JP4066866B2 - Ignition timing control device for internal combustion engine - Google Patents

Description

ただし、Ｖｃ：隙間容積［ｍ３］、
ε ：圧縮比、
Ｄ ：シリンダボア径［ｍ］、
ＳＴ ：ピストンの全ストローク［ｍ］、
Ｈ ：ピストンピン７６のＴＤＣからの距離［ｍ］、
ＣＮＤ ：コネクティングロッド７４の長さ［ｍ］、
ＣＲoff ：結節点７５のシリンダ中心軸７３からのオフセット距離［ｍ］、
ＰＩＳoff：クランクシャフト回転中心７２のシリンダ中心 軸７３からのオフセット距離［ｍ］、
θivc ：吸気弁閉時期のクランク角［ｄｅｇＡＴＤＣ］、
θoff ：ピストンピン７６とクランクシャフト回転中心７２とを結ぶ線がＴＤＣにおいて垂直線となす角 度［ｄｅｇ］、
Ｘ ：結節点７５とピストンピン７６との水平距離［ ｍ ］ 、
吸気弁閉時期のクランク角θivcは前述のように、エンジンコントローラ３１から吸気ＶＴＣ機構２７への指令信号によって決まるので、既知である。式（２）〜（６）にこのときのクランク角θivc（＝ＩＶＣ）を代入すれば、燃焼室５の吸気弁閉時期における容積ＶＩＶＣを算出することができる。したがって、実用上は燃焼室５の吸気弁閉時期における容積ＶＩＶＣは吸気弁閉時期ＩＶＣをパラメータとするテーブルで設定したものを用いる。吸気ＶＴＣ機構２７を備えないときには定数で与えることができる。
【００３９】
ステップ１３では、燃焼室５の吸気弁閉時期ＩＶＣにおける温度（つまり圧縮開始時期温度）ＴＩＮＩ［Ｋ］を算出する。燃焼室５の吸気弁閉時期ＩＶＣにおける温度ＴＩＮＩは、燃焼室５に流入する新気と燃焼室５に残留する不活性ガスとが混じったガスの温度であり、燃焼室５に流入する新気の温度は吸気コレクタ２内の新気温度ＴＣＯＬに等しく、また燃焼室５内に残留する不活性ガスの温度は排気ポート部近傍の排気温度ＴＥＸＨで近似できるので、吸気コレクタ２内の新気温度ＴＣＯＬ、排気温度ＴＥＸＨ、燃焼室５内に残留する不活性ガスの割合である内部不活性ガス率ＭＲＥＳＦＲから次式により求めることができる。
【００４０】

ステップ１４では、燃焼室５内の混合気の燃えやすさを表す反応確率ＲＰＲＯＢＡ［％］を算出する。反応確率ＲＰＲＯＢＡは無次元の値であり、残留不活性ガス率ＭＲＥＳＦＲ、冷却水温ＴＷＫ［Ｋ］、目標当量比ＴＦＢＹＡの３つのパラメータに依存するので、次式により表すことができる。
【００４１】
ＲＰＲＯＢＡ＝ｆ３（ＭＲＥＳＦＲ、ＴＷＫ、ＴＦＢＹＡ）…（８）
具体的に説明すると、ＭＲＥＳＦＲ、ＴＷＫ、ＴＦＢＹＡの３つのパラメータの組み合わせによって得られる反応確率の最大値を１００％とし、これらのパラメータと反応確率ＲＰＲＯＢＡの関係を実験的に求め、求めた反応確率ＲＰＲＯＢＡをパラメータに応じたテーブルとしてエンジンコントローラ３１のメモリに予め格納しておく。ステップ１４ではパラメータに応じてこのテーブルを検索することにより反応確率ＲＰＲＯＢＡを求める。
【００４２】
具体的には、冷却水温ＴＷＫに応じて図７に示すような特性を有する水温補正係数のテーブルと、同様に設定された内部不活性ガス率補正係数のテーブル（図示しない）と、目標当量比Ｔｆｂｙａに応じて図８に示すような特性を有する当量比補正係数のテーブルを予めメモリに格納しておく。各補正係数の最大値はそれぞれ１．０であり、３種類の補正係数の積に反応確率の最大値１００％を掛け合わせることで、反応確率ＲＰＲＯＢＡを算出する。
【００４３】
各テーブルを説明すると、図７に示す水温補正係数は冷却水温ＴＷＫが高いほど大きく、冷却水温ＴＷＫが８０℃以上では１．０になる。図８に示す当量比補正係数は目標当量比ＴＦＢＹＡが１．０のとき、つまり理論空燃比のときに最大値の１．０となり、目標当量比が１．０より大きくても小さくても当量比補正係数は減少する。内部不活性ガス率補正係数は図示しないが、内部不活性ガス率ＭＲＥＳＦＲがゼロの場合に１．０となる。
【００４４】
ステップ１５では、基準クランク角θＰＭＡＸ［ｄｅｇＡＴＤＣ］を算出する。前述のように基準クランク角θＰＭＡＸはあまり変動しないが、それでもエンジン回転速度ＮＲＰＭの上昇に応じて進角する傾向があるため、基準クランク角θＰＭＡＸはエンジン回転速度ＮＲＰＭの関数として次式で表すことができる。
【００４５】
θＰＭＡＸ＝ｆ４（ＮＲＰＭ）…（９）
具体的にはエンジン回転速度ＮＲＰＭから、エンジンコントローラ３１のメモリに予め格納された図９に示す特性のテーブルを検索することにより基準クランク角θＰＭＡＸを求める。算出を容易にするために、基準クランク角θＰＭＡＸを一定とみなすことも可能である。
【００４６】
最後にステップ１６では、点火無駄時間相当クランク角ＩＧＮＤＥＡＤ［ｄｅｇ］を算出する。点火無駄時間相当クランク角ＩＧＮＤＥＡＤは、エンジンコントローラ３１から点火指令信号、すなわち点火コイル１３の一次電流を遮断する信号を出力したタイミングから点火プラグ１４が実際に点火するまでのクランク角区間で、次式により表すことができる。
【００４７】
ＩＧＮＤＥＡＤ＝ｆ５（ＤＥＡＤＴＩＭＥ、ＮＲＰＭ）…（１０）
ここでは、点火無駄時間ＤＥＡＤＴＩＭＥを２００μｓｅｃとする。（１０）式は、エンジン回転速度ＮＲＰＭから点火無駄時間ＤＥＡＤＴＩＭＥに相当するクランク角である点火無駄時間相当クランク角ＩＧＮＤＥＡＤを算出するためのものである。
【００４８】
図１０は初期燃焼期間ＢＵＲＮ１［ｄｅｇ］を算出するためのもの、また図１２は主燃焼期間ＢＵＲＮ２［ｄｅｇ］を算出するためのもので、一定時間毎（例えば１０ｍｓｅｃ毎）に実行する。図１０、図１２は図５に続けて実行する。
図１０、図１２はどちらを先に実行してもかまわない。
【００４９】
まず図１０から説明すると、ステップ２１では、前回燃焼開始時期ＭＢＴＣＹＣＬ［ｄｅｇＢＴＤＣ］、図５のステップ１２で算出されている燃焼室５の吸気弁閉時期における容積ＶＩＶＣ［ｍ３］、図５のステップ１３で算出されている燃焼室５の吸気弁閉時期における温度ＴＩＮＩ［Ｋ］、エンジン回転速度ＮＲＰＭ［ｒｐｍ］、図５のステップ１４で算出されている反応確率ＲＰＲＯＢＡ［％］を読み込む。
【００５０】
ここで、前回燃焼開始時期ＭＢＴＣＹＣＬは、基本点火時期ＭＢＴＣＡＬの［ｄｅｇＢＴＤＣ］の１サイクル前の値であり、その算出については後述する。
【００５１】
ステップ２２では燃焼室５の燃焼開始時期における容積Ｖ０［ｍ３］を算出する。前述したように、ここでの点火時期（燃焼開始時期）は今回のサイクルで演算する基本点火時期ＭＢＴＣＡＬではなく基本点火時期の１サイクル前の値である。すなわち、基本点火時期の１サイクル前の値であるＭＢＴＣＹＣＬから次式により燃焼室５の燃焼開始時期における容積Ｖ０を算出する。
【００５２】
Ｖ０＝ｆ６（ＭＢＴＣＹＣＬ）…（１１）
具体的には前回燃焼開始時期ＭＢＴＣＹＣＬにおけるピストン６のストローク位置と、燃焼室５のボア径から、燃焼室５のＭＢＴＣＹＣＬにおける容積Ｖ０を算出する。図５のステップ１２では、燃焼室５の吸気弁閉時期ＩＶＣにおける容積ＶＩＶＣを、吸気弁閉時期をパラメータとする吸気弁閉時期容積のテーブルを検索することにより求めたが、ここではＭＢＴＣＹＣＬをパラメータとする前回燃焼開始時期容積のテーブルを検索することにより、燃焼室５の前回燃焼開始時期ＭＢＴＣＹＣＬにおける容積Ｖ０を求めればよい。
【００５３】
ステップ２３では燃焼開始時期における有効圧縮比Ｅｃを算出する。有効圧縮比Ｅｃは無次元の値であり、次式に示すように燃焼室５の燃焼開始時期における容積Ｖ０を燃焼室５の吸気弁閉時期における容積ＶＩＶＣで除した値である。
【００５４】
Ｅｃ＝ｆ７（Ｖ０、ＶＩＶＣ）＝Ｖ０／ＶＩＶＣ…（１２）
ステップ２４では吸気弁閉時期ＩＶＣから燃焼開始時期に至る間の燃焼室５内の温度上昇率ＴＣＯＭＰを次式に示すように有効圧縮比Ｅｃに基づいて算出する。
【００５５】
ＴＣＯＭＰ＝ｆ８（Ｅｃ）＝Ｅｃ＾（κ−１）…（１３）
ただし、κ：比熱比、
（１３）式は断熱圧縮されるガスの温度上昇率の式である。なお、（１３）式右辺の「＾」は累乗計算を表している。
【００５６】
κは断熱圧縮されるガスの定圧比熱を定容比熱で除した値で、断熱圧縮されるガスが空気であればκ＝１．４であり、簡単にはこの値を用いればよい。ただし、混合気に対してκの値を実験的に求めることで、一層の算出精度の向上が可能である。
【００５７】
図１１は（１３）式を図示したものである。従って、このような特性のテーブルを予めエンジンコントローラ３１のメモリに格納しておき、有効圧縮比Ｅｃに基づき当該テーブルを検索することにより温度上昇率ＴＣＯＭＰを求めることも可能である。
【００５８】
ステップ２５では、燃焼室５の燃焼開始時期における温度Ｔ０［Ｋ］を、燃焼室５の吸気弁閉時期における温度ＴＩＮＩに温度上昇率ＴＣＯＭＰを乗じることで、つまり
Ｔ０＝ＴＩＮＩ×ＴＣＯＭＰ…（１４）
の式により算出する。
【００５９】
ステップ２６では、次式（公知）により層流燃焼速度ＳＬ１［ｍ／ｓｅｃ］を算出する。
【００６０】
ＳＬ１＝ＳＬstd×（Ｔ０×Ｔstd）２．１８×（Ｐ０／Ｐstd）−０．１６…（１５）
ただし、Ｔstd ：基準温度［Ｋ］、
Ｐstd ：基準圧力［Ｐａ］、
ＳＬstd：基準温度Ｔstdと基準圧力Ｐstdにおける基準層流燃焼速 度［ｍ／ｓｅｃ］、
Ｔ０ ：燃焼室５の燃焼開始時期における温度［Ｋ］、
Ｐ０ ：燃焼室５の燃焼開始時期における圧力［Ｐａ］、
基準温度Ｔstdと基準圧力Ｐstdと基準層流燃焼速度ＳＬstdは実験により予め定められる値である。
【００６１】
燃焼室５の通常の圧力である２ｂａｒ以上の圧力下では、（１５）式の圧力項（Ｐ０／Ｐstd）−０．１６は小さな値となる。従って、圧力項（Ｐ０／Ｐstd）−０．１６を一定値として、基準層流燃焼速度ＳＬstdを基準温度Ｔstdのみで規定することも可能である。
【００６２】
従って、基準温度Ｔstdが５５０［Ｋ］で、基準層流燃焼速度ＳＬstdが１．０［ｍ／ｓｅｃ］で、圧力項が０．７である場合の燃焼開始時期における温度Ｔ０と層流燃焼速度ＳＬ１との関係は近似的に次式で定義することができる。
【００６３】

