 4 is controlled by the control unit 41 and is closed in the low engine speed and low load range (the opening degree is decreased), the flow rate of the intake air sucked into the combustion chamber is increased and the combustion chamber is strong. A swirl is generated. However, when the swirl becomes strong, the heat exchange rate of the working gas in the cylinder increases, and the working gas temperature relatively decreases.
[0034]
The hollow combustion chamber formed in the piston is a large-diameter toroidal combustion chamber. In this, the piston cavity is formed in a cylindrical shape from the crown surface to the bottom of the piston without restricting the inlet, and a conical part is formed in the center of the bottom, and this conical part enters the piston cavity at the later stage of the compression stroke. In order not to give resistance to the swirl that flows while swirling, the mixing of air and fuel is further improved.
[0035]
Thus, the swirl generated by the swirl control valve 4 is diffused from the inside of the piston cavity to the outside of the cavity as the piston descends during the combustion process due to the cylindrical piston cavity that does not restrict the inlet. The swirl is maintained even outside the cavity.
[0036]
The exhaust passage 53 includes the variable capacity turbocharger 2 downstream of the branch point of the EGR passage 54. The turbocharger 2 is provided with a variable nozzle 2d driven by a step motor 2c at the scroll inlet of the exhaust turbine 2a. The variable nozzle 2d is controlled by the control unit 41, and the nozzle opening degree is controlled to increase the flow rate of the exhaust gas introduced into the exhaust turbine 2a on the low rotation side so that a predetermined supercharging pressure can be obtained from the low engine rotation range. On the high rotation side, the nozzle opening (full open state) is controlled so that exhaust is introduced into the exhaust turbine 2a without resistance. Further, the variable nozzle 2d is controlled to a nozzle opening degree at which a desired supercharging pressure is obtained depending on the operating conditions.
[0037]
In this embodiment, the nozzle opening degree of the variable nozzle 2d is explained by a method of driving by the step motor 2c. However, it is driven by a diaphragm actuator and an electromagnetic solenoid for adjusting the control negative pressure to the actuator or by a DC motor. A method may be used. Further, the nozzle opening may be feedback controlled based on a signal from the nozzle position sensor.
[0038]
The engine is provided with the common rail fuel injection device 10 shown in FIG.
[0039]
This is mainly composed of a fuel tank 11, a supply pump 14, a common rail (accumulation chamber) 16, and a fuel injection nozzle 17 provided for each cylinder. The high-pressure fuel generated in the high-pressure supply pump 14 is stored in the common rail 16, and the fuel is injected. By opening and closing the nozzle needle by the three-way valve 25 in the nozzle 17, the start and end of injection can be freely controlled. The fuel pressure in the common rail 16 is always controlled to the optimum value required by the engine by the pressure sensor 32 and the discharge amount control mechanism of the supply pump 14.
[0040]
Control of the fuel injection amount, fuel injection timing, common rail pressure (fuel injection pressure), and the like is performed by a control unit 41 constituted by a microprocessor. For this reason, the control unit 41 receives signals from the accelerator opening sensor 33, the sensor 34 for detecting the engine speed and the crank angle, the sensor 35 for cylinder discrimination, and the water temperature sensor 36. The control unit 41 calculates the target fuel injection amount Qf and the target injection timing according to the engine speed and the accelerator opening, and controls the on-time of the three-way valve 25 in the nozzle corresponding to the target fuel injection amount Qf. In addition, the ON timing of the three-way valve 25 is controlled corresponding to the target injection timing. Further, the fuel pressure of the common rail 16 is feedback-controlled through the discharge amount control mechanism of the supply pump 14 so that the common rail pressure detected by the pressure sensor 32 matches the target pressure.
[0041]
The target injection timing is delayed from the normal injection timing in order to realize low temperature premixed combustion. As will be described later, the fuel injection is set to start within a predetermined range after the compression top dead center at the crank angle. As a result, the ignition delay period of the injected fuel is lengthened, during which the fuel vaporization is promoted, and it is possible to ignite in a state sufficiently mixed with air. As a result, low-temperature premixed combustion is performed under a low oxygen concentration due to exhaust gas recirculation, and NOx can be reduced without increasing particulates.
[0042]
Reference numeral 1 denotes a NOx reduction catalyst (for example, a copper-based zeolite catalyst).
[0043]
From the viewpoint of supercharging pressure control, EGR control also physically plays a role of supercharging pressure control. That is, the supercharging pressure also changes by changing the EGR amount. On the contrary, if the supercharging pressure is changed, the exhaust pressure changes, so the EGR amount also changes, and the supercharging pressure and the EGR amount cannot be controlled independently. In addition, there is a disturbance in control of each other.
[0044]
Therefore, the supercharging pressure and the control negative pressure supplied to the EGR valve are selectively detected by the intake pressure sensor by time sharing, and the EGR amount is controlled and supercharged based on the control negative pressure and the supercharging pressure. Although techniques for controlling the pressure are disclosed, this technique particularly deteriorates control responsiveness during a transition.
[0045]
By the way, there are five variables: intake pressure (compressor outlet pressure) Pm, exhaust pressure (turbine inlet pressure) Pexh, atmospheric pressure (compressor inlet pressure) Pa, EGR valve effective area equivalent value Aegr, and variable nozzle effective area equivalent value Avnt. If it can be known, the exhaust amount Qexh and the EGR amount Qegr can be calculated. Of the five variables, variables other than the exhaust pressure are relatively easy to detect, but exhaust pressure is generally difficult to obtain for sensors with durability at high exhaust temperatures and oxidizing atmospheres, and in-vehicle sensors As expensive. In addition, it is difficult to obtain sufficient responsiveness to provide durability under the use conditions as described above. Therefore, in order to control the supercharging pressure and the EGR amount with high accuracy and without impairing responsiveness and stability, a means for estimating the exhaust pressure is required.
[0046]
For this reason, the control unit 41 uses the four elements of the intake air amount Qas0, the fuel injection amount Qf, the variable nozzle effective area equivalent value Avnt, and the exhaust temperature Texh to directly and simply calculate the exhaust pressure Pexh. Calculate (estimate) with the formula.
[0047]
Further, EGR control is performed using the estimated exhaust pressure Pexh. For example, the target EGR rate Megr is calculated according to the engine speed and load (see FIG. 38), the required EGR amount Tqe is calculated based on the target EGR rate Megr (see FIG. 41), and the estimated exhaust pressure is calculated. The required opening area Tav of the EGR valve 57 is calculated from the difference between Pexh and the intake pressure Pm and the required EGR amount Tqe (see FIG. 42), and the EGR valve opening is controlled so as to be the required opening area Tav.
[0048]
This control performed by the control unit 41 will be described in detail below.
[0049]
The supercharging pressure control and the EGR control, which will be described in detail below, have already been proposed by another application almost simultaneously with the present application.
[0050]
First, the supercharging pressure control will be described. FIG. 4 shows a calculation flow of the command opening of the variable nozzle 2d, which is executed every 10 msec. Note that the command opening calculation method shown in FIG. 4 is basically a known method.
[0051]
In Step 1, the rotational speed Ne, the fuel injection amount Qf, the compressor inlet pressure Pa, and the actual supercharging pressure Pm_ist are read.
[0052]
Here, the actual supercharging pressure Pm_ist is the same as an intake pressure (compressor outlet pressure) Pm, which will be described later in EGR control, and this intake pressure Pm is obtained by an intake pressure sensor 72 (see FIG. 1) provided in the collector 52a. The compressor inlet pressure Pa is detected by an atmospheric pressure sensor 73 (see FIG. 1) provided upstream of the air flow meter 55. The calculation of the fuel injection amount Qf will be described later.
[0053]
In step 2, the basic boost pressure MPM is searched by searching a map having the contents shown in FIG. 5 from the rotational speed Ne and the fuel injection amount Qf. In step 3, a table having the contents shown in FIG. 6 is searched from the compressor inlet pressure Pa. Thus, an atmospheric pressure correction value of the supercharging pressure is obtained, and a value obtained by multiplying the atmospheric pressure correction value by the basic supercharging pressure MPM in step 4 is calculated as the target supercharging pressure Pm_sol.
[0054]
In step 5, the PI correction amount STEP of the nozzle opening is performed by PI control so that the actual boost pressure Pm_ist matches the target boost pressure Pm_sol. ist is calculated.
[0055]
In step 6, the basic opening degree MSTEP of the variable nozzle is retrieved by searching a map having the contents shown in FIG. 7 from the rotational speed Ne and the fuel injection amount Qf. In step 7, a table having the contents shown in FIG. By searching, an atmospheric pressure correction value of the nozzle opening is obtained, and a value obtained by multiplying this correction value by the basic opening MSTEP is a target opening STEP in step 8. Calculate as sol.
[0056]
In step 9, a D (differential) correction amount is calculated from the actual boost pressure Pm_ist and the rotational speed Ne, and this and the above-described PI correction amount STEP. ist is the target opening STEP in step 10 The value added to sol is calculated as VNTstep1.
[0057]
In step 11, a predetermined map (not shown) is searched from the engine speed Ne and the actual supercharging pressure Pm_ist to obtain the upper and lower limit values of the limiter. The upper and lower limit values are calculated as the command opening VNTstep.
[0058]
The variable nozzle command opening VNTstep thus obtained is converted into a step number (a control amount to be given to the step motor 2c as a variable nozzle actuator) by searching a predetermined table (not shown). Step motor 2c is driven such that command opening degree VNTstep is obtained.
[0059]
Next, with respect to EGR control, a rough block diagram of the control is shown in FIG. 9, and detailed flowcharts and maps and tables used for the flow are shown in FIGS. 11 to 34 and FIGS.
[0060]
Here, the control method performed by the control unit 41 is model reference control (one of the controls using a model of a multivariable input control system). Therefore, the sensors other than the accelerator opening sensor 33, the crank angle sensors 34 and 35, and the water temperature sensor 36 are the air flow meter 55, the intake air temperature sensor 71 provided in the vicinity of the air flow meter 55, and the present embodiment. With the intake pressure sensor 72 alone, various parameters required for control (for example, exhaust pressure described later) are all predicted and calculated in the control unit 41. The image of the model reference control is that each block in FIG. 9 instantaneously performs an operation given to each block while exchanging parameters with surrounding blocks. In recent years, the theoretical analysis of model reference control has progressed rapidly, so that it can be applied to engine control, and it has been confirmed by experiments that it is at a level that is practically acceptable.
[0061]
More specifically, (1) sampling of sensor detection values such as the air flow meter 55 is performed at regular intervals (see steps 1 to 3 in FIG. 12, FIG. 16, FIG. 18), and (2) parameter calculation in model reference control. Basically, every time a Ref signal (crank angle reference position signal) is input (steps 4 to 7, FIG. 13, FIG. 14, FIG. 21, FIG. 22, FIG. 25, FIG. 31, FIG. 34, FIG. 36, FIG. 38, FIG. 41, FIG. 42), (3) The output to the final actuator is executed at regular intervals. In the following description, the job for each input of the Ref signal is described as a job for every fixed time (see FIG. 11).
[0062]
Further, the calculation of each parameter in the above (2) is performed in the order shown in FIG. In FIG. 10, it does not take a long time to perform all the processes, and all the processes are completed in an instant by the input of the Ref signal. In the figure, “n−1” after the symbol means the previous value (that is, the value calculated before the 1Ref signal).
[0063]
Hereinafter, the calculation of each parameter will be described in the order shown in FIG.
[0064]
The EGR control itself has already been disclosed in Japanese Patent Application No. 10-31460 (hereinafter referred to as “prior application device”).
[0065]
FIG. 11 shows a flow of cycle processing of the cylinder intake fresh air amount, the fuel injection amount, and the cylinder intake gas temperature. In step 1, the cylinder intake fresh air amount Qac, the fuel injection amount Qf, and the cylinder intake gas temperature Tn are read. The calculation of the cylinder intake fresh air amount Qac, the fuel injection amount Qf, and the cylinder intake gas temperature Tn will be described later with reference to FIGS. 12, 22, and 21, respectively.
[0066]
In step 2, using these Qac, Qf, and Tn, Qexh = Qac · Z^{-(CYLN # -1)}, Qf0 = Qf · Z^{-(CYLN # -2)}, Tn0 = Tn · Z^{-(CYLN # -1)}The cycle processing is performed according to the following equation, which performs correction based on the phase difference with respect to the read timing of the air flow meter 55. However, CYLN # is the number of cylinders. For example, in a 4-cylinder engine, fuel injection is 180 CA × (cylinder number−2) with respect to the air flow meter reading timing, and therefore, delay processing is performed by subtracting 2 from the number of cylinders.
[0067]
FIG. 12 is a flow for calculating the cylinder intake fresh air amount Qac.
[0068]
In step 1, the output voltage of the air flow meter (AMF) 55 is read, and in step 2, the intake air amount is calculated from the output voltage by table conversion. In Step 3, a weighted average process is performed on the calculated intake air amount in order to smooth out the influence of the intake pulsation.
[0069]
In Step 4, the engine speed Ne is read. In Step 5, the cylinder intake air amount (per intake stroke) Qac0 is calculated from the rotational speed Ne and the weighted average value Qas0 of the intake air amount.
[0070]
[Expression 1]
Qac0 = (Qas0 / Ne) × KCON #
Where KCON # is a constant,
Calculate with the following formula.
[0071]
In step 6, a delay process for n times of Qac0 is performed, and a value Qac0 · Z after the delay process is performed.^-NIs calculated as the cylinder fresh air amount (per intake stroke) Qacn at the collector 52a inlet. This takes into account the delay of the intake air from the air flow meter 55 to the collector 52a inlet.
[0072]
In step 7, the volume ratio Kvol and the previous value Kin of the volume efficiency equivalent value_n-1From the above-mentioned cylinder fresh air quantity Qacn at the collector 52a inlet
[0073]
[Expression 2]
Qac = Qac_n-1× (1-Kvol × Kin_n-1) + Qacn × Kvol × Kin_n-1
However, Qac_n-1: The previous value of Qac,
Kin_n-1: The previous value of Kin,
The cylinder intake fresh air amount (per intake stroke) Qac is obtained by performing a delay process according to the following equation. This takes into account the delay in the intake air from the collector 52a inlet to the cylinder.
[0074]
FIG. 13 is a flow for calculating the cylinder intake EGR amount Qec.
[0075]
The contents of this calculation are the same as the calculation method of the cylinder intake fresh air amount Qac shown in FIG. Qe which is the previous value of the EGR (flow) amount Qe obtained in step 1 as described later (see FIG. 36)._n-1And in step 2, the engine speed Ne is read.
[0076]
In step 4, Qe_n-1, Ne, and constant KCON #, the cylinder intake EGR amount (per intake stroke) Qecn at the collector 52a inlet
[0077]
[Equation 3]
Qecn = (Qe_n-1/ Ne) x KCON #
Where KCON # is a constant,
Calculate with the following formula. Further, in step 5, the value Qecn at the collector inlet 52a, the volume ratio Kvol, and the previous value Kin of the volume efficiency equivalent value are obtained._n-1Using,
[0078]
[Expression 4]
Qec = Qec_n-1× (1-Kvol × Kin_n-1) + Qecn × Kvol × Kin_n-1
However, Qec_n-1: The previous value of Qec,
Kin_n-1: The previous value of Kin,
The cylinder intake EGR amount (per intake stroke) Qec is calculated by performing a delay process according to the following equation. This takes into account the delay of EGR gas from the collector 52a inlet to the cylinder.
[0079]
In the prior application device, the weighted average processing is performed on the EGR amount Qe in order to smooth the influence of exhaust pulsation. However, in this embodiment, the weighted average processing on Qe is not performed. This is due to the following reason. Although the weighted average processing value of Qe is used to calculate the influence of exhaust pulsation, the cylinder intake EGR amount Qec including the error associated with the weighted average is calculated. Therefore, in the present embodiment, the calculation accuracy of Qec is increased as much as possible by calculating Qec with Qe having pulsation.
[0080]
FIG. 14 is a flow for calculating the volumetric efficiency equivalent value Kin.
[0081]
In step 1, the cylinder intake fresh air amount Qac, the cylinder intake EGR amount Qec, the intake pressure Pm, and the previous value of the intake gas temperature, Tn_n-1Of which Pm and Tn_n-115 to obtain a gas density ROUqcyl by searching a map having the contents shown in FIG. 15 in step 2, and in step 3 using this gas density ROUqcyl and cylinder gas weight Qcyl (= Qac + Qec).
[0082]
[Equation 5]
Kin = Qcyl / (Vc / ROUqcyl)
Vc: 1 cylinder volume,
The volume efficiency equivalent value Kin is calculated by the following formula (volume efficiency definition formula).
[0083]
Here, the calculation method of the volumetric efficiency equivalent value Kin is different from that of the prior application device (simpler than the prior application device). This is because the intake air pressure sensor 72 is added in the present embodiment, and the volumetric efficiency can be calculated from the definition formula using this sensor detection value. Thereby, in this embodiment, the adaptation man-hour can be reduced about the calculation of volumetric efficiency.
[0084]
FIG. 16 is a flowchart for calculating (detecting) the intake pressure (in the collector).
[0085]
In step 1, the output voltage Pm_v of the intake pressure sensor 72 is read, and in step 2, a table having the contents shown in FIG. 17 is retrieved from the output voltage Pm_v to convert it into pressure Pm_0. An average process is performed, and the weighted average value Pm1 is calculated as the intake pressure Pm.
[0086]
Unlike the prior application device in which the intake pressure sensor is not provided, in the present embodiment, since the intake pressure sensor is provided, the calculation of the intake pressure Pm is simplified.
[0087]
Here, the reason why the intake pressure sensor is newly added is as follows. While the turbocharger was not a variable displacement type in the prior application device, the turbocharger of this embodiment is a variable displacement type, so the nozzle opening is newly added as an unknown number (degree of freedom). The unknown has increased by one. Therefore, in order to make the unknown the same as the prior application device, an intake pressure sensor 72 is provided (in the prior application device, the intake pressure is also an unknown, but in this embodiment, the intake pressure is not an unknown).
[0088]
FIG. 18 is a flow for calculating the intake fresh air temperature Ta.
[0089]
In step 1, the output voltage Ta_v of the intake air temperature sensor 71 is read, and in step 2, a table having the same characteristics as in FIG.
[0090]
In step 3, it is determined whether the intake air temperature sensor 71 is mounted on the upstream side or the downstream side of the intercooler 3.
[0091]
As shown in FIG. 1, when the intake air temperature sensor 71 is on the upstream side of the intercooler 3, the routine proceeds to step 4 where Pm, which is the previous value of the intake air pressure, is reached._n-1Based on the pressure correction coefficient Ktmpi, Ktmpi = Pm_n-1X Calculated from the equation of PA #. However, PA # is a constant.
[0092]
In step 5, the intake fresh air temperature Ta at the inlet of the collector 52a is determined based on the pressure correction coefficient Ktmpi.
[0093]
[Formula 6]
Ta = Ta0 × Ktmpi + TOFF #
Where TOFF # is a constant
Calculated by the formula (approximate formula) This calculation is a temperature change prediction calculation based on the laws of thermodynamics.
[0094]
The intake air temperature may be corrected by the vehicle speed, the intake air amount, or the like. At this time, a table having the characteristics shown in FIGS. 19 and 20 is created in advance, and each table is searched from the vehicle speed and the intake air amount (Qas0), whereby the intake air temperature correction value Kvsp, An intake air amount correction value Kqa of temperature is obtained, and instead of the above equation 7,
[0095]
[Expression 7]
Ta = Kvsp × Kqa × Ta0 × Ktmpi + TOFF #
The intake fresh air temperature Ta may be obtained by the following equation.
[0096]
On the other hand, if an intake air temperature sensor is mounted on the downstream side of the intercooler 3, since both the temperature rise due to supercharging and the temperature drop due to the intercooler have already been factored in, the process proceeds to step 6 to set the value of Ta0. After the intake fresh air temperature Ta is set as it is, the processing is terminated.
[0097]
FIG. 21 is a flow for calculating the cylinder intake gas temperature Tn. In Step 1, Texh is the previous value of the cylinder intake fresh air amount Qac, the intake fresh air temperature Ta, the cylinder intake EGR amount Qec, and the exhaust temperature._n-1Of these, in step 2, the previous value Texh of the exhaust temperature_n-1Is multiplied by the exhaust gas temperature drop coefficient Ktlos in the EGR passage 54 to calculate the cylinder intake EGR gas temperature Te.
[0098]
[Equation 8]
Tn = (Qac × Ta + Qec × Te) / (Qac + Qec)
The average temperature of the cylinder intake fresh air and the cylinder intake EGR gas is obtained by the following formula, and this is set as the cylinder intake temperature Tn.
[0099]
FIG. 22 is a flow for calculating the fuel injection amount Qf. In step 1, the engine speed Ne and the control lever opening degree CL (determined by the accelerator pedal opening degree) CL are read. In step 2, the map containing FIG. 23 is searched from these Ne and CL to obtain the basic fuel injection amount Mqdrv. .
[0100]
In step 3, various corrections are made on the basic fuel injection amount based on the engine coolant temperature, etc., and the fuel injection is further performed on the corrected value Qf 1 on the basis of the map containing FIG. 24 in step 4. The amount is limited by the maximum value Qf1MAX, and the value after the limitation is calculated as the fuel injection amount Qf.
[0101]
FIG. 25 is a flow for calculating the exhaust temperature Texh. In steps 1 and 2, the cycle processing value Qf0 of the fuel injection amount and the cycle processing value Tn0 of the cylinder intake gas temperature are read. Further, in step 3, Pexh which is the previous value of the exhaust pressure._n-1Is read.
[0102]
In step 4, the exhaust temperature basic value Texhb is obtained by searching a table having the contents shown in FIG. 26 from the cycle processing value Qf0 of the fuel injection amount.
[0103]
In step 5, the intake air temperature correction coefficient Ktexh1 of the exhaust gas temperature is calculated from the cycle processing value Tn0 of the intake gas temperature, and Ktexh1 = (Tn0 / TA #).^{KN #}(Where TA # and KN # are constants), and in step 6, the exhaust pressure correction coefficient Ktexh2 of the exhaust temperature is set to the previous value Pexh of the exhaust pressure._n-1To Ktexh2 = (Pexh_n-1/ PA #)^{(# Ke-1) / # Ke}Where PA # and #Ke are constants, respectively. These two correction coefficients Ktexh1 and Ktexh2 may be obtained by table search (see FIGS. 27 and 28).
[0104]
Next, in step 7, the exhaust temperature swirl correction coefficient Ktexh3 is determined by searching a table having the contents shown in FIG. 29 from the swirl valve opening position (two positions of fully open and fully closed) and the engine speed Ne. In FIG. 8, the nozzle opening correction coefficient Ktexh4 of the exhaust temperature is obtained by searching a map having the contents shown in FIG. 30 from the command opening VNTstep and the exhaust amount Qexh.
[0105]
In step 9, the exhaust gas temperature Texh is calculated by multiplying the exhaust gas basic value Texhb by the four correction coefficients Ktexh1, Ktexh2, Ktexh3, and Ktexh4.
[0106]
Here, in this embodiment, since two correction coefficients Ktexh3 and Ktexh4 that are not present in the prior application apparatus are newly introduced, the calculation accuracy of the exhaust gas temperature Texh is improved in this embodiment. The reason for improving the calculation accuracy of the exhaust gas temperature Texh is as follows. As will be described later with reference to the flow of FIG. 34, the exhaust temperature Texh is used for calculating the exhaust pressure Pexh. Therefore, improvement in the calculation accuracy of the exhaust temperature Texh leads to improvement in the calculation accuracy of the exhaust pressure Pexh. Therefore, in order to improve the calculation accuracy of the exhaust pressure Pexh, two correction coefficients Ktexh3 and Ktexh4 are newly introduced. .
[0107]
The process of FIG. 25 approximates the following expression derived from the thermodynamic expression.
[0108]
[Equation 9]

