JP4175922B2 - Intake air amount estimation device for internal combustion engine - Google Patents

Description

【００５２】
【数２】

【００５３】
ここで、数１はスロットル弁４３上流側（大気側）の圧力が下流側（吸気管側）の圧力より高い場合の通常の吸気流れ（順流）方向のスロットル弁通過吸気流量を、数２はスロットル弁４３上流側の圧力が下流側の圧力より低く逆流が生じている場合のスロットル弁通過空気流量を、それぞれ表している。
【００５４】
数１、数２において、Ｃt（θｔ）は流量係数、Ａｔ（θｔ）はスロットル弁開口面積であり、Ｃt（θｔ）、Ａｔ（θｔ）はそれぞれスロットル弁開度θｔの関数である。Ｐａはスロットル弁上流側の圧力で本実施形態では大気圧を、Ｐｍはスロットル弁下流側の圧力で本実施形態では吸気管圧力を表す。また、Ｔａ、Ｔｍは、それぞれスロットル弁上流側と下流側の吸気温度であり、本実施形態ではそれぞれ大気温度と吸気管内温度、Ｒはガス定数、κは吸気の比熱比を表している。なお、本実施形態ではＲとκは一定値として扱う。
【００５５】
本実施形態では、流量係数Ｃt（θｔ）及びスロットル弁開口面積Ａｔ（θｔ）は予めスロットル弁開度θ毎に算出され、数値マップの形でＥＣＵ８０のＲＯＭ８２に格納されている。
数１、数２に示すように、本実施形態のスロットル弁モデルは、大気圧Ｐａ及び吸気管圧力Ｐｍ、大気温度Ｔａまたは吸気管内温度Ｔｍ、及びスロットル弁開度θtを入力として、スロットル弁通過吸気流量ｍｔを出力とするモデル計算式である。
【００５６】
（２）吸気管モデル
図３は吸気管モデルを示す。吸気管モデルは、一定容積の空間内に流入する吸気流量と流出する吸気流量とに基づいて、質量保存則とエネルギー保存則とを用いて吸気管圧力Ｐｍ及び温度Ｔｍを算出する。
【００５７】
スロットル弁４３から吸気弁３２に至るサージタンク４４を含む吸気管４１の容積をＶｍとすると、図３に示すように容積部Ｖｍに流入する吸気流はスロットル弁通過吸気流であるため、流入吸気流量はｍｔ、吸気温度はＴａとなる。
【００５８】
また、容積部Ｖｍから流出する吸気流量は吸気弁を通過して気筒に充填される吸気流量ｍｃで、この吸気流の温度は吸気管温度Ｔｍとなる。
【００５９】
従って、質量保存則から下記数３が、エネルギー保存則から下記数４が、それぞれ導かれる。
【００６０】
【数３】

【００６１】
【数４】

【００６２】
本実施形態では、ＥＣＵ８０での演算処理を行うため、数３、数４を積分して離散化した、以下の数５、数６式を用いる。
【００６３】
【数５】

【００６４】
【数６】

【００６５】
ここで、添字ｋは今回計算実行時の値を、ｋ−１は前回計算実行時の値を表し、Δｔは離散化する際の差分時間（例えば８ｍｓ程度のメッシュ）を表している。また、数５、数６式におけるｍｃＡＶＥは：後述する筒内吸気流量ｍｃの吸気行程１サイクル分にわたる時間平均値である。
【００６６】
吸気管温度Ｔｍは、数５より求めたＰｍ／Ｔｍと、数６から求めたＰｍとを用いてＴｍ＝Ｐｍ／（Ｐｍ／Ｔｍ）として算出される。
（３）吸気弁モデル
吸気弁モデルは、図４に示すように吸気弁３２を通過する吸気流を表すモデルであり、基本的にはスロットル弁モデルと同様に絞り（吸気弁開口面積）を通過する吸気流として表される。
【００６７】
従って、吸気弁モデルは前述のスロットル弁モデル（数１、数２）と同様に下記の数７、数８の式で表される。
【００６８】
【数７】

【００６９】
【数８】

【００７０】
ここで、ｍｃは吸気弁を通過する吸気流量、すなわち気筒内に単位時間内に充填される吸気量、Ｐｍ、Ｔｍは吸気管圧力及び温度、Ｐｃ、Ｔｃはシリンダ内圧力及び温度である。また、Ｃｉｖ（Ｌ）は流量係数であり、吸気弁のリフト量Ｌの関数、Ａｉｖ（Ｌ）は吸気弁３２の開口面積であり、同様にリフト量Ｌの関数となる。
【００７１】
本実施形態では、積値Ｃiｖ（Ｌ）・Ａｉｖ（Ｌ）の値は各Ｌの値に対して計算され、数値テーブルの形でＥＣＵ８０のＲＯＭ８２に格納されている。
スロットル弁モデル（数１、数２）と同様に数７は吸気管圧力Ｐｍがシリンダ内圧力Ｐｃより大きく、気筒内に吸気が流入している場合を、数８はＰｍがＰｃより小さく気筒内から吸気が吸気管に逆流している場合を、それぞれ表している。
【００７２】
数７、数８に示すように、本実施形態の吸気弁モデルは、吸気管圧力Ｐｍ、シリンダ内圧力Ｐｃ、吸気管内温度Ｔｍまたはシリンダ内温度Ｔｃ、及び吸気弁リフト量Ｌを入力として、吸気弁通過吸気流量ｍｃを出力する。
前述の吸気管モデル計算式数５、数６で使用される平均吸気流量ｍｃＡＶＥは、数７、数８で算出される吸気弁通過空気流量ｍｃ吸気行程全体にわたって積分した値Ｍｃを、をクランク角１８０°に相当する時間Ｔ（１８０ＣＡ）で除した値である。
【００７３】
すなわち、ｍｃＡＶＥ＝Ｍｃ／Ｔ（１８０ＣＡ）
（４）シリンダモデル
本実施形態のシリンダモデルは、図５に示すように基本的には体積Ｖｃ（可変）の閉空間（Ｖｃはクランク角の関数）に吸気弁から単位時間当たり流量ｍｃ、温度Ｔｍの吸気が流入するモデルとして表されるが、吸気弁開弁中に排気ポートから排気弁を通過してシリンダ内に逆流する排気の流れを考慮して、吸気弁からの吸気流以外に流量ｍｅ、温度Ｔｅの排気が流入するモデルとして与えられる。
【００７４】
すなわち、本実施形態のシリンダモデルでは、ｍｃ、ｍｅ、Ｔｍ、Ｔｅ（及びクランク角）を入力とし、シリンダ内圧力Ｐｃ、温度Ｔｃ及び吸気行程中に筒内に充填された吸気の総量Ｍｃを出力する。
【００７５】
シリンダモデルは、基本的にはエネルギ保存則に基づく下記数９で表される。
【００７６】
【数９】