ステップ２７では、初期燃焼におけるガス流動の乱れ強さＵ１を算出する。このガス流動の乱れ強さＵ１は無次元の値であり、燃焼室５に流入する新気の流速と燃料インジェクタ２１の噴射燃料のペネトレーションとに依存する。
【００６４】
燃焼室５に流入する新気の流速は、吸気通路の形状と、吸気弁１５の作動状態と、吸気弁１５を設ける吸気ポート４の形状に依存する。噴射燃料のペネトレーションは燃料インジェクタ２１の噴射圧力と、燃料噴射期間と、燃焼噴射タイミングに依存する。
【００６５】
最終的に、初期燃焼におけるガス流動の乱れ強さＵ１は、エンジン回転速度ＮＲＰＭの関数として次式で表すことができる。
【００６６】
Ｕ１＝ｆ１０（ＮＲＰＭ）＝Ｃ１×ＮＲＰＭ…（１７）
ただし、Ｃ１：定数、
乱れ強さＵ１を回転速度ＮＲＰＭをパラメータとするテーブルから求めることも可能である。
【００６７】
ステップ２８では層流燃焼速度ＳＬ１と乱れ強さＵ１から、初期燃焼におけるガスの燃焼速度ＦＬＡＭＥ１［ｍ／ｓｅｃ］を次式により算出する。
【００６８】
ＦＬＡＭＥ１＝ＳＬ１×Ｕ１…（１８）
燃焼室５内にガス乱れがあるとガスの燃焼速度が変化する。（１８）式はこのガス乱れに伴う燃焼速度への寄与（影響）を考慮したものである。
【００６９】
ステップ２９では、次式により初期燃焼期間ＢＵＲＮ１［ｄｅｇ］を算出する。
【００７０】

ただし、ＡＦ１：火炎核の反応面積（固定値）［ｍ２］、
ここで、（１９）式右辺のＢＲ１は燃焼開始時期より初期燃焼期間ＢＵＲＮ１の終了時期までの燃焼質量割合の変化量であり、ここではＢＲ１＝２％に設定している。（１９）式右辺の（ＮＲＰＭ×６）は単位をｒｐｍからクランク角（ｄｅｇ）に変換するための措置である。火炎核の反応面積ＡＦ１は実験的に設定される。
【００７１】
次に図１２のフローに移ると、ステップ３１では回転速度ＮＲＰＭ、図５のステップ１４で算出されている反応確率ＲＰＲＯＢＡを読み込む。
【００７２】
ステップ３２では主燃焼におけるガス流動の乱れ強さＵ２を算出する。このガス流動の乱れ強さＵ２も初期燃焼におけるガス流動の乱れ強さＵ１と同様に、エンジン回転速度ＮＲＰＭの関数として次式で表すことができる。
【００７３】
Ｕ２＝ｆ１１（ＮＲＰＭ）＝Ｃ２×ＮＲＰＭ…（２０）
ただし、Ｃ２：定数、
乱れ強さＵ２を回転速度をパラメータとするテーブルから求めることも可能である。
【００７４】
ステップ３３では、層流燃焼速度ＳＬ２［ｍ／ｓｅｃ］と主燃焼におけるガス流動の乱れ強さＵ２とから、主燃焼における燃焼速度ＦＬＡＭＥ２［ｍ／ｓｅｃ］を次式により算出する。
【００７５】
ＦＬＡＭＥ２＝ＳＬ２×Ｕ２…（２１）
ただし、ＳＬ２：層流燃焼速度［ｍ／ｓｅｃ］、
（２１）式は（１８）式と同様、ガス乱れに伴う燃焼速度への寄与を考慮したものである。
【００７６】
前述のように主燃焼期間ＢＵＲＮ２の長さは燃焼室５内の温度や圧力の変化の影響を受けにくい。従って、層流燃焼速度ＳＬ２には予め実験的に求めた固定値を適用する。
【００７７】
ステップ３４では、主燃焼期間ＢＵＲＮ２［ｄｅｇ］を（１９）式に類似した次式で算出する。
【００７８】

ただし、Ｖ２：燃焼室５の主燃焼期間開始時容積［ｍ３］、
ＡＦ２：火炎核の反応面積［ｍ２］
ここで、（２２）式右辺のＢＲ２は主燃焼期間の開始時期より終了時期までの燃焼質量割合の変化量である。初期燃焼期間の終了時期に燃焼質量割合が２％になり、その後、主燃焼期間が開始し、燃焼質量割合が６０％に達して主燃焼期間が終了すると考えているので、ＢＲ２＝６０％−２％＝５８％を設定している。ＡＦ２は火炎核の成長行程における平均の反応面積であり、（１９）式のＡＦ１と同様に、予め実験的に定めた固定値とする。燃焼室５の主燃焼期間開始時における容積Ｖ２も固定値である。
【００７９】
図１３は基本点火時期ＭＢＴＣＡＬ［ｄｅｇＢＴＤＣ］を算出するためのもので、一定時間毎（例えば１０ｍｓｅｃ毎）に実行する。図１０、図１２のうち遅く実行されるフローに続けて実行する。
【００８０】
ステップ４１では、図１０のステップ２９で算出されている初期燃焼期間ＢＵＲＮ１、図１２のステップ３４で算出されている主燃焼期間ＢＵＲＮ２、図５のステップ１６で算出されている点火時期無駄時間相当クランク角ＩＧＮＤＥＡＤ、図５のステップ１５で算出されている基準クランク角θＰＭＡＸを読み込む。
【００８１】
ステップ４２では、初期燃焼期間ＢＵＲＮ１と主燃焼期間ＢＵＲＮ２の合計を燃焼期間ＢＵＲＮ［ｄｅｇ］として算出する。
【００８２】
ステップ４３では次式により基本点火時期ＭＢＴＣＡＬ［ｄｅｇＢＴＤＣ］を算出する。
【００８３】
ＭＢＴＣＡＬ＝ＢＵＲＮ−θＰＭＡＸ＋ＩＧＮＤＥＡＤ…（２３）
ステップ４４では、この基本点火時期ＭＢＴＣＡＬから点火無駄時間相当クランク角ＩＧＮＤＥＡＤを差し引いた値を前回燃焼開始時期ＭＢＴＣＹＣＬ［ｄｅｇＢＴＤＣ］として算出する。
【００８４】
このようにして算出した基本点火時期ＭＢＴＣＡＬは点火レジスタに移され、実際のクランク角がこの基本点火時期ＭＢＴＣＡＬと一致したタイミングでエンジンコントローラ３１より一次電流を遮断する点火信号が点火コイル１３に出力される。
【００８５】
また、今サイクルの点火時期指令値としてステップ４３で算出された基本点火時期ＭＢＴＣＡＬが用いられたとすると、次サイクルの点火時期になるまでの間、ステップ４４で算出された前回燃焼開始時期ＭＢＴＣＹＣＬが図１０のステップ２２において用いられる。
【００８６】
以上のように、本実施形態においては、燃焼室５内の未燃ガス量などの質量計算を行わずにＭＢＴの得られる点火時期である基本点火時期ＭＢＴＣＡＬを算出するので、計算負荷を小さく抑えることができる。
【００８７】
また、上記（１９）式に示したように初期燃焼期間ＢＵＲＮ１を、燃焼開始時期における燃焼室容積Ｖ０と、混合気の燃焼のしやすさを表す反応確率ＲＰＲＯＢＡと、燃焼速度ＦＬＡＭＥ１の関数で表している。ここで、燃焼開始時期における燃焼室容積Ｖ０が大きいほど、反応確率ＲＰＲＯＢＡが小さいほど、燃焼速度ＦＬＡＭＥ１が遅いほど、それぞれ初期燃焼期間ＢＵＲＮ１が長くなり、結果として基本点火時期ＭＢＴＣＡＬが進角する。
【００８８】
同様に、上記（２２）式に示したように主燃焼期間ＢＵＲＮ２を、主燃焼期間の開始時期における燃焼室容積Ｖ２と、混合気のしやすさを表す反応確率ＲＰＲＯＢＡと、燃焼速度ＦＬＡＭＥ２の関数で表している。ここで、主燃焼期間開始時期における燃焼室容積Ｖ２が大きいほど、反応確率ＲＰＲＯＢＡが小さいほど、燃焼速度ＦＬＡＭＥ２が遅いほど、それぞれ主燃焼期間ＢＵＲＮ２が長くなり、結果として点火時期ＭＢＴＣＡＬが進角する。
【００８９】
このように、燃焼期間ＢＵＲＮ１とＢＵＲＮ２を、燃焼期間に影響を与える様々なパラメータの関数として算出することで、燃焼期間ＢＵＲＮ１とＢＵＲＮ２を正確に算出することができる。結果として、燃焼期間ＢＵＲＮ１とＢＵＲＮ２に基づき算出される基本点火時期ＭＢＴＣＡＬも高精度に算出することができる。また、燃焼期間ＢＵＲＮを温度や圧力が大きく影響を受けやすい火炎核成長期間に相当する初期燃焼期間ＢＵＲＮ１と、温度や圧力の影響の少ない主燃焼期間とに分けて算出しているので、燃焼期間ＢＵＲＮの算出精度が向上する。燃焼期間ＢＵＲＮを３以上にさらに分割することで、算出精度のさらなる向上も可能である。
【００９０】
実施形態では、初期燃焼期間ＢＵＲＮ１の算出に用いる燃焼速度ＦＬＡＭＥ１を層流燃焼速度ＳＬ１と乱れ強さＵ１の積として、また主燃焼期間ＢＵＲＮ２の算出に用いる燃焼速度ＦＬＡＭＥ２を層流燃焼速度ＳＬ２と乱れ強さＵ２の積としてそれぞれ算出しているが、特開平１０−３０５３５号公報に記載されているように足し算による算出方法で求めても良い。
【００９１】
実施形態では、初期燃焼期間を燃焼質量割合でゼロから２％まで（つまりＢＲ１＝２％）、主燃焼期間を燃焼質量割合で２〜６０％まで（つまりＢＲ２＝５８％）と規定したが、本発明は必ずしもこの数値に限定されるものでない。
【００９２】
次に本発明の特徴である乱れ強さの算出方法について図１４と図１５に基づき、説明する。
【００９３】
図１４に基づいて吸気バルブ開（ＩＶＯ）時に燃焼期間での乱れ強さを推定する方法について説明する。この方法は、インテークマニフォールド圧力（以下、インマニ圧という。）、バルブタイミング、バルブリフト量から乱れ強さを算出する方法で、所定の間隔で算出される。
【００９４】
まず、インマニ圧Ｐｏ（Ｐａ）、吸気バルブの開弁期間θｏ（ｄｅｇ）、エンジン回転速度Ｎ（ｒｐｍ）からシリンダ内への吸入空気量Ｑｏ（ｍｍ³）を下式で算出する。
【００９５】
【数２４】

但し、上式中のｆ１（θｏ、Ｎ）は、予め実験等により算出した結果を図１６に示すようなマップとして記憶し、マップを参照して算出してもよい。
【００９６】
次に、吸気バルブの開弁期間θｏ中のバルブリフト量Ｌｏ（ｍｍ）から平均吸気バルブ開口面積Ｓｏ（ｍｍ²）を下式で算出する。
【００９７】
【数２５】

但し、上式中のｆ２（Ｌｏ）は、予め実験等により算出した結果を図１７に示すようなマップとして記憶し、マップを参照して算出してもよい。またバルブリフト量Ｌｏは、図１８に示すように吸気バルブ開弁期間θｏ中のリフト量を積算して、リフト量の積算値から平均値を求め、この平均値をリフト量Ｌｏとしてもよい。
【００９８】
算出したシリンダ内への吸入空気量Ｑｏ、平均吸気バルブ開口面積Ｓｏ、エンジン回転速度Ｎから吸気バルブ通過平均速度（吸入空気流速）Ｖｏ（ｍ／ｓｅｃ）を下式で算出する。
【００９９】
【数２６】

そして、吸気バルブ開状態で推定した燃焼期間の乱れ強さはＵｏは、
【０１００】
【数２７】

で算出される。
【０１０１】
ここで、Ｋｃ：スワールコントロールバルブ補正係数
Ｋｔ：タンブルコントロールバルブ補正係数
Ｋｑ：スキッシュ補正係数
であり、予め測定した値を用いる。また、構成にスワールコントロールバルブ、タンブルコントロールバルブ、スキッシュのいずれかの構成がない場合には、ない構成の補正係数を１として演算する。
【０１０２】
但し、上式中のｆ３（Ｖｏ、Ｎ）は、予め実験等により算出した結果を図１９に示すようなマップとして記憶し、マップを参照して算出してもよい。
【０１０３】
このように吸気バルブ開の状態で求めた燃焼期間の平均乱れ強さはＵｏは、シリンダ内への吸入空気量Ｑｏ、平均吸気バルブ開口面積Ｓｏ、エンジン回転速度Ｎから算出される吸気バルブ通過平均速度Ｖｏに基づき設定されるので、エンジンの過渡運転状態でシリンダ内への吸入空気量Ｑｏ、平均吸気バルブ開口面積Ｓｏ、エンジン回転速度Ｎが変化しても、変化に応じて乱れ強さＵｏを算出することができる。したがって、エンジンが過渡状態にある場合でも、エンジンの運転状態の変化に応じて乱れ強さを設定し、精度よく点火時期を算出することができる。
【０１０４】
図１５に基づいて吸気バルブ閉（ＩＶＣ）時の状態で燃焼期間での乱れ強さを推定する方法について説明する。この方法は、吸気バルブ開の状態での算出方法と同様に吸気バルブ閉時のインテークマニフォールド圧力（以下、インマニ圧という。）、バルブタイミング、バルブリフト量から乱れ強さを算出する方法で、所定のタイミングで算出される。
【０１０５】
まず、インマニ圧Ｐｃ（Ｐａ）、吸気バルブの開弁期間θｃ（ｄｅｇ）、エンジン回転速度Ｎ（ｒｐｍ）からシリンダ内への吸入空気量Ｑｃ（ｍｍ³）を下式で算出する。
【０１０６】
【数２８】