FIG. 31 is a calculation flow of the effective area equivalent value Avnt of the variable nozzle 2d. In step 1, the command opening VNTstep, the total exhaust weight Qtotal (= Qas0 + Qf), and the exhaust temperature Texh are read.
[0109]
Of these, in step 2 from the total exhaust weight Qtotal and the exhaust temperature Texh
[0110]
[Expression 10]
Wexh = Qtotal × Texh / Tstd [m³/ Sec]
Where Tstd: standard atmospheric temperature,
The exhaust gas flow rate equivalent value Wexh is calculated by the following formula.
[0111]
In step 3, a friction loss ξfric is calculated by searching a table having the contents shown in FIG. 32 from a value obtained by taking the square root of the exhaust gas flow velocity equivalent value Wexh. In step 4, a map having the contents shown in FIG. 33 is searched from the command opening degree VNTstep and the total gas weight Qtotal to calculate the nozzle loss ξconv. Then, these two losses ξfric and ξconv are multiplied by the command opening VNTstep in step 5, that is,
[0112]
[Expression 11]
Avnt = VNTstep × ξfric × ξconv
The effective area equivalent value Avnt of the variable nozzle is calculated by the following equation.
[0113]
FIG. 34 is a flowchart for calculating the exhaust pressure (turbine inlet pressure) Pexh.
[0114]
In step 1, the weighted average value Qas0 of the intake air amount, the fuel injection amount Qf, the effective area equivalent value Avnt, the exhaust gas temperature Texh, and the atmospheric pressure (compressor inlet pressure) Pa are read. Using these parameters, in step 2,
[0115]
[Expression 12]
Pexh0 = Kpexh × {(Qas0 + Qfuel) / Avnt}²× Texh + Pa
Where Kpexh is a constant,
The exhaust pressure Pexh0 is calculated by the following equation, and a weighted average process is performed on the exhaust pressure in step 3, and the weighted average value is obtained as the exhaust pressure Pexh.
[0116]
Here, each calculation method of the effective area equivalent value Avnt and the exhaust pressure Pexh0 is obtained as follows.
[0117]
<1> Basic equation of flow when the channel area is reduced
Consider an ideal fluid flowing in a pipe whose sectional area gradually decreases as shown in FIG.
[0118]
When the fluid pressure, flow velocity, area, and specific gravity are P, w, A, and ρ, respectively, the inlet is subscript 1, the outlet is subscript 2, and Bernoulli's theorem is applied to the inlet and outlet cross sections.
w₁ ²/ 2 + P₁/ Ρ = w₂ ²/ 2 + P₂/ Ρ (1a)
From the continuous formula
A₁× w₁= A₂× w₂                                                                        ... (1b)
Therefore, from both equations, w₁If you delete
w₂= 1 / {1- (A₂/ A₁)²}^1/2× {2 (P₁-P₂) / Ρ}^1/2[M / sec] (2)
Since the flow rate Q flowing per unit time is constant from the continuous equation,