但し、上式中のｆ１（θｃ、Ｎ）は、予め実験等により算出した結果を図１６に示すようなマップとして記憶し、マップを参照して算出してもよい。
【０１０７】
次に、吸気バルブの開弁期間θｃ中のバルブリフト量Ｌｃ（ｍｍ）から平均吸気バルブ開口面積Ｓｃ（ｍｍ²）を下式で算出する。
【０１０８】
【数２９】

但し、上式中のｆ２（Ｌｃ）は、予め実験等により算出した結果を図１７に示すようなマップとして記憶し、マップを参照して算出してもよい。またバルブリフト量Ｌｃは、図１８に示すように吸気バルブ開弁期間θｃ中のリフト量を積算して、リフト量の積算値から平均値を求め、この平均値をリフト量Ｌｃとしてもよい。
【０１０９】
算出したシリンダ内への吸入空気量Ｑｃ、平均吸気バルブ開口面積Ｓｃ、エンジン回転速度Ｎから吸気バルブ通過平均速度（吸入空気流速）Ｖｃ（ｍ／ｓｅｃ）を下式で算出する。
【０１１０】
【数３０】

そして、吸気バルブ閉時の状態での燃焼期間の平均乱れ強さはＵｃは、
【０１１１】
【数３１】

で算出される。
【０１１２】
ここで、Ｋｃ：スワールコントロールバルブ補正係数
Ｋｔ：タンブルコントロールバルブ補正係数
Ｋｑ：スキッシュ補正係数
であり、予め測定した値を用いる。また、構成にスワールコントロールバルブ、タンブルコントロールバルブ、スキッシュのいずれかの構成がない場合には、ない構成の補正係数を１として演算する。
【０１１３】
但し、上式中のｆ３（Ｖｃ、Ｎ）は、予め実験等により算出した結果を図１９に示すようなマップとして記憶し、マップを参照して算出してもよい。
【０１１４】
このように吸気バルブ閉時の燃焼期間の乱れ強さＵｃは、シリンダ内への吸入空気量Ｑｃ、平均吸気バルブ開口面積Ｓｃ、エンジン回転速度Ｎから算出される吸気バルブ通過平均速度Ｖｃに基づき設定されるので、エンジンの過渡運転状態でシリンダ内への吸入空気量Ｑｃ、平均吸気バルブ開口面積Ｓｃ、エンジン回転速度Ｎが変化しても、変化に応じて乱れ強さＵｃを設定することができる。エンジンが過渡状態にある場合でも、エンジンの運転状態の変化に応じて乱れ強さを設定するし、精度よく点火時期を算出することができる。
【０１１５】
次に算出した平均乱れ強さＵｏ、Ｕｃを用いて初期燃焼期間での乱れ強さを算出する方法について図２０を用いて説明する。この演算は一定間隔で実行される。
【０１１６】
まず、平均乱れ強さＵｏから初期燃焼期間での乱れ強さＵｏ１を下式で算出する。
【０１１７】
【数３２】

但し、上式中のｆ４（Ｕｏ）は、予め実験等により算出した結果を図２１に示すようなマップとして記憶し、マップを参照して算出してもよい。
【０１１８】
次に、平均乱れ強さＵｃから初期燃焼期間での乱れ強さＵｃ１を下式で算出する。
【０１１９】
【数３３】

但し、上式中のｆ４（Ｕｃ）は、予め実験等により算出した結果を図２１に示すようなマップとして記憶し、マップを参照して算出してもよい。
【０１２０】
さらに算出した吸気バルブ開状態または閉状態に推定した乱れ強さに基づく初期燃焼期間での乱れ強さＵｏ１、Ｕｃ１から、初期燃焼期間の乱れ強さＵ１を下式から算出する。
【０１２１】
【数３４】

ここで、係数ｋ１は下式で算出される。
【０１２２】
【数３５】

但し、演算を実施する計算機の性能上、後述の演算が不可能な場合にはｋ１＝０．５と設定してもよい。
【０１２３】
係数ｋｐは、次式で算出される。
【０１２４】
【数３６】

ここで、Ｔｏ：吸気バルブの開き始め時間、Ｔｃ：吸気バルブの閉じ時間、Ｐａｖｅ：平均インマニ圧（Ｐａ）とする。
【０１２５】
但し、平均インマニ圧は式（３７）で算出できる。
【０１２６】
【数３７】

ここでＰ（ｔ）：インマニ圧力（Ｐａ）である。
【０１２７】
次に係数ｋθは、次式で算出される。
【０１２８】
【数３８】

ここで、Ｔｏ：吸気バルブの開き始め時間、Ｔｃ：吸気バルブの閉じ時間、θａｖｅ：平均バルブ開期間（ｄｅｇ）とする。
【０１２９】
但し、平均バルブ開期間は式（３９）で算出できる。
【０１３０】
【数３９】

ここでθ（ｔ）：バルブタイミングである。
【０１３１】
次に係数ｋ_Lは、次式で算出される。
【０１３２】
【数４０】

ここで、Ｔｏ：吸気バルブの開き始め時間、Ｔｃ：吸気バルブの閉じ時間、Ｌａｖｅ：平均リフト量（ｍｍ）とする。
【０１３３】
但し、平均リフト量は式（４１）で算出できる。
【０１３４】
【数４１】

ここでＬ（ｔ）：バルブリフト量（ｍｍ）である。
【０１３５】
次に算出した乱れ強さＵｏ、Ｕｃを用いて主燃焼期間での乱れ強さを算出する方法について図２２を用いて説明する。この演算は一定間隔で実行される。
【０１３６】
まず、乱れ強さＵｏから主燃焼期間での乱れ強さＵｏ２を下式で算出する。
【０１３７】
【数４２】

但し、上式中のｆ５（Ｕｏ）は、予め実験等により算出した結果を図２３に示すようなマップとして記憶し、マップを参照して算出してもよい。
【０１３８】
次に、乱れ強さＵｃから主燃焼期間での乱れ強さＵｃ２を下式で算出する。
【０１３９】
【数４３】

但し、上式中のｆ５（Ｕｃ）は、予め実験等により算出した結果を図２３に示すようなマップとして記憶し、マップを参照して算出してもよい。
【０１４０】
さらに算出した吸気バルブ開状態または閉状態に推定した乱れ強さに基づく初期燃焼期間での乱れ強さＵｏ１、Ｕｃ１から、主燃焼期間の乱れ強さＵ２を下式から算出する。
【０１４１】
【数４４】

ここで、係数ｋ１は前述と同様にして算出される。
【０１４２】
このようにして初期燃焼期間の乱れ強さＵ１と主燃焼期間の乱れ強さＵ２が算出され、これら乱れ強さに基づき式（１２）あるいは式（１６）により燃焼速度ＦＬＡＭＥ１、ＦＬＡＭＥ２が算出され、さらに燃焼期間ＢＵＲＮ、点火時期ＭＢＴＣＡＬが算出される。
【０１４３】
したがって、本発明においては、乱れ強さの算出にあたり、エンジンの過渡状態のときに運転状態の変化に対応する吸気バルブを通過する吸入空気流速を用いることで、内燃機関の過渡特性に応じて乱れ強さを変化するようにしたので、燃焼期間、さらには点火時期を精度よく算出し、ＭＢＴの得られる点火時期との差異を抑制することができる。
【０１４４】
図２４は、インマニ圧、シリンダ内吸入ガス量、バルブ通過流速、乱れ強さ等の変化を示すタイミングチャートである。
【０１４５】
＃２気筒を代表して説明すると＃２気筒の吸気バルブ開時ではインマニ圧が低い、つまり吸入ガス量が少ない状態であり、この吸気バルブ開状態で推定した吸入ガス量は少なく（図中一点鎖線で示す。）、吸気バルブ閉時にはインマニ圧が上昇し、この吸気バルブ閉状態で推定した吸入ガス量は上昇する（図中、実線で示す。）。吸気バルブ通過流速も吸入ガス量と同様の傾向を示し、また、乱れ強さも同様である。
【０１４６】
また、乱れ強さが大きくなることで、吸気バルブ閉時の燃焼速度も速くなるが、ＭＢＴ演算のための乱流燃焼速度は、吸気バルブ開状態で推定した燃焼速度と吸気バルブ閉時に推定した燃焼速度とから求められるため、図中破線で示すような燃焼速度となる。
【０１４７】
また図２５は、エンジンのある気筒での吸気行程中に吸気ポート圧力が上昇した際のシリンダ内圧の一例を示す。
【０１４８】
図中▲１▼の曲線は、吸気バルブ開時のエンジン状態が、吸気バルブが閉じるまで保持され、かつ、この状態に基づき算出されたＭＢＴで点火した場合のシリンダ内圧の変化を示す。
【０１４９】
図中▲２▼の曲線は、吸気バルブ閉時のエンジン状態が、吸気バルブ開の状態から吸気バルブが閉じるまで継続され、かつ、この状態に基づき算出されたＭＢＴで点火した場合のシリンダ内圧の変化を示す。
【０１５０】
なお、これら▲１▼と▲２▼のＭＢＴで点火した場合、基準クランク位置θＰｍａｘ（上死点後１２〜１５ｄｅｇクランク角度）で、シリンダ内圧は最大となる。
【０１５１】
▲４▼の曲線は、▲１▼のＭＢＴで点火した場合のシリンダ内圧の変化を示している。吸気ポート内の圧力が上昇するのに伴って、実際のエンジン状態での乱れ強さは、吸気バルブ開でのエンジン状態で算出した乱れ強さより上昇するため、燃焼速度も上昇する。したがって、吸気バルブ開での状態で算出したＭＢＴで点火した場合には、実際の要求点火時期より早過ぎ、▲４▼の曲線で示すようにノッキングを発生する恐れがある。
【０１５２】
対して、▲３▼の曲線は、これまで説明した算出方法により吸気バルブ開状態と閉状態とでの乱れ強さを考慮してＭＢＴを算出した場合のシリンダ内圧の変化を示したものである。燃焼速度と最大筒内圧は▲１▼と▲２▼の曲線の間に位置し、基準クランク位置θＰｍａｘで筒内圧最大となるが、ノックは発生しない。
【０１５３】
図２６は、ある気筒の吸気工程中に吸気ポート圧力が上昇した場合の燃焼質量割合の一例を示す図である。
【０１５４】
図中▲１▼の曲線は、吸気バルブ開のエンジン状態が、吸気バルブが閉じるまで保持され、かつ、この状態に基づき算出されたＭＢＴで点火した場合の燃焼質量割合の変化を示す。図中▲２▼の曲線は、吸気バルブ閉時のエンジン状態が、吸気バルブ開の状態から吸気バルブが閉じるまで継続され、かつ、この状態に基づき算出されたＭＢＴで点火した場合の燃焼質量割合の変化を示す。
【０１５５】
ＭＢＴで点火した場合は、基準クランク位置θＰｍａｘでの燃焼質量割合Ｒｍａｘはほぼ一定（６０％）であり、曲線▲１▼、▲２▼ともこの点を通過する。
【０１５６】
吸気ポート圧力が上昇するのに伴い、実際のエンジンの乱れ強さは、吸気バルブ開状態で推定した乱れ強さより上昇するため燃焼速度も上昇する（傾きが大きくなる。）。よって、吸気バルブ開状態で推定した乱れ強さに基づくＭＢＴで点火した場合、▲４▼の曲線のように基準クランク位置θＰｍａｘでの燃焼質量割合Ｒｍａｘは所定値より大きくなり、ノッキングを発生する恐れが生じる。
【０１５７】
一方、本発明で算出した乱れ強さを考慮したＭＢＴで点火した場合には、燃焼質量割合の変化は、▲３▼の曲線となり、基準クランク位置θＰｍａｘでの燃焼質量割合Ｒｍａｘは、所定値Ｒｍａｘとなる。
【０１５８】
以上説明した本実施の形態における点火時期制御装置の全体的な作用を説明する。吸気バルブが閉じられた時点のシリンダ内容積ＶＩＶＣと筒内温度ＴＩＮＩが算出される（ステップＳ１２、Ｓ１３）。そして燃焼ガスの燃焼のしやすさを所定条件における燃焼状態を１とした場合における比で示した反応確率ＲＰＲＯＢＡが算出される（ステップＳ１４）。
【０１５９】
燃焼速度は、筒内における雰囲気温度、燃焼速度の他に残留ガス率ＥＧＲＲＥＭ、水温ＴＷＫ、当量比に依存し、燃焼速度がこれらの値により変化する状態を実験で予め求めマップ化して保有しておくことができる。勿論、これらの因子に依存した相関関係を関数式でもつことも可能である。
【０１６０】
次に、ＭＢＴで点火した場合、燃焼ガスの圧力のピーク値である基準クランク角θＰｍａｘは略一定であり、エンジン高回転側で進角側にシフトする傾向があるためステップＳ１５の如く算出する。
【０１６１】
次にシリンダ内容積Ｖ₀、有効圧縮比Ｅｃから吸気バルブが閉じられた時点の筒内温度ＴＩＮＩからの温度上昇率を算出し、燃焼ガスが点火される時点の雰囲気温度Ｔ０が算出される（ステップＳ２１〜Ｓ２５）。算出した雰囲気温度Ｔ０に基づいて、層流燃焼速度ＳＬ１を算出する（ステップＳ２６）。
【０１６２】
燃焼速度は層流燃焼速度が早いほど早くなる傾向にあるため、筒内の乱流速度も燃焼速度に大きく依存する。乱流速度は吸気ポート形状等に左右され、一般にエンジン回転数が高くなるほど速くなる。このためこれを乱れ強さＵ１として算出し（ステップＳ２７）、層流燃焼速度ＳＬ１と掛け合わせて燃焼速度ＦＬＡＭＥ１を算出する（ステップＳ２８）。
【０１６３】
そして点火してから火炎核が形成されるまでの燃焼期間ＢＵＲＮ１を算出する（ステップＳ２９）。火炎核の形成には、燃焼ガス質量の内２％の燃焼によってなされることがわかっているため、燃焼ガス質量の２％を算出して用いる。火炎核形成段階では、点火時点のシリンダ内容積が大きいほど形成に時間が長くなるためこれを分子とし、反応確率ＲＰＲＯＢＡが大きいぼど時間が短く、燃焼速度が早いほど時間が短くなるため、これらを分母とした演算式を用いて演算することにより、正確且つ容易に火炎核形成段階の燃焼時間を求めることができる。
【０１６４】
火炎核が形成された後は、燃焼ガスが急速に燃焼される急速燃焼段階となり火炎核は急速に成長する。圧力がピークとなる６０％の燃焼ガスが燃焼される期間が演算される（ステップＳ３１〜ステップＳ３４）。ここでも火炎核の成長時点のシリンダ内容積が大きいほど形成に時間が長くなるためこれを分子とし、反応確率ＲＰＲＯＢＡが大きいほど時間が短く、燃焼速度が早いほど時間が短くなるため、これらを分母とした演算式を用いて演算することにより、正確且つ容易に火炎核成長段階の燃焼時間を求めることができる。
【０１６５】
以上の如く算出した火炎核形成段階の初期燃焼期間と火炎核成長段階の主燃焼期間とを足すことにより燃焼ガス質量６０％の燃焼に必要な期間を算出する事ができる（ステップＳ４２）。そして点火無駄時間とシリンダ内最大圧力となる基準クランク角θＰｍａｘを考慮してＭＢＴ点火時期を算出し、イグニッションコイルに出力指令を出すことができる（ステップＳ４３）。
【０１６６】
本実施の形態によれば、所定のクランク角までにシリンダ内で燃焼する燃焼ガス量を算出する燃焼ガス量算出手段と、燃焼ガスの層流状態での燃焼速度である層流燃焼速度を算出する層流燃焼速度算出手段と、エンジン過渡状態に応じて変化する燃焼の乱れを示す指標である乱れ強さを算出する乱れ強さ算出手段と、前記層流燃焼速度と乱れ強さからシリンダ内燃焼速度を算出し、この燃焼速度と前記シリンダ内の燃焼ガス量とから燃焼期間を算出し、この燃焼期間に基づいてＭＢＴの得られる基本点火時期を算出する点火時期演算手段とを備えたので、内燃機関の運転状態が過渡状態であっても、運転状態に応じて乱れ強さを設定するので、過渡状態においてもＭＢＴを得るための点火時期を精度よく算出することができる。
【０１６７】
また、燃焼期間を算出する際に、圧力・温度に対する燃焼期間の感度が大きい火炎核形成段階と、圧力・温度に対する燃焼期間の感度が鈍くなる主燃焼期間とに分けて演算することで、燃焼期間をより正確に演算することができる。この際、分子に燃焼ガス質量、シリンダ容積、分母に反応面積、燃焼速度といった演算式を用いており、同一次元の単位が分子、分母に用いられることが無いので無駄な演算をすることなく正確に算出できる。
【０１６８】
燃焼状態を無次元の反応確率で置き換えて演算に用いており、当該反応確率は実験等で容易に求めることができるため正確な燃焼期間を算出することができる。また、本実施の形態では、吸気バルブの作動角とリフト量を連続的に可変にできる可変動弁機構を備えた内燃機関へ適用した例を示したが、吸気バルブが閉じた時点のシリンダ内容積が変化しても容易にＭＢＴを算出することができるため、例えば、バルブ開閉に電磁石を用い、バルブタイミングを任意に変化させることができるいわゆる電磁駆動吸排気弁等の可変動弁システムへの適用も行える。なお、吸気バルブの作動角とリフト量を連続的に可変にできる可変動弁機構の一例として特開平１１−１０７７２５号公報、特開２００３−４１９７６号公報に記載のものがある。
【０１６９】
本実施形態では燃焼質量割合から燃焼期間を求めるにあたり、燃焼期間ＢＵＲＮを初期燃焼期間ＢＵＲＮ１（燃焼質量割合２％まで）と燃焼主期間ＢＵＲＮ２（燃焼質量割合２％から６０％まで）とに分けて演算し、点火時期ＭＢＴＣＹＣＬを求めたが、さらに細分化してもよく、または細分化を行った上で、燃焼期間（燃焼質量割合）をたとえば、初期燃焼期間ＢＵＲＮ１と燃焼主期間ＢＵＲＮ２とで分けて加算するようにしてもよい。細分化して燃焼期間を求めることによって、燃焼期間の算出精度は向上できる。
【０１７０】
本発明は、上記した実施形態に限定されるものではなく、本発明の技術的思想の範囲内でさまざまな変更がなしうることは明白である。
【図面の簡単な説明】
【図１】本発明の内燃機関のシステム図である。
【図２】エンジンコントローラで実行される点火時期制御のブロック図である。
【図３】燃焼室の圧力変化図である。
【図４】燃焼質量割合の変化を説明する特性図である。
【図５】物理量の算出を説明するためのフローチャートである。
【図６】エンジンのクランクシャフトとコネクティングロッドの位置関係を説明するダイアグラムである。
【図７】水温補正係数の特性図である。
【図８】当量比補正係数の特性図である。
【図９】基準クランク角の特性図である。
【図１０】初期燃焼期間の算出を説明するためのフローチャートである。
【図１１】温度上昇率の特性図である。
【図１２】主燃焼期間の算出を説明するためのフローチャートである。
【図１３】基本点火時期の算出を説明するためのフローチャートである。
【図１４】吸気バルブ開状態で推定する平均乱れ強さの算出方法を説明する図である。
【図１５】吸気バルブ閉状態で推定する平均乱れ強さの算出方法を説明する図である。
【図１６】変数ｆ１を算出するためのマップの一例である。
【図１７】変数ｆ２を算出するためのマップの一例である。
【図１８】吸気バルブの開弁期間とリフト量を説明するための図である。
【図１９】変数ｆ３を算出するためのマップの一例である。
【図２０】初期燃焼期間の乱れ強さの算出方法を説明する図である。
【図２１】変数ｆ４を算出するためのマップの一例である。
【図２２】主燃焼期間の乱れ強さの算出方法を説明する図である。
【図２３】変数ｆ５を算出するためのマップの一例である。
【図２４】乱れ強さの変化を示すタイミングチャートである。
【図２５】シリンダ内圧の変化を説明するための図である。
【図２６】燃焼割合の変化を説明するための図である。
【符号の説明】
１ 内燃機関
２ 吸気コレクタ
３ インジェクタ
４ 吸気マニホールド
５ 排気マニホールド
６ マフラ
７ エアクリーナ
８ 電子制御スロットル
９ 吸気ダクト
２０ コントロールユニット[0001]
BACKGROUND OF THE INVENTION
The present invention relates to an ignition timing control device for an internal combustion engine.
[0002]
[Prior art]
As a conventional ignition timing control device for an internal combustion engine, a minimum ignition advance value (hereinafter referred to as MBT) for obtaining the best torque from the combustion time in the cylinder can be obtained without measuring the pressure in the cylinder. There is one that calculates the ignition timing (for example, see Patent Document 1).
[0003]
This conventional technique calculates the total gas mass (MASSC) and unburned gas density (ROU) in the cylinder when calculating the MBT calculation value, and then calculates MASSC / ROU to minimize the ratio of unburned gas. MBT is obtained from the ignition timing. Since the unburned gas density is a value obtained by dividing the unburned gas mass by the unburned gas volume, it depends on changes in the mass and volume of the unburned gas. Theoretically, if these values are detected, the mass of unburned gas can be accurately obtained. However, since it is difficult to estimate the volume of unburned gas in the cylinder, there is The unburned gas density is calculated based on the charging efficiency ηc.
[0004]
[Patent Document 1]
Japanese Patent Laid-Open No. 10-30535
[0005]
[Problems to be solved by the invention]
However, the function of only the charging efficiency ηc corresponding to the mass cannot accurately calculate the unburned gas density, and it is difficult to correct the unburned gas volume that varies depending on the operating conditions. Remains. Further, since both MASSC and ROU are masses, and the values obtained by dividing the masses are dimensionless, it is wasteful to compute these mass values individually and further divide them. This increases the calculation load.
[0006]
[Means for Solving the Problems]
The present invention relates to an ignition timing control device for an internal combustion engine. Calculating a combustion speed, calculating a combustion period from the combustion speed and the amount of combustion gas in the cylinder, calculating a basic ignition timing for obtaining MBT based on the combustion period; Calculate the laminar combustion speed, which is the combustion speed in the laminar state of the combustion gas, calculate the turbulence intensity, which is an index indicating the combustion turbulence that changes according to the engine transient state, Calculate the combustion speed of the combustion gas in the cylinder from the turbulence intensity, calculate the volume in the cylinder at the start of combustion gas combustion, calculate the amount of combustion gas combusted in the cylinder by the predetermined crank angle, A reaction probability indicating the ease of combustion of the combustion gas in the cylinder with respect to the ease of combustion of the combustion gas in the cylinder, and calculating the combustion speed, the volume in the cylinder, the amount of combustion gas, and the reaction probability, Based on this, the combustion period is calculated, and based on this combustion period, the basic ignition timing at which MBT is obtained is calculated It is characterized by doing.
[0007]
SUMMARY OF THE INVENTION Accordingly, an object of the present invention is to provide an ignition timing control device for an internal combustion engine that solves the above problems and calculates the ignition timing at which MBT can be obtained more accurately.
[0008]
[Means for Solving the Problems]
The present invention relates to an ignition timing control device for an internal combustion engine. A combustion speed is calculated, a combustion period is calculated from the combustion speed and the amount of combustion gas in the cylinder, and a basic ignition timing at which MBT is obtained is calculated based on the combustion period.
[0009]
【The invention's effect】
According to the present invention, even if the operating state of the internal combustion engine is in a transient state, the turbulence intensity is set according to the operating state, so that the ignition timing for obtaining MBT can be accurately calculated even in the transient state. it can.
[0010]
DETAILED DESCRIPTION OF THE INVENTION
FIG. 1 is a schematic diagram for explaining the system of the present invention. Intake air is introduced into an internal combustion engine 1 via an intake manifold 2 in which an intake collector 2 and an injector 3 are arranged. Exhaust gas is discharged into the atmosphere from the exhaust manifold 5 provided with the three-way catalyst 21 through the muffler 6.
[0011]
Air purified by an air cleaner 7 is supplied to the intake collector 2 through an intake duct 9. An electronically controlled throttle 8 having a throttle opening sensor and a throttle motor is provided upstream of the intake collector 2.
[0012]
The fuel gas evaporated in the fuel tank 10 is sent to the intake collector 2 together with the air introduced into the canister 11 after being introduced into the canister 11. Further, an EGR passage 12 that connects the three-way catalyst 21 downstream and the intake collector 2 is installed, and a part of the exhaust gas is recirculated into the intake air.
[0013]
The internal combustion engine 1 is provided with a position sensor 13 for detecting the position of the crank angle and a water temperature sensor 14 for detecting the cooling water temperature. The intake duct 9 has an air flow meter 16 for detecting the amount of intake air incorporating the intake air temperature sensor. Provided. Instead of the air flow meter 16, a sensor for detecting the temperature and pressure of the intake manifold 4 may be provided, and the intake air amount may be calculated based on these sensors.
[0014]
The exhaust manifold 5 is provided with an oxygen concentration sensor 17 for detecting the oxygen concentration in the exhaust gas. An EGR control valve 18 is provided in the middle of the EGR passage 12, and a canister purge air control valve 19 for controlling the purge air discharged from the canister 11. Is installed.
[0015]
The internal combustion engine 1 is further provided with a variable valve mechanism 22 that can continuously change the operating angle and lift amount of the intake valve, and an operating angle sensor 15 that detects the actual operating angle of the intake valve. Based on the output signal of the angle sensor 15, the opening timing (IVO) and closing timing (IVC) of the intake valve can be calculated.
[0016]
The output signals of the position sensor 13, the water temperature sensor 14, the operating angle sensor 15, the air flow meter 16 and the intake air temperature sensor incorporated therein are input to the control unit 20, and further from the output signal of the oxygen concentration sensor 17 and a battery (not shown). The voltage signal is also input.
[0017]
Based on these input data, the control unit 20 outputs an ignition timing control signal to the ignition coil, an air-fuel ratio control signal to the fuel injector 3, an intake valve operating angle control signal to the variable valve mechanism 22, and canister control. A signal is output to the canister purge control valve 19 and an EGR control signal is output to the EGR control valve 18 to maintain the internal combustion engine 1 in an appropriate operating state.
[0018]
FIG. 2 is a block diagram of the ignition timing control performed in the engine controller 31. The ignition timing control section 51 and the ignition timing control section 61 are mainly composed. The ignition timing calculation unit 51 further includes an initial combustion period calculation unit 52, a main combustion period calculation unit 53, a combustion period calculation unit 54, a basic ignition timing calculation unit 55, a previous combustion start timing calculation unit 56, and an ignition timing command value calculation unit 57. Become.
[0019]
The initial combustion period calculation unit 52 calculates a period from when the air-fuel mixture is ignited until flame nuclei are formed as the initial combustion period BURN1. The main combustion period calculation unit 53 calculates a period from when the flame kernel is formed until the combustion pressure reaches the maximum value Pmax as the main combustion period BURN2. The combustion period calculation unit 54 calculates the sum of the initial combustion period BURN1 and the main combustion period BURN2 as a combustion period BURN from ignition to the maximum combustion pressure Pmax. The basic ignition timing calculation unit 55 calculates an ignition timing (this ignition timing is referred to as “basic ignition timing”) MBTCAL from which MBT is obtained based on the combustion period BURN.
[0020]
The ignition timing control unit 61 controls the energization angle and the non-energization angle to the ignition coil 13 so that the ignition plug 14 ignites the air-fuel mixture in the combustion chamber 5 at the basic ignition timing MBTCAL calculated in this way. To do.
[0021]
As described above, the combustion period BURN is calculated by dividing the combustion period BURN into the initial combustion period BURN1 and the main combustion period BURN2, and the basic ignition timing MBTCAL is obtained according to the combustion period BURN, based on the result obtained from the combustion analysis. Is. Hereinafter, the ignition timing control based on the combustion analysis will be further described.
[0022]
As shown in FIG. 3, the crank angle at which the combustion pressure of the air-fuel mixture reaches the maximum value Pmax when the air-fuel mixture is ignited with MBT (minimum advance angle value at which maximum torque is obtained) is defined as a reference crank angle θPMAX [degATDC]. The reference crank angle θPMAX is substantially constant regardless of the combustion method, and is generally in the range of 12 to 15 degrees after compression top dead center and at most 10 to 20 degrees after compression top dead center.
[0023]
FIG. 4 shows changes in the combustion mass ratio R obtained by the combustion analysis in the combustion chamber in the spark ignition engine. The combustion mass ratio R representing the ratio of the combustion mass to the fuel supplied to the combustion chamber is 0% at the time of ignition and reaches 100% by complete combustion. The combustion mass ratio Rmax at the reference crank angle θPMAX is constant and about 60%.
[0024]
The combustion period until the combustion mass ratio R reaches 0% to approximately 60% corresponding to the reference crank angle θPMAX is an initial combustion period in which there is almost no change in the combustion mass ratio or the combustion pressure immediately after ignition, and the combustion It can be divided into a mass ratio and a main combustion period in which the combustion pressure increases rapidly. The initial combustion period is a stage from the start of combustion until flame nuclei are formed, and the flame nuclei are formed at a timing of 2% to 10% in terms of the combustion mass ratio. During this period, the rate of increase in combustion pressure and combustion temperature is small, and the initial combustion period is long with respect to changes in the combustion mass ratio. The length of the initial combustion period is susceptible to changes in temperature and pressure in the combustion chamber.
[0025]
On the other hand, in the main combustion period, the flame propagates from the flame kernel to the outside, and the combustion speed increases rapidly. Therefore, the change in the combustion mass ratio during the main combustion period is larger than the change in the combustion mass ratio during the initial combustion period.
[0026]
In the engine controller 31, the period until the combustion mass ratio reaches 2% is defined as the initial combustion period BURN1 [deg], and after the end of the initial combustion period BURN1, the interval until the reference crank angle θPMAX is reached (combustion chamber volume ratio is 2). To about 60%) is distinguished as the main combustion period BURN2 [deg]. Then, a combustion period BURN [deg] that is the sum of the initial combustion period BURN1 and the main combustion period BURN2 is calculated, a reference crank angle θPMAX [degATDC] is subtracted from the combustion period BURN, and an ignition dead time equivalent crank described later is further calculated. The crank angle position to which the angle IGNDEAD [deg] is added is set as the basic ignition timing MBTCAL [degBTDC], which is the ignition timing at which MBT is obtained.
[0027]
The pressure and temperature in the combustion chamber 5 during the initial combustion period in which flame nuclei are formed are almost equivalent to the pressure and temperature at the time of ignition, but the ignition timing is calculated from this, but accurate ignition is performed from the beginning. The time cannot be set. Therefore, as shown in FIG. 2, the previous combustion start timing calculation unit 56 calculates the previous value of the basic ignition timing as the previous combustion start timing MBTCYCL [degBTDC], and gives this value to the initial combustion period calculation unit 52. In addition, the initial combustion period calculation unit 52 cyclically repeats the calculation of the initial combustion period so as to obtain a highly accurate result without a time delay.
[0028]
Next, the calculation of the ignition timing command value QADV executed by the engine controller 31 will be described in detail with reference to the following flowchart.
[0029]
FIG. 5 is for calculating various physical quantities necessary for calculating the ignition timing, and is executed at regular time intervals (for example, every 10 msec).
[0030]
First, in step 11, the intake valve closing timing IVC [degBTDC], the collector internal temperature TCOL [K] detected by the temperature sensor 43, the exhaust gas temperature TEXH [K] detected by the temperature sensor 45, and the internal inert gas rate MRESFR [ %], The coolant temperature TWK [K] detected by the temperature sensor 37, the target equivalence ratio TFBYA, the engine speed NRPM [rpm] detected by the crank angle sensor, and the dead time DEADTIME [μsec].
[0031]
Here, the crank angle sensor includes a position sensor 33 that detects the position of the crankshaft 7 and an operating angle sensor 34 that detects the position of the intake camshaft 25. Based on signals from these two sensors 33 and 34, The engine speed NRPM [rpm] is calculated.
[0032]
The intake valve closing timing IVC is known from the command value given to the intake VTC mechanism 27. Alternatively, the actual intake valve closing timing may be detected by the operating angle sensor 34.
[0033]
The internal inert gas ratio MRESFR is a value obtained by dividing the amount of inert gas remaining in the combustion chamber by the total amount of gas in the combustion chamber, and the calculation thereof will be described later. The ignition dead time DEADTIME is a constant value.
[0034]
The target equivalent ratio TFBYA is calculated in a fuel injection amount calculation flow (not shown). The target equivalent ratio TFBYA is an unnamed number, and is a value represented by the following expression when the theoretical air-fuel ratio is 14.7.
[0035]
TFBYA = 14.7 / target air-fuel ratio (1)
For example, from equation (1), when the target air-fuel ratio is the stoichiometric air-fuel ratio, TFBYA = 1.0, and when the target air-fuel ratio is a lean value such as 22.0, TFBYA is a positive value less than 1.0. is there.
[0036]
In step 12, the volume of the combustion chamber 5 at the intake valve closing timing IVC (that is, the volume at the compression start timing) VIVC [m3] is calculated. The volume VIVC of the combustion chamber 5 when the intake valve is closed is determined by the stroke position of the piston 6. The stroke position of the piston 6 is determined by the crank angle position of the engine.
[0037]
Referring to FIG. 6, consider a case where the rotation center 72 of the crankshaft 71 of the engine is offset from the center axis 73 of the cylinder. Assume that the connecting rod 74, the connecting point 74 between the connecting rod 74 and the crankshaft 71, and the piston pin 76 that connects the connecting rod 74 and the piston are in the relationship shown in the figure. At this time, the volume VIVC of the combustion chamber 5 at the closing timing of the intake valve can be expressed by the following equations (2) to (6).
[0038]