It can be expressed by the following formula.
[0119]
1 / {1- (A on the right side of equation (3)₂/ A₁)²}^1/2Is the efficiency ηn, the following basic equation of flow is obtained.
[0120]
Q = ηn × A₂× {2ρ × (P₁-P₂)}^1/2                            ... (4)
<2> State equation of turbocharger
Next, the relationship between the compressor 2b and the work balance in the turbocharger 2 is examined. The symbols used below are as shown in FIG.
[0121]
The effective power Lc of the compressor 2b is
Lc = Qas0 × Wc / ηc [W] (5)
However, Qas0: intake fresh air weight flow [kg / sec],
Wc: Compressor theoretical work [J / kg],
ηc: Compressor efficiency equivalent value.
[0122]
The effective power Lt of the turbine 2a is
Lt = ηt × Qtotal × Wt [W] (6)
Where Qtotal: total exhaust weight flow rate [kg / sec]
Wt: Turbine theoretical work [J / kg],
ηt: turbine efficiency equivalent value.
[0123]
Since the turbine 2a and the compressor 2b are directly connected via a shaft, if the actual powers Lc and Lt of the compressor 2b and the turbine 2a are equal (bearing friction is included in the efficiency), the state equation of the turbocharger 2 As follows.
[0124]

<3> Exhaust pressure prediction formula when the flow path area is reduced
Applying the above equation (4) to the left side of the equation (7),

Where Avnt: equivalent value of effective area of variable nozzle,
Pexh: Exhaust pressure,
Pa: Atmospheric pressure equivalent value,
ρe: exhaust density,
VNTstep: command opening,
ηn: efficiency (loss),
To get the formula
[0125]
When formula (8a) is arranged for the exhaust pressure Pexh,

Here, according to the theoretical formula, the exhaust density ρe is
ρe = ρstd × (Ta / Texh) × (Pexh / Pa) (10)
However, ρstd: density of standard atmosphere (≈1.1679 g / cm³),
Ta: compressor inlet temperature,
Texh: exhaust temperature,
Pexh: Exhaust pressure,
Pa: atmospheric pressure,
However, in this theoretical formula, the exhaust pressure Pexh is used to obtain the exhaust density ρe, so it is not good.
ρe≈ρstd × (Ta / Texh) = Tstd / Texh (11)
However, Tstd: temperature of standard atmosphere (≈298.15K),
The approximate expression is used. The reason why it can be approximated is that if the exhaust pressure Pexh increases, the exhaust temperature Texh also increases, so that the change in the exhaust pressure Pexh can be considered to be included in the exhaust temperature Texh.
[0126]
Therefore, the following formula is obtained by substituting the formula (11) into the formula (9).
[0127]