Where Vc: gap volume [m3],
ε: compression ratio,
D: cylinder bore diameter [m],
ST: Full piston stroke [m],
H: distance [m] of the piston pin 76 from the TDC,
CND: length of connecting rod 74 [m],
CRoff: offset distance [m] of the nodal point 75 from the cylinder center axis 73,
PISoff: Crankshaft rotation center 72 offset distance [m] from cylinder center axis 73
θivc: Intake valve closing timing crank angle [degATDC],
θoff: an angle [deg] between a line connecting the piston pin 76 and the crankshaft rotation center 72 and a vertical line in TDC,
X: horizontal distance [m] between the nodal point 75 and the piston pin 76,
As described above, the crank angle θivc at the time of closing the intake valve is known because it is determined by the command signal from the engine controller 31 to the intake VTC mechanism 27. By substituting the crank angle θivc (= IVC) at this time into the equations (2) to (6), the volume VIVC of the combustion chamber 5 at the closing timing of the intake valve can be calculated. Therefore, in practice, the volume VIVC of the combustion chamber 5 at the intake valve closing timing is set by a table using the intake valve closing timing IVC as a parameter. When the intake VTC mechanism 27 is not provided, a constant value can be given.
[0039]
In step 13, the temperature (that is, the compression start timing temperature) TINI [K] of the combustion chamber 5 at the intake valve closing timing IVC is calculated. The temperature TINI of the combustion chamber 5 at the intake valve closing timing IVC is a temperature of a gas in which fresh air flowing into the combustion chamber 5 and inert gas remaining in the combustion chamber 5 are mixed, and fresh air flowing into the combustion chamber 5. Is equal to the fresh air temperature TCOL in the intake collector 2 and the temperature of the inert gas remaining in the combustion chamber 5 can be approximated by the exhaust temperature TEXH in the vicinity of the exhaust port portion. From the TCOL, the exhaust gas temperature TEXH, and the internal inert gas ratio MRESFR which is the ratio of the inert gas remaining in the combustion chamber 5, it can be obtained by the following equation.
[0040]