Where Kpexh: constant.
[0128]
Here, the compressor theoretical work Wc and the turbine theoretical work Wt on the right side of the equation (12a) are given by the following equations.
[0129]
[Formula 13]

Now, from the equation (12a), an arithmetic expression for the exhaust pressure Pexh was obtained. Η in the equation (12a)
The calculation of c, ηt, Wc, and Wt is complicated (requires the ability of the ECU), and the expression (14) needs to know the exhaust pressure Pexh to be obtained from now, so further consideration will be given.
[0130]
Now, there is the following relationship between the total exhaust weight Qtotal, the intake fresh air amount Qas0, and the fuel injection amount Qf (all units are [kg / sec]).
[0131]
Qtotal = Qas0 + Qf (15)
Applying the above equation (4) to the left side of the equation (15),
Avnt × {2 × ρe × (Pexh-Pa)}^1/2= Qas0 + Qf (16a)
Avnt = ηn × VNTstep (16b)
When the both sides of the equation (16a) are squared and the exhaust pressure Pexh is arranged, the following equation is obtained.
[0132]
Pexh = {(Qas0 + Qf) / Avnt}²× (1 / ρe) + Pa (17)
Again, the following final expression is obtained by substituting Expression (11), which is an approximate expression of the exhaust density ρe, into Expression (17).
[0133]

Where Kpexh: constant.
[0134]
The expression (18a) is equivalent to the above expression (12a), and the calculation formula of the exhaust pressure Pexh by the expression (18a) includes the ratio of theoretical work (Wc / Wt) of the compressor 2b and the turbine 2a and the respective efficiency. The product (ηc × ηt) is included, and if the equation (18a) is used, the theoretical work Wc, Wt and the efficiency ηc, ηt of the turbocharger 2 are considered even if they are unknown. Therefore, the efficiency ηn of the gas flowing through the variable nozzle 2d can be obtained after that.
[0135]
<4> Efficiency of gas flowing through the nozzle ηn
The effective area equivalent value Avnt of the variable nozzle 2d including the efficiency ηn is given by the above equations (8b) and (16b), and the efficiency ηn can be expressed by the following equation.
[0136]

Where ξconv: nozzle loss,
ξfric: friction loss.
[0137]
In the equation (19), the nozzle loss ξconv is a loss determined for each opening degree of the nozzle.₂/ A₁)²}^1/2Becomes efficient.
[0138]
However, if the change in flow velocity is large, 1 / {1- (A₂/ A₁)²}^1/2If this value is regarded as the nozzle loss ξ conv as it is, it often does not match the actual nozzle loss, so it is described by providing a table of efficiency with respect to the nozzle opening (see FIG. 33).
[0139]
Further, the friction loss ξfric in the equation (19) is established by the Hagen-Poiseille equation when the flow inside the nozzle is regarded as a laminar flow, and the friction loss ξfric is proportional to the square root of the flow velocity. there,
Wexh = Qtotal / ρe (20)
The volume flow equivalent value Wexh is calculated by the following equation, and the square root of the volume flow equivalent value Wexh is set as the exhaust flow velocity, thereby searching for the friction loss ξfric (see FIG. 32).
[0140]
Again, substituting equation (11), which is an approximate expression of exhaust density ρe, into equation (20),

In this way, the nozzle effective area equivalent value Avnt is calculated by the equation (19), and the exhaust pressure Pexh is predicted by the equations (18a) and (18b) using Qas0, Qf, Texh, Pa in addition to this Avnt. That is why. FIG. 35 shows the experimental results of examining the correlation between the actual measured value and the predicted value of the exhaust pressure. From the figure, it can be seen that the predicted value is sufficiently accurate.
[0141]
Next, FIG. 36 is a flow for calculating the EGR (flow) amount Qe. In step 1, the above-described intake pressure Pm, exhaust pressure Pexh, and EGR valve actual lift amount Lifts as the EGR valve actual opening are read. Alternatively, the target EGR valve lift amount may be used if the actual EGR valve lift amount is uniquely determined by giving the target value as in the step motor.
[0142]
In step 2, a table having the contents shown in FIG. 37 is retrieved from the EGR valve actual lift amount Lifts to obtain an opening area equivalent value Ave of the EGR valve 57.
[0143]
In step 3, the EGR flow rate Qe is calculated from the intake pressure Pm, the exhaust pressure Pexh, and the opening area equivalent value Ave of the EGR valve 57.
[0144]
[Expression 14]
Qe = Ave × {(Pexh−Pm) × KR #}^1/2
However, KR #: correction coefficient (constant),
Calculate with the following formula.
[0145]
FIG. 38 is a flow for calculating the target EGR rate Megr. In step 1, the engine speed Ne, the fuel injection amount Qf, and the cylinder intake gas temperature Tn are read, and a map having the contents shown in FIG. 39 is retrieved from Ne and Qf to obtain the target EGR rate basic value Megr0. In step 3, a table having the contents shown in FIG. 40 is retrieved from the cylinder intake gas temperature Tn to obtain a target EGR rate correction value Hegr, and the target EGR rate correction value Hegr is multiplied by the target EGR rate basic value Megr0 to obtain the target EGR rate. Calculate Megr.
[0146]
FIG. 41 is a calculation flow of the required EGR (flow) amount Tqe. In step 1, the engine speed Ne, the target EGR rate Megr, the cylinder intake fresh air amount Qac, and the cycle processing value Qf0 of the fuel injection amount are read. To calculate the target inhalation EGR amount Mqec.
[0147]
In step 3, Kin × Kvol is used as a weighted average coefficient for the target inhalation EGR amount Mqec.
[0148]
[Expression 15]
Rqec = Rqec_n-1× (1-Kin × Kvol) + Mqec × Kin × Kvol
However, Rqec_n-1: The previous value of Rqec,
An intermediate processing value (weighted average value) Rqec is calculated by the following equation, and the intermediate processing value Rqec and the above-described target inhalation EGR amount Mqec are used in step 4.
[0149]
[Expression 16]
Tqec = Mqec × GKQEC + Rqec_n-1× (1-GKQEC)
However, Rqec_n-1: The previous value of Rqec,
GKQEC: Lead compensation gain,
The target cylinder intake EGR amount Tqec is obtained by performing the advance processing according to the following equation. Since there is a delay in the intake system with respect to the required value (that is, a delay corresponding to the capacity of the EGR valve 57 → the collector 52a → the intake manifold → the intake valve), in steps 3 and 4, advance processing for this delay is performed.
[0150]
In step 5, from this target cylinder suction EGR amount Tqec,
[0151]
[Expression 17]
Tqe = (Tqec / Ne) × KCON #
Where KCON # is a constant,
The required EGR amount Tqe is calculated by performing unit conversion (per cylinder → per unit time) by the following formula.
[0152]
FIG. 42 is a flow for calculating the command EGR valve lift amount Liftt as the command EGR valve opening degree. In step 1, the intake pressure Pm, the exhaust pressure Pexh, and the required EGR amount Tqe are read. In step 2, the required opening area Tav of the EGR valve 57 is set to
[0153]
[Expression 18]
Tav = Tqe / {(Pexh−Pm) × KR #}^1/2
However, KR #: correction coefficient (constant),
The following formula (the law of fluid mechanics) is used.
[0154]
In step 3, a table having the contents shown in FIG. 43 is retrieved from the required opening area Tav of the EGR valve 57 to obtain the EGR valve target lift amount Mlift as the target EGR valve opening, and the step is performed with respect to the target lift amount Mlift. 4, the advance process for the operation delay of the EGR valve 57 is performed, and the value after the advance process is obtained as the command EGR valve lift amount Liftt.
[0155]
The command EGR valve lift amount Liftt thus obtained is output to the step motor 57a by a flow (not shown), and the EGR valve 57 is driven.
[0156]
This is the end of the description of EGR control.
[0157]
As described above, in the embodiment of the present invention, the above-described direct and simple above-described four elements, that is, the intake air amount (weighted average value) Qas0, the fuel injection amount Qf, the effective area equivalent value Avnt of the variable nozzle, and the exhaust gas temperature Texh. Since the exhaust pressure Pexh can be calculated using the equation (12), the exhaust pressure can be estimated without delay in response even when a variable capacity turbocharger is provided.
[0158]
Further, the effective area equivalent value Avnt is given by the product of the efficiency ηn of the gas flowing through the variable nozzle 2d and the command opening VNTstep given to the step motor 2c for driving the variable nozzle 2d, and therefore the gas flowing through the variable nozzle 2d. The efficiency ηn can be considered.
[0159]
Further, since the efficiency ηn of the gas flowing through the variable nozzle 2d is the product of the friction loss ξfric and the nozzle loss ξconv, the friction loss and the nozzle loss can be considered separately.
[0160]
Further, since the friction loss ξfric is given as a value proportional to the square root of the exhaust flow velocity equivalent value Wexh, the friction loss ξfric can be given accurately even if the exhaust flow velocity is different.
[0161]
Further, when the change in the flow velocity is large, the loss to the constricted pipe (1 / {1- (A₂/ A₁)²}^1/2In this embodiment, the nozzle loss ξconv is set to a value corresponding to the command opening degree VNTstep and the total exhaust weight Qtotal. Even when the change in flow velocity is large, the actual nozzle loss can be matched well.
[0162]
Further, the exhaust temperature nozzle opening correction coefficient Ktexh4 is calculated according to the command opening VNTstep and the exhaust amount Qexh, and the exhaust temperature basic value Texhb is corrected with this correction coefficient Ktexh4. And the calculation accuracy of the exhaust pressure Pexh is improved by this improvement. Similarly, when the intake port is provided with a swirl valve, an exhaust temperature swirl correction coefficient Ktexh3 is calculated according to the opening position of the swirl valve and the engine speed Ne, and the exhaust gas basic value is calculated using the correction coefficient Ktexh3. Since Texhb is corrected, the calculation accuracy of the exhaust temperature Texh is improved even when the intake port is provided with a swirl valve, and the calculation accuracy of the exhaust pressure Pexh is improved by this improvement.
[0163]
The calculation formula of the exhaust pressure (turbine inlet pressure) is not limited to the above formula 12. This will be described with reference to the calculation flow of FIG. 46. This flow replaces FIG.
[0164]
In FIG. 34, the exhaust pressure was calculated on the assumption that the flow of gas passing through the nozzle 2d is a flow when the flow path area is reduced (see FIG. 44), whereas in FIG. The flow of the passing gas is determined on the assumption that the flow of the ideal gas flows when the ideal gas is adiabatically changed (see FIG. 48). Specifically, in FIG.
[0165]
[Equation 19]
Pexhr = Kpexhn × {(Qas0 + Qfuel) / Avnt}²× Texh
Where Kpexhn: constant,
The turbine inlet exhaust pressure equivalent value Pexhr is calculated by the following equation, and the exhaust pressure Pexh0 is obtained by searching a map having the contents shown in FIG. 47 in step 12 from the Pexhr and the atmospheric pressure Pa. The rest is the same as FIG. 34, and a weighted average process is performed on this Pexh0 in step 3, and the weighted average value is obtained as the exhaust pressure Pexh.
[0166]
Here, how the exhaust pressure calculation method of Formula 19 is obtained will be described next.
[0167]
<5> Basic formula of flow in case of tapered nozzle
Considering the flow passing through the turbine nozzle, there is almost no external heat input / output and work, and therefore, the energy of the fluid is considered to change from the decrease in internal energy to kinetic energy and extrusion work. The engine exhaust can be regarded as an ideal gas because of its low pressure and high temperature. Therefore, the flow of exhaust gas passing through the turbine nozzle can be considered as “the ideal gas flows with adiabatic change”.
[0168]
Now, in a tapered nozzle such as a turbine nozzle, as shown in FIG. 48, the pressure, specific volume, flow velocity, area, temperature, specific heat ratio, and gas constant are P, v, w, A, T, κ, Let R be the subscript 1 for the entrance and subscript 2 for the exit.
[0169]
[Expression 20]