In step 14, a reaction probability RPROBA [%] representing the flammability of the air-fuel mixture in the combustion chamber 5 is calculated. The reaction probability RPROBA is a dimensionless value and depends on the three parameters of the residual inert gas ratio MRESFR, the cooling water temperature TWK [K], and the target equivalent ratio TFBYA, and can be expressed by the following equation.
[0041]
RPROBA = f3 (MRESFR, TWK, TFBYA) (8)
More specifically, the maximum value of the reaction probability obtained by the combination of the three parameters MRESFR, TWK, and TFBYA is set to 100%, the relationship between these parameters and the reaction probability RPROBA is experimentally obtained, and the obtained reaction probability RPROBA is obtained. Are stored in advance in the memory of the engine controller 31 as a table corresponding to the parameters. In step 14, the reaction probability RPROBA is obtained by searching this table according to the parameters.
[0042]
Specifically, a table of water temperature correction coefficients having characteristics as shown in FIG. 7 according to the cooling water temperature TWK, a table of internal inert gas rate correction coefficients (not shown) set similarly, and a target equivalent ratio A table of equivalence ratio correction coefficients having characteristics as shown in FIG. 8 according to Tfbya is stored in advance in the memory. The maximum value of each correction coefficient is 1.0, and the reaction probability RPROBA is calculated by multiplying the product of the three types of correction coefficients by the maximum value of 100% of the reaction probability.
[0043]
Explaining each table, the water temperature correction coefficient shown in FIG. 7 becomes larger as the cooling water temperature TWK is higher, and becomes 1.0 when the cooling water temperature TWK is 80 ° C. or higher. The equivalence ratio correction coefficient shown in FIG. 8 is the maximum value of 1.0 when the target equivalence ratio TFBYA is 1.0, that is, the stoichiometric air-fuel ratio. The ratio correction factor decreases. Although the internal inert gas rate correction coefficient is not shown, it is 1.0 when the internal inert gas rate MRESFR is zero.
[0044]
In step 15, a reference crank angle θPMAX [degATDC] is calculated. As described above, the reference crank angle θPMAX does not fluctuate very much, but still has a tendency to advance as the engine speed NRPM increases. Therefore, the reference crank angle θPMAX can be expressed by the following equation as a function of the engine speed NRPM. it can.
[0045]
θPMAX = f4 (NRPM) (9)
Specifically, the reference crank angle θPMAX is obtained by searching a table of characteristics shown in FIG. 9 stored in advance in the memory of the engine controller 31 from the engine speed NRPM. In order to facilitate calculation, the reference crank angle θPMAX can be regarded as constant.
[0046]
Finally, in step 16, the ignition dead time equivalent crank angle IGNDEAD [deg] is calculated. The crank angle IGNDEAD corresponding to the ignition dead time is a crank angle section from the timing at which the ignition command signal, that is, the signal for cutting off the primary current of the ignition coil 13 is output from the engine controller 31 until the ignition plug 14 actually ignites. Can be represented by
[0047]
IGNDEAD = f5 (DEADTIME, NRPM) (10)
Here, the ignition dead time DEADTIME is set to 200 μsec. Equation (10) is for calculating the ignition dead time equivalent crank angle IGNDEAD that is the crank angle corresponding to the ignition dead time DEADTIME from the engine speed NRPM.
[0048]
FIG. 10 is for calculating the initial combustion period BURN1 [deg], and FIG. 12 is for calculating the main combustion period BURN2 [deg], which is executed at regular intervals (for example, every 10 msec). 10 and 12 are executed following FIG.
Either of FIGS. 10 and 12 may be executed first.
[0049]
First, referring to FIG. 10, in step 21, the previous combustion start timing MBTCYCL [degBTDC], the volume VIVC [m3] at the intake valve closing timing of the combustion chamber 5 calculated in step 12 of FIG. 5, and step 13 of FIG. The temperature TINI [K] at the closing timing of the intake valve of the combustion chamber 5 calculated in step S5, the engine speed NRPM [rpm], and the reaction probability RPROBA [%] calculated in step 14 of FIG.
[0050]
Here, the previous combustion start timing MBTCYCL is a value one cycle before [degBTDC] of the basic ignition timing MBTCAL, and the calculation thereof will be described later.
[0051]
In step 22, the volume V0 [m3] at the combustion start timing of the combustion chamber 5 is calculated. As described above, the ignition timing (combustion start timing) here is not the basic ignition timing MBTCAL calculated in the current cycle but a value one cycle before the basic ignition timing. That is, the volume V0 at the combustion start timing of the combustion chamber 5 is calculated from MBTCYCL, which is a value one cycle before the basic ignition timing, by the following equation.
[0052]
V0 = f6 (MBTCYCL) (11)
Specifically, the volume V0 of MBTCYCL in the combustion chamber 5 is calculated from the stroke position of the piston 6 at the previous combustion start timing MBTCYCL and the bore diameter of the combustion chamber 5. In step 12 of FIG. 5, the volume VIVC of the combustion chamber 5 at the intake valve closing timing IVC is obtained by searching a table of intake valve closing timing volumes using the intake valve closing timing as a parameter. Here, MBTCYCL is set as a parameter. The volume V0 of the combustion chamber 5 at the previous combustion start time MBTCYCL may be obtained by searching the table of the previous combustion start time volume.
[0053]
In step 23, an effective compression ratio Ec at the combustion start timing is calculated. The effective compression ratio Ec is a dimensionless value, and is a value obtained by dividing the volume V0 of the combustion chamber 5 at the combustion start timing by the volume VIVC of the combustion chamber 5 at the intake valve closing timing, as shown in the following equation.
[0054]
Ec = f7 (V0, VIVC) = V0 / VIVC (12)
In step 24, the temperature increase rate TCOMP in the combustion chamber 5 from the intake valve closing timing IVC to the combustion start timing is calculated based on the effective compression ratio Ec as shown in the following equation.
[0055]
TCOMP = f8 (Ec) = Ec ^ (κ−1) (13)
Where κ: specific heat ratio,
Equation (13) is an equation for the rate of temperature rise of the adiabatic compressed gas. Note that “^” on the right side of the equation (13) represents power calculation.
[0056]
κ is a value obtained by dividing the constant pressure specific heat of the gas adiabatically compressed by the constant volume specific heat. If the gas adiabatically compressed is air, κ = 1.4, and this value may be used simply. However, the calculation accuracy can be further improved by experimentally determining the value of κ for the air-fuel mixture.
[0057]
FIG. 11 illustrates equation (13). Therefore, it is possible to obtain the temperature increase rate TCOMP by storing a table of such characteristics in advance in the memory of the engine controller 31 and searching the table based on the effective compression ratio Ec.
[0058]
In step 25, the temperature T0 [K] at the combustion start timing of the combustion chamber 5 is multiplied by the temperature TINI at the intake valve closing timing of the combustion chamber 5 by the temperature increase rate TCOMP, that is,
T0 = TINI × TCOMP (14)
It is calculated by the following formula.
[0059]
In step 26, a laminar combustion speed SL1 [m / sec] is calculated by the following equation (known).
[0060]
SL1 = SLstd × (T0 × Tstd) 2.18 × (P0 / Pstd) −0.16 (15)
Where Tstd: reference temperature [K],
Pstd: Reference pressure [Pa]
SLstd: reference laminar combustion speed [m / sec] at reference temperature Tstd and reference pressure Pstd,
T0: temperature [K] at the combustion start timing of the combustion chamber 5;
P0: pressure [Pa] at the combustion start timing of the combustion chamber 5;
The reference temperature Tstd, the reference pressure Pstd, and the reference laminar combustion speed SLstd are values determined in advance by experiments.
[0061]
Under a pressure of 2 bar or more, which is a normal pressure in the combustion chamber 5, the pressure term (P0 / Pstd) −0.16 in the equation (15) becomes a small value. Accordingly, it is also possible to define the reference laminar combustion speed SLstd only by the reference temperature Tstd with the pressure term (P0 / Pstd) −0.16 as a constant value.
[0062]
Accordingly, when the reference temperature Tstd is 550 [K], the reference laminar combustion rate SLstd is 1.0 [m / sec], and the pressure term is 0.7, the temperature T0 and the laminar combustion rate at the combustion start timing The relationship with SL1 can be approximately defined by the following equation.
[0063]

In step 27, the turbulence intensity U1 of the gas flow in the initial combustion is calculated. The turbulence intensity U1 of the gas flow is a dimensionless value and depends on the flow rate of fresh air flowing into the combustion chamber 5 and the penetration of the injected fuel from the fuel injector 21.
[0064]
The flow rate of fresh air flowing into the combustion chamber 5 depends on the shape of the intake passage, the operating state of the intake valve 15, and the shape of the intake port 4 where the intake valve 15 is provided. The penetration of the injected fuel depends on the injection pressure of the fuel injector 21, the fuel injection period, and the combustion injection timing.
[0065]
Finally, the turbulence intensity U1 of the gas flow in the initial combustion can be expressed by the following equation as a function of the engine speed NRPM.
[0066]
U1 = f10 (NRPM) = C1 × NRPM (17)
Where C1: constant,
It is also possible to obtain the turbulence intensity U1 from a table using the rotational speed NRPM as a parameter.
[0067]
In step 28, the gas combustion rate FLAME1 [m / sec] in the initial combustion is calculated from the laminar combustion rate SL1 and the turbulence intensity U1 by the following equation.
[0068]
FLAME1 = SL1 × U1 (18)
If there is gas turbulence in the combustion chamber 5, the gas combustion speed changes. Equation (18) takes into account the contribution (influence) to the combustion speed associated with this gas turbulence.
[0069]
In step 29, the initial combustion period BURN1 [deg] is calculated by the following equation.
[0070]

However, AF1: Reaction area of flame kernel (fixed value) [m2],
Here, BR1 on the right side of the equation (19) is the amount of change in the combustion mass ratio from the combustion start timing to the end timing of the initial combustion period BURN1, and here BR1 = 2% is set. (NRPM × 6) on the right side of equation (19) is a measure for converting the unit from rpm to crank angle (deg). The reaction area AF1 of the flame kernel is set experimentally.
[0071]
Next, in the flow of FIG. 12, in step 31, the rotational speed NRPM and the reaction probability RPROBA calculated in step 14 of FIG. 5 are read.
[0072]
In step 32, the turbulence intensity U2 of the gas flow in the main combustion is calculated. The turbulence intensity U2 of the gas flow can be expressed by the following equation as a function of the engine speed NRPM, similarly to the turbulence intensity U1 of the gas flow in the initial combustion.
[0073]
U2 = f11 (NRPM) = C2 × NRPM (20)
Where C2 is a constant,
It is also possible to obtain the turbulence intensity U2 from a table using the rotation speed as a parameter.
[0074]
In step 33, the combustion speed FLAME2 [m / sec] in the main combustion is calculated from the laminar combustion speed SL2 [m / sec] and the gas flow turbulence intensity U2 in the main combustion by the following equation.
[0075]
FLAME2 = SL2 × U2 (21)
However, SL2: Laminar burning velocity [m / sec],
Equation (21) considers the contribution to the combustion speed associated with gas turbulence, as in Equation (18).
[0076]
As described above, the length of the main combustion period BURN2 is not easily affected by changes in temperature and pressure in the combustion chamber 5. Therefore, a fixed value obtained experimentally in advance is applied to the laminar combustion speed SL2.
[0077]
In step 34, the main combustion period BURN2 [deg] is calculated by the following equation similar to the equation (19).
[0078]

V2: Volume at the start of the main combustion period of the combustion chamber 5 [m3],
AF2: Reaction area of flame kernel [m2]
Here, BR2 on the right side of Equation (22) is the amount of change in the combustion mass ratio from the start timing to the end timing of the main combustion period. Since the combustion mass ratio becomes 2% at the end of the initial combustion period, and then the main combustion period starts and the combustion mass ratio reaches 60% and the main combustion period ends, BR2 = 60% − 2% = 58% is set. AF2 is an average reaction area in the growth process of the flame kernel, and is set to a fixed value experimentally determined in advance, like AF1 in the equation (19). The volume V2 at the start of the main combustion period of the combustion chamber 5 is also a fixed value.
[0079]
FIG. 13 is for calculating the basic ignition timing MBTCAL [degBTDC], and is executed at regular intervals (for example, every 10 msec). This is executed following the flow that is executed later in FIGS.
[0080]
In step 41, the initial combustion period BURN1 calculated in step 29 of FIG. 10, the main combustion period BURN2 calculated in step 34 of FIG. 12, and the ignition timing dead time equivalent crank calculated in step 16 of FIG. The angle IGNDEAD, the reference crank angle θPMAX calculated in step 15 of FIG. 5, is read.
[0081]
In step 42, the sum of the initial combustion period BURN1 and the main combustion period BURN2 is calculated as the combustion period BURN [deg].
[0082]
In step 43, the basic ignition timing MBTCAL [degBTDC] is calculated by the following equation.
[0083]
MBTCAL = BURN−θPMAX + IGNDEAD (23)
In step 44, the value obtained by subtracting the ignition dead time equivalent crank angle IGNDEAD from the basic ignition timing MBTCAL is calculated as the previous combustion start timing MBTCYCL [degBTDC].
[0084]
The basic ignition timing MBTCAL calculated in this way is transferred to the ignition register, and an ignition signal for cutting off the primary current is output from the engine controller 31 to the ignition coil 13 at a timing when the actual crank angle coincides with the basic ignition timing MBTCAL. The
[0085]
If the basic ignition timing MBTCAL calculated in step 43 is used as the ignition timing command value for the current cycle, the previous combustion start timing MBTCYCL calculated in step 44 is displayed until the ignition timing for the next cycle is reached. 10 steps 22 are used.
[0086]
As described above, in the present embodiment, the basic ignition timing MBTCAL, which is the ignition timing at which MBT is obtained, is calculated without performing mass calculation such as the amount of unburned gas in the combustion chamber 5, so the calculation load is kept small. be able to.
[0087]
Further, as shown in the above equation (19), the initial combustion period BURN1 is expressed as a function of the combustion chamber volume V0 at the combustion start timing, the reaction probability RPROBA indicating the ease of combustion of the air-fuel mixture, and the combustion speed FLAME1. ing. Here, the larger the combustion chamber volume V0 at the combustion start timing, the smaller the reaction probability RPROBA, and the slower the combustion speed FLAME1, the longer the initial combustion period BURN1 and, as a result, the basic ignition timing MBTCAL advances.
[0088]
Similarly, as shown in the above equation (22), the main combustion period BURN2 is a function of the combustion chamber volume V2 at the start timing of the main combustion period, the reaction probability RPROBA indicating the ease of air-fuel mixture, and the combustion rate FLAME2. It is represented by Here, the larger the combustion chamber volume V2 at the start timing of the main combustion period, the smaller the reaction probability RPROBA, and the slower the combustion speed FLAME2, the longer the main combustion period BURN2, and as a result, the ignition timing MBTCAL is advanced.
[0089]
Thus, the combustion periods BURN1 and BURN2 can be accurately calculated by calculating the combustion periods BURN1 and BURN2 as functions of various parameters that affect the combustion period. As a result, the basic ignition timing MBTCAL calculated based on the combustion periods BURN1 and BURN2 can also be calculated with high accuracy. In addition, since the combustion period BURN is calculated by dividing it into an initial combustion period BURN1 corresponding to a flame nucleus growth period in which temperature and pressure are greatly affected and a main combustion period in which the influence of temperature and pressure is small, the combustion period BURN1 is calculated. The calculation accuracy of BURN is improved. The calculation accuracy can be further improved by further dividing the combustion period BURN into three or more.
[0090]
In the embodiment, the combustion speed FLAME1 used for the calculation of the initial combustion period BURN1 is the product of the laminar combustion speed SL1 and the turbulence intensity U1, and the combustion speed FLAME2 used for the calculation of the main combustion period BURN2 is turbulent with the laminar combustion speed SL2. Although calculated as the product of the strength U2, each may be obtained by a calculation method by addition as described in JP-A-10-30535.
[0091]
In the embodiment, the initial combustion period is defined as a combustion mass ratio from zero to 2% (that is, BR1 = 2%), and the main combustion period is defined as a combustion mass ratio from 2 to 60% (that is, BR2 = 58%). The present invention is not necessarily limited to this value.
[0092]
Next, a method for calculating the turbulence intensity, which is a feature of the present invention, will be described with reference to FIGS.
[0093]
A method of estimating the turbulence intensity during the combustion period when the intake valve is opened (IVO) will be described with reference to FIG. This method is a method of calculating turbulence intensity from an intake manifold pressure (hereinafter referred to as intake manifold pressure), valve timing, and valve lift amount, and is calculated at predetermined intervals.
[0094]
First, the intake air amount Qo (mm) from the intake manifold pressure Po (Pa), the valve opening period θo (deg) of the intake valve, and the engine rotational speed N (rpm). ^Three ) Is calculated by the following equation.
[0095]
[Expression 24]

However, f1 (θo, N) in the above equation may be calculated by storing the result calculated in advance through experiments or the like as a map as shown in FIG. 16 and referring to the map.
[0096]
Next, the average intake valve opening area So (mm) is calculated from the valve lift Lo (mm) during the intake valve opening period θo. ² ) Is calculated by the following equation.
[0097]
[Expression 25]

However, f2 (Lo) in the above equation may be calculated by storing the result calculated in advance through experiments or the like as a map as shown in FIG. 17 and referring to the map. Further, as shown in FIG. 18, the valve lift amount Lo may be obtained by integrating the lift amount during the intake valve opening period θo, obtaining an average value from the integrated value of the lift amount, and setting this average value as the lift amount Lo.
[0098]
The intake valve passing average speed (intake air flow speed) Vo (m / sec) is calculated from the calculated intake air amount Qo into the cylinder, average intake valve opening area So, and engine speed N by the following equation.
[0099]
[Equation 26]

And the turbulence intensity of the combustion period estimated in the intake valve open state is Uo
[0100]
[Expression 27]

Is calculated by
[0101]
Where, Kc: swirl control valve correction coefficient
Kt: Tumble control valve correction factor
Kq: Squish correction coefficient
The value measured in advance is used. Further, when the configuration does not include any of the swirl control valve, the tumble control valve, and the squish, the correction coefficient of the configuration without the calculation is set to 1.
[0102]
However, f3 (Vo, N) in the above equation may be calculated with reference to a map stored in advance as a map as shown in FIG.
[0103]
In this way, the average turbulence intensity during the combustion period obtained with the intake valve opened is Uo is the intake valve passage average calculated from the intake air amount Qo into the cylinder, the average intake valve opening area So, and the engine speed N. Since it is set based on the speed Vo, even if the intake air amount Qo into the cylinder, the average intake valve opening area So, and the engine speed N change in the transient operation state of the engine, the turbulence intensity Uo is changed according to the change. Can be calculated. Therefore, even when the engine is in a transient state, the turbulence intensity can be set according to the change in the engine operating state, and the ignition timing can be accurately calculated.
[0104]
A method of estimating the turbulence intensity during the combustion period in the state when the intake valve is closed (IVC) will be described based on FIG. This method calculates the turbulence intensity from the intake manifold pressure (hereinafter referred to as intake manifold pressure), valve timing, and valve lift amount when the intake valve is closed, similar to the calculation method when the intake valve is open. It is calculated at the timing of
[0105]
First, the intake air amount Qc (mm) from the intake manifold pressure Pc (Pa), the valve opening period θc (deg) of the intake valve, and the engine rotational speed N (rpm). ^Three ) Is calculated by the following equation.
[0106]
[Expression 28]

However, f1 (θc, N) in the above equation may be calculated with reference to a map stored in advance as a map as shown in FIG.
[0107]
Next, the average intake valve opening area Sc (mm) is calculated from the valve lift amount Lc (mm) during the intake valve opening period θc. ² ) Is calculated by the following equation.
[0108]
[Expression 29]

However, f2 (Lc) in the above equation may be calculated by storing the result calculated in advance through experiments or the like as a map as shown in FIG. 17 and referring to the map. Further, as shown in FIG. 18, the valve lift amount Lc may be obtained by integrating the lift amount during the intake valve opening period θc, obtaining an average value from the integrated value of the lift amount, and using this average value as the lift amount Lc.
[0109]
The intake valve passing average speed (intake air flow speed) Vc (m / sec) is calculated from the calculated intake air amount Qc into the cylinder, average intake valve opening area Sc, and engine speed N by the following equation.
[0110]
[30]

The average turbulence intensity during the combustion period when the intake valve is closed is Uc
[0111]
[31]

Is calculated by
[0112]
Where, Kc: swirl control valve correction coefficient
Kt: Tumble control valve correction factor
Kq: Squish correction coefficient
The value measured in advance is used. Further, when the configuration does not include any of the swirl control valve, the tumble control valve, and the squish, the correction coefficient of the configuration without the calculation is set to 1.
[0113]
However, f3 (Vc, N) in the above equation may be calculated with reference to a map stored in advance as a map as shown in FIG.
[0114]
Thus, the turbulence intensity Uc during the combustion period when the intake valve is closed is set based on the intake valve passing average speed Vc calculated from the intake air amount Qc into the cylinder, the average intake valve opening area Sc, and the engine speed N. Therefore, even if the intake air amount Qc, the average intake valve opening area Sc, and the engine speed N change in the transient operation state of the engine, the turbulence intensity Uc can be set according to the change. . Even when the engine is in a transient state, the turbulence intensity can be set according to the change in the operating state of the engine, and the ignition timing can be accurately calculated.
[0115]
Next, a method for calculating the turbulence intensity in the initial combustion period using the calculated average turbulence intensity Uo and Uc will be described with reference to FIG. This calculation is performed at regular intervals.
[0116]
First, the turbulence intensity Uo1 in the initial combustion period is calculated from the average turbulence intensity Uo by the following equation.
[0117]
[Expression 32]

However, f4 (Uo) in the above equation may be calculated with reference to a map stored in advance as a map as shown in FIG.
[0118]
Next, the turbulence intensity Uc1 in the initial combustion period is calculated from the average turbulence intensity Uc by the following equation.
[0119]
[Expression 33]

However, f4 (Uc) in the above equation may be calculated with reference to a map stored in advance as a map as shown in FIG.
[0120]
Further, the turbulence intensity U1 in the initial combustion period is calculated from the following equation from the turbulence intensity Uo1 and Uc1 in the initial combustion period based on the calculated turbulence intensity estimated in the intake valve open state or the closed state.
[0121]
[Expression 34]

Here, the coefficient k1 is calculated by the following equation.
[0122]
[Expression 35]

However, k1 = 0.5 may be set when the calculation described later is impossible due to the performance of the computer that performs the calculation.
[0123]
The coefficient kp is calculated by the following equation.
[0124]
[Expression 36]

Here, To: intake valve opening start time, Tc: intake valve closing time, Pave: average intake manifold pressure (Pa).
[0125]
However, the average intake manifold pressure can be calculated by Expression (37).
[0126]
[Expression 37]

Here, P (t): intake manifold pressure (Pa).
[0127]
Next, the coefficient kθ is calculated by the following equation.
[0128]
[Formula 38]

Here, To is the intake valve opening start time, Tc is the intake valve closing time, and θave is the average valve opening period (deg).
[0129]
However, the average valve opening period can be calculated by equation (39).
[0130]
[39]

Here, θ (t) is the valve timing.
[0131]
Next, the coefficient k _L Is calculated by the following equation.
[0132]
[Formula 40]

Here, To: intake valve opening start time, Tc: intake valve closing time, and Love: average lift amount (mm).
[0133]
However, the average lift amount can be calculated by equation (41).
[0134]
[Expression 41]

Here, L (t): valve lift amount (mm).
[0135]
Next, a method for calculating the turbulence intensity in the main combustion period using the calculated turbulence intensities Uo and Uc will be described with reference to FIG. This calculation is performed at regular intervals.
[0136]
First, the turbulence intensity Uo2 in the main combustion period is calculated from the turbulence intensity Uo by the following equation.
[0137]
[Expression 42]

However, f5 (Uo) in the above equation may be calculated with reference to a map stored in advance as a map as shown in FIG.
[0138]
Next, the turbulence intensity Uc2 in the main combustion period is calculated from the turbulence intensity Uc by the following equation.
[0139]
[Equation 43]

However, f5 (Uc) in the above equation may be calculated with reference to a map stored in advance as a map as shown in FIG.
[0140]
Further, the turbulence intensity U2 in the main combustion period is calculated from the following equation from the turbulence intensity Uo1 and Uc1 in the initial combustion period based on the calculated turbulence intensity estimated in the intake valve open state or the closed state.
[0141]
(44)