It is. Further, the following equation is established from the basic equation of steady flow energy.
[0170]
[Expression 21]
Substituting equation (31) into equation 19,
[0171]
[Expression 22]
Or P₁V₁= RT₁From
[0172]
[Expression 23]
For tapered nozzles, inlet flow velocity w₁Is the exit velocity w₂If it is omitted because it is very small compared to₂Is given by:
[0173]
[Expression 24]

Since the flow rate Q flowing through each section of the nozzle per unit time is constant from the continuous equation,
Q = A₂× w₂/ V₂= Ρe × A₂× w₂[Kg / sec] (33)
It is. In addition, since the fluid flowing in the nozzle is considered to change adiabatically with an ideal gas, from the above equation (31),
[0174]
[Expression 25]

It is.
[0175]
Substituting (32) and (34) into (33),
[0176]
[Equation 26]

Equation (35) is a basic equation for the flow when the nozzle is a tapered nozzle.
[0177]
<6> Examination of exhaust pressure prediction formula for tapered nozzle
Refer to FIG. From equation (15)
Qas0 + Qf = Qtotal [kg / sec] (36)
It is. Applying the above equation (35), which is a nozzle flow equation whose area is reduced, to the right side of the equation (36),
[0178]
[Expression 27]

To get the formula
[0179]
Here, the turbine inlet exhaust pressure equivalent value Pexhr is
[0180]
[Expression 28]