Here, the coefficient k1 is calculated in the same manner as described above.
[0142]
In this way, the turbulence intensity U1 in the initial combustion period and the turbulence intensity U2 in the main combustion period are calculated, and based on these turbulence intensity, the combustion speeds FLAME1 and FLAME2 are calculated by the equation (12) or the equation (16). Further, the combustion period BURN and the ignition timing MBTCAL are calculated.
[0143]
Therefore, in the present invention, when calculating the turbulence intensity, the turbulence according to the transient characteristics of the internal combustion engine is used by using the intake air flow velocity that passes through the intake valve corresponding to the change in the operating state when the engine is in the transient state. Since the strength is changed, it is possible to accurately calculate the combustion period and further the ignition timing, and to suppress the difference from the ignition timing at which MBT is obtained.
[0144]
FIG. 24 is a timing chart showing changes in intake manifold pressure, in-cylinder intake gas amount, valve passage flow velocity, turbulence intensity, and the like.
[0145]
When the # 2 cylinder is described as a representative, when the intake valve of the # 2 cylinder is open, the intake manifold pressure is low, that is, the intake gas amount is small, and the intake gas amount estimated in the intake valve open state is small (one point in the figure) The intake manifold pressure increases when the intake valve is closed, and the intake gas amount estimated when the intake valve is closed increases (indicated by the solid line in the figure). The intake valve passage flow rate shows the same tendency as the intake gas amount, and the turbulence intensity is also the same.
[0146]
In addition, as the turbulence intensity increases, the combustion speed when the intake valve is closed increases, but the turbulent combustion speed for the MBT calculation is estimated when the intake valve is open and when the intake valve is closed. Since it is calculated | required from a combustion rate, it becomes a combustion rate as shown with a broken line in a figure.
[0147]
FIG. 25 shows an example of the cylinder internal pressure when the intake port pressure rises during the intake stroke in a cylinder of the engine.
[0148]
The curve (1) in the figure shows the change in the cylinder internal pressure when the engine state when the intake valve is open is maintained until the intake valve is closed and ignition is performed with MBT calculated based on this state.
[0149]
The curve (2) in the figure shows the cylinder internal pressure when the engine state when the intake valve is closed continues from the open state of the intake valve until the intake valve is closed, and ignition is performed with MBT calculated based on this state. Showing change.
[0150]
When ignition is performed with the MBTs (1) and (2), the cylinder internal pressure becomes maximum at the reference crank position θPmax (12 to 15 deg crank angle after top dead center).
[0151]
The curve of (4) shows the change in the cylinder internal pressure when ignition is performed with the MBT of (1). As the pressure in the intake port increases, the turbulence intensity in the actual engine state increases from the turbulence intensity calculated in the engine state with the intake valve open, so the combustion speed also increases. Therefore, when ignition is performed with MBT calculated with the intake valve opened, knocking may occur as shown by the curve (4), which is earlier than the actual required ignition timing.
[0152]
On the other hand, curve (3) shows the change in the cylinder internal pressure when MBT is calculated in consideration of the turbulence intensity between the intake valve open state and the closed state by the calculation method described so far. . The combustion speed and the maximum in-cylinder pressure are located between the curves (1) and (2), and the in-cylinder pressure becomes maximum at the reference crank position θPmax, but no knock occurs.
[0153]
FIG. 26 is a diagram showing an example of the combustion mass ratio when the intake port pressure increases during the intake process of a certain cylinder.
[0154]
The curve (1) in the figure shows the change in the combustion mass ratio when the engine state with the intake valve opened is maintained until the intake valve is closed and ignition is performed with MBT calculated based on this state. The curve (2) in the figure shows the combustion mass ratio when the engine state when the intake valve is closed continues from the state where the intake valve is open until the intake valve is closed, and ignition is performed with MBT calculated based on this state. Shows changes.
[0155]
When ignition is performed with MBT, the combustion mass ratio Rmax at the reference crank position θPmax is substantially constant (60%), and the curves (1) and (2) pass this point.
[0156]
As the intake port pressure increases, the actual engine turbulence intensity increases from the turbulence intensity estimated in the intake valve open state, so the combustion speed also increases (the inclination increases). Therefore, when ignition is performed with MBT based on the turbulence intensity estimated in the intake valve open state, the combustion mass ratio Rmax at the reference crank position θPmax becomes larger than a predetermined value as shown by the curve (4), and knocking may occur. Occurs.
[0157]
On the other hand, when ignition is performed with MBT in consideration of the turbulence intensity calculated in the present invention, the change in the combustion mass ratio becomes a curve of (3), and the combustion mass ratio Rmax at the reference crank position θPmax is a predetermined value Rmax. It becomes.
[0158]
The overall operation of the ignition timing control apparatus in the present embodiment described above will be described. The in-cylinder volume VIVC and the in-cylinder temperature TINI at the time when the intake valve is closed are calculated (steps S12 and S13). Then, the reaction probability RPROBA indicated by the ratio when the combustion state of the combustion gas is set to 1 when the combustion state under the predetermined condition is 1 is calculated (step S14).
[0159]
The combustion speed depends on the residual gas ratio EGRREM, the water temperature TWK, and the equivalence ratio in addition to the atmospheric temperature and the combustion speed in the cylinder. I can leave. Of course, it is also possible to have a correlation depending on these factors as a functional expression.
[0160]
Next, when ignition is performed with MBT, the reference crank angle θPmax, which is the peak value of the pressure of the combustion gas, is substantially constant, and tends to shift toward the advance side on the high engine speed side, and is calculated as in step S15.
[0161]
Next, the cylinder internal volume V ₀ The temperature increase rate from the in-cylinder temperature TINI when the intake valve is closed is calculated from the effective compression ratio Ec, and the ambient temperature T0 when the combustion gas is ignited is calculated (steps S21 to S25). Based on the calculated atmospheric temperature T0, a laminar combustion speed SL1 is calculated (step S26).
[0162]
Since the combustion speed tends to increase as the laminar combustion speed increases, the turbulent flow speed in the cylinder also greatly depends on the combustion speed. The turbulent velocity depends on the shape of the intake port and the like, and generally increases as the engine speed increases. Therefore, this is calculated as the turbulence intensity U1 (step S27), and is multiplied by the laminar combustion speed SL1 to calculate the combustion speed FLAME1 (step S28).
[0163]
Then, a combustion period BURN1 from ignition to formation of a flame kernel is calculated (step S29). Since it is known that the formation of flame nuclei is performed by combustion of 2% of the mass of the combustion gas, 2% of the mass of the combustion gas is calculated and used. In the flame nucleation stage, the larger the cylinder volume at the time of ignition, the longer the formation time, so this is the numerator. By calculating using an arithmetic expression with the denominator as the denominator, it is possible to accurately and easily determine the combustion time at the flame nucleus formation stage.
[0164]
After the flame nuclei are formed, the flame nuclei rapidly grow in a rapid combustion stage in which the combustion gas is rapidly burned. A period during which 60% of the combustion gas at which the pressure reaches a peak is burned is calculated (steps S31 to S34). Again, the larger the volume in the cylinder at the time of growth of the flame kernel, the longer the formation time, so this is the numerator. The larger the reaction probability RPROBA, the shorter the time, and the faster the combustion speed, the shorter the time. Thus, the combustion time at the flame nucleus growth stage can be obtained accurately and easily.
[0165]
By adding the initial combustion period of the flame nucleus formation stage calculated as described above and the main combustion period of the flame nucleus growth stage, it is possible to calculate a period necessary for combustion with a combustion gas mass of 60% (step S42). Then, the MBT ignition timing is calculated in consideration of the dead time for ignition and the reference crank angle θPmax that is the maximum pressure in the cylinder, and an output command can be issued to the ignition coil (step S43).
[0166]
According to the present embodiment, the combustion gas amount calculating means for calculating the amount of combustion gas combusted in the cylinder up to a predetermined crank angle, and the laminar combustion speed that is the combustion speed in the laminar flow state of the combustion gas are calculated. Laminar combustion speed calculating means, turbulence intensity calculating means for calculating turbulence intensity which is an index indicating combustion turbulence that changes according to the engine transient state, and the laminar combustion speed and turbulence intensity in the cylinder Since there is provided an ignition timing calculating means for calculating a combustion speed, calculating a combustion period from the combustion speed and the amount of combustion gas in the cylinder, and calculating a basic ignition timing for obtaining MBT based on the combustion period. Even when the operating state of the internal combustion engine is in a transient state, the turbulence intensity is set according to the operating state, so that the ignition timing for obtaining MBT can be accurately calculated even in the transient state.
[0167]
Also, when calculating the combustion period, it is calculated by dividing into the flame nucleation stage where the sensitivity of the combustion period with respect to pressure and temperature is large, and the main combustion period where the sensitivity of the combustion period with respect to pressure and temperature becomes dull. The period can be calculated more accurately. In this case, calculation formulas such as combustion gas mass and cylinder volume for numerator, reaction area and combustion rate for denominator are used, and units of the same dimension are not used for numerator and denominator. Can be calculated.
[0168]
The combustion state is replaced with a dimensionless reaction probability and used for calculation. Since the reaction probability can be easily obtained by experiments or the like, an accurate combustion period can be calculated. Further, in the present embodiment, an example is shown in which the present invention is applied to an internal combustion engine having a variable valve mechanism that can continuously vary the operating angle and lift amount of the intake valve, but the cylinder contents when the intake valve is closed are shown. Since the MBT can be easily calculated even if the product changes, for example, an electromagnet is used to open and close the valve, and the valve timing can be arbitrarily changed. Application is also possible. Examples of a variable valve mechanism that can continuously vary the operating angle and lift amount of the intake valve are disclosed in Japanese Patent Application Laid-Open Nos. 11-107725 and 2003-41976.
[0169]
In this embodiment, when determining the combustion period from the combustion mass ratio, the combustion period BURN is divided into an initial combustion period BURN1 (up to a combustion mass ratio of 2%) and a combustion main period BURN2 (from a combustion mass ratio of 2% to 60%). The ignition timing MBTCYCL was calculated and calculated, but it may be further subdivided, or after subdividing, the combustion period (combustion mass ratio) is divided into, for example, an initial combustion period BURN1 and a main combustion period BURN2. You may make it add. By subdividing and determining the combustion period, the calculation accuracy of the combustion period can be improved.
[0170]
The present invention is not limited to the above-described embodiment, and it is obvious that various modifications can be made within the scope of the technical idea of the present invention.
[Brief description of the drawings]
FIG. 1 is a system diagram of an internal combustion engine of the present invention.
FIG. 2 is a block diagram of ignition timing control executed by an engine controller.
FIG. 3 is a pressure change diagram of a combustion chamber.
FIG. 4 is a characteristic diagram illustrating a change in combustion mass ratio.
FIG. 5 is a flowchart for explaining calculation of a physical quantity.
FIG. 6 is a diagram for explaining the positional relationship between an engine crankshaft and a connecting rod.
FIG. 7 is a characteristic diagram of a water temperature correction coefficient.
FIG. 8 is a characteristic diagram of an equivalence ratio correction coefficient.
FIG. 9 is a characteristic diagram of a reference crank angle.
FIG. 10 is a flowchart for explaining calculation of an initial combustion period.
FIG. 11 is a characteristic diagram of a rate of temperature increase.
FIG. 12 is a flowchart for explaining calculation of a main combustion period.
FIG. 13 is a flowchart for explaining calculation of basic ignition timing.
FIG. 14 is a diagram illustrating a method for calculating an average turbulence intensity estimated in an intake valve open state.
FIG. 15 is a diagram for explaining a method for calculating an average turbulence intensity estimated in the intake valve closed state;
FIG. 16 is an example of a map for calculating a variable f1.
FIG. 17 is an example of a map for calculating a variable f2.
FIG. 18 is a diagram for explaining a valve opening period and a lift amount of an intake valve.
FIG. 19 is an example of a map for calculating a variable f3.
FIG. 20 is a diagram illustrating a method for calculating the turbulence intensity during the initial combustion period.
FIG. 21 is an example of a map for calculating a variable f4.
FIG. 22 is a diagram illustrating a method for calculating the turbulence intensity during the main combustion period.
FIG. 23 is an example of a map for calculating a variable f5.
FIG. 24 is a timing chart showing changes in turbulence intensity.
FIG. 25 is a diagram for explaining a change in cylinder internal pressure;
FIG. 26 is a diagram for explaining a change in a combustion ratio.
[Explanation of symbols]
1 Internal combustion engine
2 Intake collector
3 Injector
4 Intake manifold
5 Exhaust manifold
6 Muffler
7 Air cleaner
8 Electronically controlled throttle
9 Air intake duct
20 Control unit

Claims (14)

Combustion gas amount calculating means for calculating the amount of combustion gas combusted in the cylinder up to a predetermined crank angle;
Laminar combustion speed calculating means for calculating a laminar combustion speed that is a combustion speed in a laminar state of the combustion gas;
Turbulence intensity calculating means for calculating the turbulence intensity, which is an index indicating the combustion turbulence that changes according to the engine transient state;
The in-cylinder combustion speed is calculated from the laminar combustion speed and the turbulence intensity, the combustion period is calculated from the combustion speed and the amount of combustion gas in the cylinder, and the basic ignition timing at which MBT is obtained based on the combustion period Basic ignition timing calculation means for calculating
Laminar combustion speed calculating means for calculating a laminar combustion speed that is a combustion speed in a laminar state of the combustion gas;
Turbulence intensity calculating means for calculating the turbulence intensity, which is an index indicating the combustion turbulence that changes according to the engine transient state;
A combustion rate calculation means for calculating the combustion rate of the combustion gas in the cylinder from the calculated laminar flow rate and turbulence intensity;
Cylinder internal volume calculation means for calculating the cylinder internal volume at the start of combustion gas combustion;
Combustion gas amount calculating means for calculating the amount of combustion gas combusted in the cylinder up to a predetermined crank angle;
Reaction probability calculating means for calculating a reaction probability indicating the ease of combustion of the combustion gas in the cylinder with respect to the ease of combustion of the combustion gas under a predetermined operating condition;
A basic ignition timing calculating means for calculating a combustion period based on the combustion speed, the volume in the cylinder, the amount of combustion gas, and the reaction probability, and calculating a basic ignition timing for obtaining MBT based on the combustion period;
An ignition timing control device for an internal combustion engine, comprising: Reference crank angle calculating means for calculating a crank angle at which the combustion pressure of the combustion gas in the cylinder is maximized;
An ignition dead time equivalent crank angle calculating means for calculating a crank angle section from the timing at which an ignition command signal of an ignition plug for igniting the combustion gas in the cylinder is output until the ignition plug ignites,
2. The ignition timing control device for an internal combustion engine according to claim 1 , wherein the basic ignition timing from which the MBT is obtained is calculated from the combustion period, the calculated reference crank angle, and a crank angle equivalent to the ignition dead time. . 3. The turbulence intensity is calculated based on an intake air flow velocity that passes through an intake valve of the internal combustion engine and flows into the cylinder when the internal combustion engine is in a transient state . An ignition timing control device for an internal combustion engine. The intake air flow rate is calculated based on an intake air amount into the cylinder when the intake valve is opened and closed, an average opening area of the intake valve, and a rotational speed of the internal combustion engine. Ignition timing control device for internal combustion engine. The ignition timing control device for an internal combustion engine according to claim 3 or 4 , wherein the intake air flow rate is calculated based on an intake air flow rate when the intake valve is opened and an intake air flow rate when the intake valve is closed . The combustion period is calculated based on the calculated value obtained by dividing the value obtained by multiplying the change amount and the engine rotational speed of the combustion mass proportion and cylinder volume by a value obtained by multiplying the reaction area of the reaction probability and the flame kernel and burn rate The ignition timing control device according to claim 1 . The reaction probability is set to a value larger than 1 when the easiness of combustion at a predetermined residual gas rate is 1, and is set to a value smaller than 1 as it is smaller than the predetermined residual gas rate. The ignition timing control device for an internal combustion engine according to claim 1 , wherein the ignition timing control device is an internal combustion engine. The reaction probability is set to a value greater than 1 when the temperature is higher than the predetermined water temperature, and is set to a value smaller than 1 when the temperature is lower than the predetermined water temperature. The ignition timing control device for an internal combustion engine according to claim 1 or 6 . The reaction probability is set to a value smaller than 1 on both the rich side and the lean side of the predetermined equivalence ratio, with the ease of combustion at a predetermined equivalence ratio being 1. An ignition timing control device for an internal combustion engine as described. 10. The internal combustion engine according to claim 1, wherein the amount of combustion gas burned up to the predetermined crank angle is the amount of combustion gas burned until the combustion pressure after ignition of the combustion gas reaches a maximum . Ignition timing control device. The combustion gas amount burned up to the predetermined crank angle is divided into a plurality of amounts, a combustion period corresponding to each combustion gas amount is calculated, and all combustion periods are summed to obtain the combustion period. The ignition timing control device for an internal combustion engine according to claim 10 . The amount of combustion gas combusted up to the predetermined crank angle is combusted after the combustion gas is ignited for the formation of flame nuclei and after the formation of the flame nuclei until the combustion pressure after the combustion gas is ignited is maximized. The ignition timing control device for an internal combustion engine according to claim 11, wherein the ignition timing control device is divided into combustion gas amounts . The atmospheric temperature at the start of combustion of the combustion gas in the cylinder and the atmospheric pressure at the start of combustion of the combustion gas in the cylinder are calculated, and the combustion speed of the combustion gas in the cylinder is a predetermined temperature of the combustion gas, The ignition timing control device for an internal combustion engine according to claim 1 or 12, wherein the ignition timing control device calculates the basic laminar combustion speed at a predetermined pressure based on the atmospheric temperature and the atmospheric pressure . 14. The temperature of the combustion gas when the intake valve is closed is calculated, and the ambient temperature is calculated by multiplying the calculated temperature by the rate of temperature increase calculated based on the effective compression ratio. ignition timing control system for an internal combustion engine according to.

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