Then, the equation (37a) is

Therefore, when the equation (39) is arranged for the turbine inlet exhaust pressure equivalent value Pexhr, the following equation is obtained.
[0181]
Pexhr = (1 / (2 × ρe)) × (κe / (κe−1)) × {(Qas0 + Qf) / Avnt}²[Pa] (40)
Here again, the following final expression is obtained by substituting Expression (11), which is an approximate expression of the exhaust density ρe, into Expression (40).
[0182]
Pexhr = Kpexhn × {(Qas0 + Qf) / Avnt}²× Texh [Pa] (41a)
Kpechn = (1 / (2 × ρe)) × (κe / (κe−1)) (41b)
Where Kpexhn: constant.
[0183]
Now, as shown in the above equation 12, if the exhaust pressure is calculated on the assumption that the flow of gas passing through the nozzle is a flow when the flow path area is reduced, the standard state (298K, 0. 1MP), a high calculation accuracy of the exhaust pressure can be obtained. However, according to experiments, when the exhaust pressure is different from the standard state (for example, when the temperature is higher than the high altitude, the standard temperature, the humidity is different from the standard state, etc.) It has been found that the calculation accuracy of is reduced. This seems to be because Equation (12) takes into account the change in specific gravity, but is not yet accurate.
[0184]
On the other hand, according to the fourth embodiment obtained by assuming that the flow of gas passing through the nozzle is a flow when the ideal gas is adiabatically changed and flows, the flow rate per unit time is calculated by an arithmetic expression. Since the pressure and pressure (that is, the change in specific gravity) can be accurately described, high calculation accuracy of the exhaust pressure can be obtained even in a state of atmospheric pressure or temperature different from the standard state.
[0185]
Moreover, since the characteristics of FIG. 47 that must be matched are simple as shown in the figure, only calculation is sufficient (no matching is required), and the calculation accuracy of the exhaust pressure is high even with calculation only on the desk ( Confirmed by experiment).
[0186]
Now, the engine is equipped with an EGR device and a variable capacity turbocharger, and the EGR control using the model reference control described above and the above-described supercharging pressure control are performed, and the vehicle is also provided with a manual transmission for acceleration. At this time, in this vehicle, it was found that combustion noise and smoke increased temporarily immediately after starting acceleration, and that combustion noise increased temporarily during shifting during acceleration. In order to analyze this, an experiment was conducted in which the fuel injection amount was increased stepwise while the engine speed was kept constant, and the fuel injection amount was decreased stepwise after a certain period. FIG. 49 shows the experimental result at this time as a model. In the figure, the timing of t1 when the fuel injection amount Qf increases stepwise corresponds to the start acceleration, and the timing of t2 when the fuel injection amount Qf decreases stepwise corresponds to shifting during acceleration.
[0187]
Note that the timing of t2 corresponds to the time of shifting during acceleration for the following reason. In a vehicle equipped with a manual transmission, for example, when the gear position is set to the first speed, the accelerator pedal is depressed to accelerate (start acceleration), and then the gear position is switched to the second speed from the accelerator pedal that was previously depressed. Release to move to the clutch pedal and depress the clutch pedal to disengage the clutch. In this clutch disengagement process, the accelerator pedal opening decreases, and the fuel supply amount decreases corresponding to the decrease in the accelerator pedal opening. However, the present invention is not limited to a vehicle equipped with a manual transmission, and the timing of t2 corresponds to the middle of a speed change during acceleration even in a vehicle equipped with an automatic transmission. This is because, in a vehicle equipped with an automatic transmission, a command for reducing the fuel injection amount is issued in order to relieve a shift shock during a shift.
[0188]
Hereinafter, the analysis is divided into (1) immediately after starting acceleration and (2) shifting during acceleration.
(1) Immediately after starting acceleration:
The exhaust pressure Pexh rises with good response due to an increase in load, whereas the intake pressure Pm rises later due to the operation delay of the turbocharger 2, so that the differential pressure ΔP (= Pexh−Pm) between the intake pressure Pm and the exhaust pressure Pexh. ) Increases as shown in the fifth stage, and as a result, the actual EGR flow rate (Qe calculated by Equation 13) temporarily increases as shown in the third stage dashed line, and then the fuel after the increase It converges with a first-order lag to the EGR flow rate corresponding to the injection amount Qf. Low-temperature premixed combustion cannot be maintained due to a temporary increase in the EGR flow rate associated with the operation delay of the turbocharger 2. The combustion state shifts to normal diesel combustion mainly consisting of diffusion combustion, and combustion noise (CPL of 1 kHz band) and smoke (ISF) increase.
(2) During shifting during acceleration:
The exhaust pressure Pexh decreases with good response due to the decrease in the load, whereas the intake pressure Pm decreases with a delay in the operation of the turbocharger 2, so that the differential pressure ΔP decreases as in the fifth stage. The flow rate temporarily decreases, and thereafter converges with a first-order lag to the EGR flow rate corresponding to the decreased fuel injection amount Qf. Even when the EGR flow rate temporarily decreases due to the operation delay of the turbocharger 2, the combustion state shifts from low-temperature premixed combustion to normal diesel combustion mainly consisting of diffusion combustion, and combustion noise increases.
[0189]
Therefore, in the control unit 41, in the case of (1), that is, when the actual supercharging pressure (= intake pressure) is smaller than the target supercharging pressure and the actual EGR flow rate is larger than the target EGR flow rate, the common rail pressure (fuel injection pressure) Is increased and the injection timing is retarded. Combustion after promoting the mixing of fuel and air by increasing the common rail pressure to promote atomization of fuel spray to reduce smoke and retard the injection timing to increase the ignition delay period This maintains low-temperature premixed combustion and suppresses both combustion noise and smoke.
[0190]
In the case of (2), that is, when the actual supercharging pressure is larger than the target supercharging pressure and the actual EGR flow rate is smaller than the target EGR flow rate, the common rail pressure is corrected to decrease and the injection timing is corrected to retard. By reducing the common rail pressure, fuel atomization is slowed down and combustion is slowed down (smoke does not deteriorate because the amount of oxygen is sufficient), and the fuel injection timing is retarded to increase the ignition delay period. Thus, by promoting the mixing of the fuel and air, the low temperature premixed combustion is maintained also in this case, thereby suppressing the combustion noise.
[0191]
This is summarized in the table of FIG. 50. In FIG. 50, the upper right column corresponds to the case (1) above, and the lower left column corresponds to the case (2) above.
[0192]
In this case, according to the combination of the comparison result of the actual supercharging pressure and the target supercharging pressure and the comparison result of the actual EGR flow rate and the target EGR flow rate, the following cases (3) and (4) can be considered. For the cases of 3) and (4), the common rail pressure and the injection timing are controlled as follows.
(3) When the actual boost pressure is greater than the target boost pressure and the actual EGR flow rate is greater than the target EGR flow rate:
The reduction correction amount of the common rail pressure and the retardation correction amount of the injection timing are made smaller than in the case (2) (corresponding to the lower right column in FIG. 50).
(4) When the actual boost pressure is smaller than the target boost pressure and the actual EGR flow rate is smaller than the target EGR flow rate:
The increase correction amount of the common rail pressure and the retardation correction amount of the injection timing are made smaller than in the case (1) (corresponding to the upper left column in FIG. 50).
[0193]
Here, the cases (3) and (4) are for dealing with the case where the actual EGR flow rate is a response that overshoots. For example, if the change in EGR flow rate during shifting during acceleration is modeled as shown in FIG. 51, in the case of a normal response of the actual EGR flow rate, the control of the common rail pressure and the injection timing in the case of (2) above is used. did. However, in the case of a response that overshoots the actual EGR flow rate exceeding the target value (response with excessive gain in the figure), even in the section from t3 to t4 when the actual EGR flow rate exceeds the target EGR flow rate, When the control is performed, the common rail pressure decrease correction amount and the fuel injection timing retardation correction amount are too large. Therefore, in order to reduce each correction amount, the control in the case (3) is performed.
[0194]
The above control performed by the control unit 41 will be described in detail below.
[0195]
FIG. 52 shows common rail pressure correction amount K It is executed every 10 msec in the main calculation flow.
[0196]
In step 1, the engine speed Ne, the fuel injection amount Qf, the cylinder intake EGR amount Qec (calculated by FIG. 13), the target cylinder intake EGR amount Tqec (calculated in step 4 of FIG. 41), the actual boost pressure Pm, the target boost pressure Read TPm.
[0197]
Here, the cylinder intake EGR amount Qec is used as the actual EGR flow rate, and the target cylinder intake EGR amount Tqec is used as the target value. However, the actual EGR flow rate Qe (calculated from FIG. 36) and the corresponding target value EGR as the target value. The flow rate Tqe (see FIG. 41) may be used. Since the actual supercharging pressure Pm_ist (detected by the sensor 72) is equal to the intake pressure Pm, Pm is used as a sign. Further, the target boost pressure Pm_sol is calculated in step 4 of FIG. 4, and a symbol TPm is used in correspondence with a symbol Pm.
[0198]
In step 2, the difference dQec between the cylinder intake EGR amount Qec and the target cylinder intake EGR amount Tqec and the difference dPm between the actual boost pressure Pm and the target boost pressure TPm are calculated, and the difference dQec, dPm is calculated in step 3 from FIG. The basic value KQB of the common rail pressure correction amount is searched by searching one of the two tables containing 53. P is calculated.
[0199]
Further, by searching a map having the contents shown in FIG. 54 in step 4 from the engine speed Ne and the fuel injection amount Qf, an operating condition reflection coefficient KQC as a correction gain is obtained. P is calculated and this coefficient KQC P and the above basic value KQB In step 5 using P
[0200]
[Expression 18]
K Pail = KQB P × KQC P
The common rail pressure correction amount K Calculate Prail.
[0201]
Where the basic value KQB P gives the common rail pressure correction amount described above with reference to FIG. That is, KQB The value of P increases in proportion to dQec when TPm> Pm and Qec> Tqec (in the case of (1) above) (see the right side of FIG. 53), and when TPm <Pm and Qec <Tqec (above) (In the case of (2)), it is a value that becomes negative and large in proportion to dQec (see the left figure of FIG. 53). In the case of TPm> Pm and Qec <Tqec (in the case of (3) above), the slope of the straight line is made gentler (see the right diagram in FIG. 53) than in the case of (2) above, and similarly TPm <Pm and Qec> In the case of Tqec (in the case of (4) above), the slope of the straight line is made gentler than in the case of (1) (see the left figure in FIG. 53).
[0202]
In FIG. 54, A is a low temperature premixed combustion region, C is a normal diesel combustion region mainly composed of diffusion combustion, B is an intermediate region between them, and combustion noise is particularly remarkable in the region B. KQC in the area of B P is set to a maximum of 1.0.
[0203]
KQC in area A The reason why P = 0.8 is that in this region, the response is slow for both the EGR flow rate and the supercharging pressure. If the correction gain is large, there is a possibility of hunting. Therefore, the correction gain is reduced to prevent this hunting. Is.
[0204]
In the area C, KQC The reason why P = 0.6 is as follows. Since this region is a region where EGR is stopped and the high boost pressure is used, in principle, this region does not need to be corrected. However, KQC Since the value of P is ultimately determined by adaptation, the value at C may be required. Therefore, a value can be entered even in the region C (0.6 is a provisional value).
[0205]
FIG. 55 shows the injection timing correction amount K. It is a calculation flow of IT.
[0206]
Steps 1 and 2 are the same as Steps 1 and 2 in FIG. 52 (the difference dQec between the cylinder intake EGR amount Qec and the target cylinder intake EGR amount Tqec, and the difference dPm between the actual boost pressure Pm and the target boost pressure TPm are set. Calculate).
[0207]
In step 3, the basic value KQB of the injection timing correction amount is obtained by searching one of two tables having the contents shown in FIG. 56 from these differences dQec and dPm. By retrieving a map containing the contents of FIG. 57 in step 4 from IT and the engine speed Ne and the fuel injection amount Qf, an operating condition reflection coefficient KQC as a correction gain is obtained. IT is calculated and these operating condition reflection coefficients KQC IT, basic value KQB In step 5 using IT
[0208]
[Equation 19]
K IT = KQB IT x KQC IT
The injection timing correction amount K_IT is calculated by the following formula.
[0209]
Where the basic value KQB IT gives the retard correction amount of the fuel injection timing described above with reference to FIG. That is, the basic value KQB IT is KQB in proportion to dQec when TPm> Pm and Qec> Tqec (case (1) above). When the value of IT is negative and large (see the right diagram in FIG. 56), and TPm <Pm and Qec <Tqec (in the case of (2) above), KQB is proportional to dQec. The value of IT is negative and large (see the left figure in FIG. 56). Further, in the case of TPm> Pm and Qec <Tqec (in the case of (3) above), the slope of the straight line is made gentler than that in the case of (2) (see the right side of FIG. 56), and similarly TPm <Pm and Qec> In the case of Tqec (in the case of (4) above), the slope of the straight line is made gentler than in the case of (1) (see the left figure in FIG. 56).
[0210]
Basic value KQB IT is given as a negative value, as will be described later in step 7 of FIG. A value obtained by adding IT to the target injection timing TIT0 is obtained as the target injection timing TIT1, and the target injection timing in this case is a value (advance amount) measured from the predetermined crank angle position to the advance side, so Correction amount K for angle correction This is because IT must be given as a negative value.
[0211]
57, KQC in the area A P = 0.8, and KQC in C region P = 0.6 is for the same reason as in FIG.
[0212]
FIG. 58 is a calculation flow of the target common rail pressure TPrail.
[0213]
Here, in step 7, the common rail pressure correction amount K described above is used. The rest is the same as the prior art in that the common rail pressure is corrected by using a rail, and the rest is the same as the prior art.
[0214]
Specifically, in step 1, the engine speed Ne, the fuel injection amount Qf, the atmospheric pressure (compressor inlet pressure) Pa, the cooling water temperature Tw, the intake fresh air temperature Ta (calculated from FIG. 18), the actual common rail pressure Prail (detected by the sensor 32). ) Plus common rail pressure correction amount K Prail is read, and the basic value TPrailB of the target common rail pressure, the cooling water temperature Tw, and the intake fresh air temperature Ta are obtained by searching a map having the contents shown in FIG. 59 in step 2 from the engine speed Ne and the fuel injection amount Qf. 60, 61 and 62 are searched from the atmospheric pressure Pa in steps 3, 4, and 5 to search the water temperature correction coefficient KPTw, the intake air temperature correction coefficient KPTa, and the atmospheric pressure correction coefficient for the target common rail pressure. Determine KPPa and in step 6
[0215]
[Expression 20]
TPrail0 = TPrailB × KPTw × KPTa × KPPa
The target common rail pressure TPrail0 is calculated by the following formula.
[0216]
Here, as shown in FIG. 60, the value of the correction coefficient KPTw is set to a value larger than 1.0 at the time of low water temperature. This is prevented because the fuel temperature is low and the state of fuel spray becomes worse at low water temperature. It is to do. As shown in FIG. 61, when the intake fresh air temperature Ta is low, the correction coefficient KPTa is set to a value larger than 1.0 because the fuel spray is less likely to vaporize when the intake fresh air temperature Ta is low. This is to reduce the size. As shown in FIG. 62, when the atmospheric pressure Pa is low, the correction coefficient KPPa is set to a value larger than 1.0. If the atmospheric pressure Pa is low, the actual compression ratio becomes low and ignition becomes difficult. This is to make it easier to ignite.
[0217]
In Step 7, the common rail pressure correction amount K With Prail
[0218]
[Expression 21]
TPrail1 = TPrail0 × K Prail
The target common rail pressure TPrail0 is corrected by the following equation, and the corrected target common rail pressure TPrail1 is set as the target common rail pressure TPrail1.
[0219]
In step 8, the PI correction amount is calculated by PI control so that the actual common rail pressure Trail matches the target common rail pressure TPrail1, and in step 9, this PI correction amount is added to the target common rail pressure TPrail1, and the target common rail pressure after the addition is calculated. Is set to the target common rail pressure TPrail2.
[0220]
In step 10, the map of FIGS. 63 and 64 is searched from the engine speed Ne and the fuel injection amount Qf to obtain the maximum common rail pressure PrailMAX and the minimum common rail pressure PrailMIN, and if TPrail2 is between the maximum value and the minimum value. The value of TPrail2 is calculated as the target common rail pressure TPrail, and when TPrail2 exceeds the maximum common rail pressure PrailMAX, the maximum common rail pressure PrailMAX is calculated.
[0221]
FIG. 65 is a calculation flow of the target injection timing TIT. Here, in step 7, the injection timing correction amount K The rest is the same as in the prior art in that the target injection timing is corrected using IT and is different from the prior art.
[0222]
Specifically, in step 1, the engine speed Ne, fuel injection amount Qf, atmospheric pressure Pa, cooling water temperature Tw, intake fresh air temperature Ta, injection timing correction amount K IT is read, and the basic value TITB of the target injection timing, the cooling water temperature Tw, the intake fresh air temperature Ta are obtained by searching a map having the contents shown in FIG. 66 in step 2 from the engine speed Ne and the fuel injection amount Qf. 67, 68, and 69 are searched from the atmospheric pressure Pa in steps 3, 4, and 5 to search the water temperature correction coefficient KITTw, the intake air temperature correction coefficient KITTa, and the atmospheric pressure correction coefficient for the target injection timing. Find KITPa and in step 6
[0223]
[Expression 22]
TIT0 = TITB × KITTw × KITTa × KITa
The target injection timing TIT0 is calculated by the following formula.
[0224]
Here, the basic value TITB is set so that fuel injection is started within a predetermined range after compression top dead center at the crank angle.
[0225]
The basic value TITB is a value (advance amount) measured from a predetermined crank angle position to the advance side. Therefore, the injection timing is advanced when the correction coefficients KITTw, KITTa, and KITPa are larger than 1.0. As shown in FIG. 67, the reason why the value of the correction coefficient KITTw is set to a value larger than 1.0 at the time of low water temperature is that the fuel temperature is low and combustion tends to be delayed at the time of low water temperature. It is for bringing. 68, the correction coefficient KITTa is set to a value larger than 1.0 when the intake fresh air temperature Ta is low, and the correction coefficient KITPa is set to a value larger than 1.0 when the atmospheric pressure Pa is low as shown in FIG. This is because of the same reason.
[0226]
In step 7, the injection timing correction amount K Using IT
[0227]
[Expression 23]
TIT1 = TIT0 + K IT
The target injection timing TIT0 is corrected by the following formula, and the corrected value is set as the target injection timing TIT1. Injection timing correction amount K Since IT is a negative value, the target injection timing is retarded by Equation 23.
[0228]
In step 8, the maps of FIGS. 70 and 71 are searched from the engine speed Ne and the fuel injection amount Qf to obtain the maximum injection timing ITMAX and the minimum injection timing ITMIN, and the target injection timing TIT1 is between the maximum value and the minimum value. If the target injection timing TIT1 exceeds the maximum injection timing ITMAX, the maximum injection timing ITMAX is set. If the target injection timing TIT1 is lower than the minimum injection timing ITMIN, the minimum injection timing ITMIN is set to the target fuel injection timing TIT. Calculate as
[0229]
Here, the operation of the present embodiment will be described with reference to FIG.
[0230]
According to this embodiment. When the actual supercharging pressure Pm is smaller than the target supercharging pressure TPm and the actual EGR flow rate is larger than the target EGR flow rate from the timing t1, the target common rail pressure TPrail is corrected to increase (see the broken line in the sixth stage in FIG. 49). ), Atomization of fuel spray is promoted, smoke is reduced, and the target injection timing TIT is retarded and the ignition delay period is increased as shown by the broken line in the seventh row of FIG. After mixing is promoted, it burns. In other words, the target common rail pressure and target injection timing are corrected so that low-temperature premixed combustion is realized, so that low-temperature premixed combustion is maintained even during start acceleration, thereby suppressing both combustion noise and smoke. (Refer to the dashed lines in the first and second stages from the bottom of FIG. 49).
[0231]
Further, according to the present embodiment, when the actual supercharging pressure Pm is larger than the target supercharging pressure TPm and the actual EGR flow rate becomes smaller than the target EGR flow rate from the timing t2, the target common rail pressure TPrail is corrected to decrease ( 49 (see the sixth stage broken line), the atomization of the fuel spray is slowed down and the combustion is slow (the smoke is not deteriorated because the oxygen amount is sufficient), and the target injection timing TIT as shown in the seventh stage broken line in FIG. Is retarded and the ignition delay period is increased, so that the mixture of fuel and air is promoted to burn. Even in this case, correction of the target common rail pressure and target injection timing is performed so that low-temperature premixed combustion is realized, and low-temperature premixed combustion is maintained even during shifting during start acceleration, thereby suppressing combustion noise. (See the dashed line in the second row from the bottom in FIG. 49).
[0232]
Although not shown, when the response is such that the actual EGR flow rate overshoots after a temporary increase immediately after the timing of t1, the control in the case of the above (1) is performed even in the section where the actual EGR flow rate is lower than the target EGR flow rate. Therefore, the increase correction amount of the target common rail pressure and the delay angle correction amount of the target injection timing are too large, or the response is such that the actual EGR flow rate overshoots after a temporary decrease immediately after the timing t2 (FIG. 51), when the control in the case (2) is performed even in the section where the actual EGR flow rate exceeds the target EGR flow rate, the target common rail pressure reduction correction amount and the target injection timing retardation correction amount are too large. However, in this case, according to the present embodiment, each correction amount is made smaller than in the cases (1) and (2), so that each correction amount is large. This makes it possible to give an appropriate correction amount even in the case of a response in which the actual EGR flow rate overshoots after a temporary increase or temporary decrease during start acceleration or during a shift during start acceleration. .
[0233]
On the other hand, a common rail fuel injection device and an EGR device are provided, an actual EGR rate is obtained from the oxygen concentration in the exhaust detected by the wide-range air-fuel ratio sensor, and at the time of acceleration when the actual EGR rate becomes higher than the target EGR rate. There is a conventional device that reduces the smoke by correcting the common rail pressure to increase, and conversely, when the actual EGR rate is lower than the target EGR rate, there is a conventional device that reduces the common rail pressure to reduce the amount of NOx generated. According to this, since a detection delay occurs in the wide-range air-fuel ratio sensor, correction of the fuel injection pressure is delayed at the initial stage of acceleration when the target EGR rate and the actual EGR rate are most different from each other, and this causes the smoke to be worsened or deteriorated. May occur.
[0234]
In this case, the comparison between the actual EGR rate and the target EGR rate may be a comparison between the actual EGR flow rate and the target EGR flow rate. Therefore, by calculating the actual EGR flow rate by model reference control, the response is good even during transient operation. Combining a device that obtains an actual EGR flow rate (see Japanese Patent Laid-Open No. 10-318047) and the above-described conventional device provides a correction amount of the fuel injection pressure in accordance with the response delay timing of the actual EGR flow rate. Can do.
[0235]
However, when a variable capacity turbocharger having a variable nozzle is provided in the turbine in order to avoid a decrease in the excess air ratio due to a large amount of EGR, a response delay of the actual boost pressure is caused by a so-called turbo lag during acceleration or shifting. Because this affects the actual EGR flow rate, in order to avoid a temporary increase in the actual EGR flow rate during acceleration, for example, by reducing the nozzle opening of the variable nozzle, the intake air flow rate is reduced and the excess air ratio is improved. However, this method may increase the temporary increase in the actual EGR flow rate during acceleration.
[0236]
On the other hand, in this embodiment, for an engine including an EGR device capable of mass EGR and a variable displacement turbocharger having a variable nozzle in the turbine, the actual EGR flow rate is controlled even during transient operation by model reference control. While calculating with good response, when the actual EGR flow rate obtained in this way is larger than the target EGR flow rate and the actual supercharging pressure is smaller than the target supercharging pressure, the target injection pressure is increased and corrected. When the actual EGR flow rate is smaller than the target EGR flow rate and the actual boost pressure is greater than the target boost pressure, the target injection pressure is corrected to decrease, so the engine in the practical operating range (low speed or low load) While improving both the exhaust composition and operability of the engine, there is a response delay of the actual EGR flow rate due to the operation delay of the EGR valve during acceleration or during shifting during acceleration. Further even when the turbo lag affects, it is possible to correct the fuel injection pressure to match the response delay of the phase of the actual EGR flow rate, it is possible to improve the correction accuracy of the injection pressure.
[0237]
In the embodiment, the case where both the target common rail pressure and the target injection timing are corrected has been described. However, only the target common rail pressure may be corrected.
[0238]
The embodiment has been described with an engine that includes a common rail fuel injection device (fuel injection pressure control device), an EGR device, and a variable displacement turbocharger, and performs low-temperature premixed combustion in a predetermined operation range (for example, a low and medium load range). However, the present invention is not limited to this, and can also be applied to an engine that performs normal diesel combustion mainly using diffusion combustion in the entire operation region.
[0239]
In the embodiment, the case where the sensor 73 for detecting the compressor inlet pressure Pa is provided has been described. However, a vehicle equipped with an engine including the EGR device and the variable capacity turbocharger is in a standard atmosphere (or an atmosphere close thereto) The compressor inlet pressure sensor is not required as long as it is operated at This is because the Pa value for the standard atmosphere can be set at this time.
[Brief description of the drawings]
FIG. 1 is a control system diagram of a first embodiment.
FIG. 2 is a system diagram of a common rail fuel injection device.
FIG. 3 is an EGR control system diagram.
FIG. 4 is a flowchart for explaining calculation of a command opening degree given to a variable nozzle actuator.
FIG. 5 is a characteristic diagram of basic supercharging pressure.
FIG. 6 is a characteristic diagram of an atmospheric pressure correction value.
FIG. 7 is a characteristic diagram of the basic opening.
FIG. 8 is a characteristic diagram of an atmospheric pressure correction value.
FIG. 9 is a block diagram of an EGR control system.
FIG. 10 is a flowchart showing the calculation order of parameters in model reference control.
FIG. 11 is a flowchart for explaining cycle processing;
FIG. 12 is a flowchart for explaining calculation of a cylinder intake fresh air amount.
FIG. 13 is a flowchart for explaining calculation of a cylinder suction EGR amount.
FIG. 14 is a flowchart for explaining calculation of a volume efficiency equivalent value.
FIG. 15 is a characteristic diagram of air density.
FIG. 16 is a flowchart for explaining calculation of intake pressure.
FIG. 17 is a characteristic diagram of pressure with respect to sensor output voltage.
FIG. 18 is a flowchart for explaining calculation of intake air temperature.
FIG. 19 is a characteristic diagram of a vehicle speed correction value of intake air temperature.
FIG. 20 is a characteristic diagram of an intake air amount correction value of intake air temperature.
FIG. 21 is a flowchart for explaining calculation of cylinder intake gas temperature;
FIG. 22 is a flowchart for explaining calculation of a fuel injection amount.
FIG. 23 is a characteristic diagram of a basic fuel injection amount.
FIG. 24 is a characteristic diagram of a maximum injection amount.
FIG. 25 is a flowchart for explaining calculation of exhaust gas temperature.
FIG. 26 is a characteristic diagram of an exhaust gas temperature basic value.
FIG. 27 is a characteristic diagram of an intake air temperature correction coefficient.
FIG. 28 is a characteristic diagram of an exhaust pressure correction coefficient.
FIG. 29 is a characteristic diagram of a swirl correction coefficient.
FIG. 30 is a characteristic diagram of a nozzle opening correction coefficient.
FIG. 31 is a flowchart for explaining calculation of a nozzle effective area equivalent value;
FIG. 32 is a characteristic diagram of friction loss.
FIG. 33 is a characteristic diagram of nozzle loss.
FIG. 34 is a flowchart for explaining calculation of exhaust pressure.
FIG. 35 is a characteristic diagram in which the correlation between the actual measured value and the predicted value of the exhaust pressure is examined.
FIG. 36 is a flowchart for explaining calculation of an EGR flow rate.
FIG. 37 is a characteristic diagram of an EGR valve opening area equivalent value.
FIG. 38 is a flowchart for explaining calculation of a target EGR rate.
FIG. 39 is a characteristic diagram of a target EGR rate basic value.
FIG. 40 is a characteristic diagram of a target EGR rate correction value.
FIG. 41 is a flowchart for explaining calculation of a required EGR amount.
FIG. 42 is a flowchart for explaining calculation of a command EGR valve lift amount.
FIG. 43 is a characteristic diagram of an EGR valve target lift amount.
FIG. 44 is a model diagram of a flow in which the flow path area is reduced.
FIG. 45 is a model diagram used for examining the mechanical balance of the intake and exhaust systems.
FIG. 46 is a flowchart for explaining another exhaust gas pressure calculation.
FIG. 47 is a characteristic diagram of the exhaust pressure Pexh0.
FIG. 48 is a model diagram of a tapered nozzle.
FIG. 49 is a waveform diagram for explaining the operation of the present embodiment.
FIG. 50 is a table showing the control content of the embodiment.
FIG. 51 is a model diagram showing a change in actual EGR flow rate during shifting during acceleration.
FIG. 52 is a flowchart for explaining calculation of a common rail pressure correction amount.
FIG. 53 is a characteristic diagram of a basic value of a common rail pressure correction amount.
FIG. 54 is a characteristic diagram of an operating condition reflection coefficient.
FIG. 55 is a flowchart for explaining calculation of an injection timing correction amount.
FIG. 56 is a characteristic diagram of the basic value of the injection timing correction amount.
FIG. 57 is a characteristic diagram of an operating condition reflection coefficient.
FIG. 58 is a flowchart for explaining calculation of a target common rail pressure.
FIG. 59 is a characteristic diagram of a target common rail pressure basic value.
FIG. 60 is a characteristic diagram of a water temperature correction coefficient.
FIG. 61 is a characteristic diagram of an intake air temperature correction coefficient.
FIG. 62 is a characteristic diagram of an atmospheric pressure correction coefficient.
FIG. 63 is a characteristic diagram of maximum common rail pressure.
FIG. 64 is a characteristic diagram of minimum common rail pressure.
FIG. 65 is a flowchart for explaining calculation of a target injection timing.
FIG. 66 is a characteristic diagram of a target injection timing basic value.
FIG. 67 is a characteristic diagram of a water temperature correction coefficient.
FIG. 68 is a characteristic diagram of an intake air temperature correction coefficient.
FIG. 69 is a characteristic diagram of an atmospheric pressure correction coefficient.
FIG. 70 is a characteristic diagram of maximum injection timing.
FIG. 71 is a characteristic diagram of minimum injection timing.
FIG. 72 is a waveform diagram when the nozzle opening is reduced during acceleration.
FIG. 73 is a claim correspondence diagram of the first invention.
[Explanation of symbols]
2 Variable capacity turbocharger
2d variable nozzle
10 Common rail fuel injection system
17 Fuel injection valve
41 Control unit
57 EGR valve

Claims (6)

An EGR valve;
A variable capacity turbocharger having a variable nozzle in the turbine;
Means for calculating a target injection pressure according to the engine load;
Means for calculating a target EGR flow rate according to the engine load;
Means for controlling the EGR valve so that the target EGR flow rate flows;
Means for calculating the actual EGR flow rate by model reference control;
Means for calculating a target boost pressure according to the engine load;
Means for controlling the variable nozzle opening so as to obtain the target supercharging pressure;
Means for detecting the actual supercharging pressure;
Means for comparing the actual supercharging pressure and the target supercharging pressure, and comparing the actual EGR flow rate and the target EGR flow rate;
A means for increasing and correcting the target injection pressure when the actual supercharging pressure is lower than the target supercharging pressure and the actual EGR flow rate is larger than the target EGR flow rate based on these comparison results;
Means for controlling the fuel injection pressure so as to be the corrected target injection pressure ,
Based on the comparison result, when the actual boost pressure is lower than the target boost pressure and the actual EGR flow rate is greater than the target EGR flow rate, the target injection timing is retarded.
A control device for a diesel engine. An EGR valve;
A variable capacity turbocharger having a variable nozzle in the turbine;
Means for calculating a target injection pressure according to the engine load;
Means for calculating a target EGR flow rate according to the engine load;
Means for controlling the EGR valve so that the target EGR flow rate flows;
Means for calculating the actual EGR flow rate by model reference control;
Means for calculating a target boost pressure according to the engine load;
Means for controlling the variable nozzle opening so as to obtain the target supercharging pressure;
Means for detecting the actual supercharging pressure;
Means for comparing the actual supercharging pressure and the target supercharging pressure, and comparing the actual EGR flow rate and the target EGR flow rate;
Means the increase correcting the target injection pressure when the actual boost pressure from these comparison results is lower than the target supercharging pressure and the actual EGR rate is larger than the target EGR rate,
Means for controlling the fuel injection pressure so as to be the corrected target injection pressure ,
Based on the comparison result, when the actual boost pressure is lower than the target boost pressure and the actual EGR flow rate is smaller than the target EGR flow rate, the target injection pressure is increased and corrected by a correction amount smaller than the target injection pressure increase correction amount.
A control device for a diesel engine. 3. The diesel engine control device according to claim 1, wherein the actual EGR flow rate is calculated based on a differential pressure between the exhaust pressure and the intake pressure. An intake air amount Qas0, a fuel injection amount Qf, an effective area equivalent value Avnt of the variable nozzle and an exhaust temperature Texh are detected, and the exhaust pressure Pexh is determined using these four elements.
Pexh = Kpexh × {(Qas0 + Qf) / Avnt)} ² × Texh + Pa
Where Pexh: exhaust pressure,
Qas0: intake air volume,
Qf: fuel injection amount,
Avnt: Effective area equivalent value of variable nozzle,
Texh: exhaust temperature at the turbine inlet,
Pa: compressor inlet pressure,
Kpexh: constant,
The diesel engine control device according to claim 3 , wherein the calculation is performed according to: An intake air amount Qas0, a fuel injection amount Qf, an effective area equivalent value Avnt of the variable nozzle, and an exhaust temperature Texh are detected, and using these four elements, a turbine inlet exhaust pressure equivalent value Pexhr is obtained.
Pexhr = Kpexhn × {(Qas0 + Qf) / Avnt)} ² × Texh
Where Pexhr: turbine inlet exhaust pressure equivalent value,
Qas0: intake air volume,
Qf: fuel injection amount,
Avnt: Effective area equivalent value of variable nozzle,
Texh: exhaust temperature at the turbine inlet,
Pa: compressor inlet pressure,
Kpexhn: constant,
4. The diesel engine control device according to claim 3 , wherein the exhaust pressure Pexh is calculated from the turbine inlet exhaust pressure equivalent value Pexhr and the compressor inlet pressure Pa. 5. 2. The diesel engine control device according to claim 1 , wherein the target injection pressure and the target injection timing are corrected so as to realize low-temperature premixed combustion.

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