【００７７】
ここで、Ｑｗはシリンダ壁面、吸気ポート壁面などから吸気に与えられる熱量である。
【００７８】
また、シリンダ内温度Ｔｃは、気体の状態方程式から下記数１０で表される。
【００７９】
【数１０】

【００８０】
ここで、Ｍｃ１は吸気弁開弁時からの充填吸気量であり、吸気弁開弁時から現在（Ｔｃ計算時）までの吸気弁と排気弁からの流入気体流量（ｍｃ＋ｍｅ）の積分値として与えられる。
【００８１】
上記数９は、このままの形ではＥＣＵ８０で演算を行うことができないため、離散化形式に直し、以下の数１１の形で使用する。
【００８２】
【数１１】

【００８３】
なお、数１１におけるΔｔは数９を離散化する際の差分時間（シリンダモデルでは吸気管モデルの場合に較べてΔｔは短く設定され、例えば１ｍｓ程度のメッシュに設定される）である。
【００８４】
また、排気弁通過流量ｍｅは、次に説明する排気弁モデルにより求められる。
（５）排気弁モデル（図７、Ｍ５０）
排気弁モデルは吸気弁モデルと同様に、排気ポートの圧力と温度、Ｐｅ、Ｔｅ及びシリンダ圧力と温度Ｐｃ、Ｔｃとクランク角を入力として排気弁通過流量ｍｅを出力するモデルである。
【００８５】
図６は、排気弁モデルを示す。図６に示すように排気弁モデルは基本的には吸気弁モデルと同一であり、排気弁通過流量は数１２または数１３で与えられる。
【００８６】
【数１２】

【００８７】
【数１３】

数１２は、排気ポートからシリンダ内に排気が逆流する場合（すなわち、シリンダ内圧力Ｐｃが排気ポート圧力Ｐｅより低い場合）、数１３はシリンダから排気ポートに排気が流れる場合（ＰｃがＰｅより高い場合）をそれぞれ示している。なお、、Ｃｅｖ（Ｌ）、Ａｅｖ（Ｌ）はそれぞれ排気弁３５の流量係数と開口面積であり、ともに排気弁のリフト量Ｌの関数となる。
【００８８】
なお、本実施形態では、排気ポート圧力Ｐｅと排気温度Ｔｅは排気ポートに設けた排気圧センサと排気温度センサ（図示せず）により検出されるが、排気系の圧損は比較的小さいため、排気圧センサで排気圧を検出する代わりにＰｅを近似的に大気圧に等しいと仮定して演算を行うことも可能である。
また、同様に排気温度Ｔｅはシリンダ内気体温度と略等しいため、排気温度Ｔｅについても排気温度センサを設けることなくシリンダ内温度Ｔｃで代用するようにしても良い。
【００８９】
図７は、上記に説明したスロットル弁モデル、吸気管モデル、吸気弁モデル、シリンダモデルと排気弁モデルから構成されるエアモデル全体を示すブロック線図である。図６を用いて、各モデル要素のモデル計算式、入力と出力、各モデル要素間の関係及び筒内充填吸気量Ｍｃの算出過程等について再度まとめておく。
【００９０】
（Ａ）スロットル弁モデル（図７、Ｍ１０）
（１）計算式
【００９１】
【数１４】

【００９２】
【数１５】

【００９３】
（２）入力
・スロットル弁開度θｔ
・大気圧Ｐａ（大気圧センサ６３出力）
・大気温度Ｔａ（吸気温度センサ６２出力）
・吸気管圧力Ｐｍ
・吸気管温度Ｔｍ（数１５（逆流時）の場合）
（３）出力
・スロットル弁通過吸気流量ｍｔ
（Ｂ）吸気管モデル（図７、Ｍ２０）
（１）計算式
【００９４】
【数１６】

【００９５】
【数１７】

【００９６】
【数１８】

【００９７】
（２）入力
・スロットル弁通過吸気流量ｍｔ(k-1)
・吸気弁通過吸気流量（平均値）ｍｃＡＶＥ(k-1)
・Ｐｍ(k-1)、Ｔｍ(k-1)、（Ｐｍ／Ｔｍ）(k-1)
（３）出力
・吸気管圧力Ｐｍ(k)
・吸気管温度Ｔｍ(k)
（Ｃ）吸気弁モデル（図７、Ｍ３０）
（１）計算式
【００９８】
【数１９】

【００９９】
【数２０】

【０１００】
（２）入力
・吸気管圧力Ｐｍ
・吸気管温度Ｔｍ
・シリンダ圧力Ｐｃ
・シリンダ内温度Ｔｃ（数２０（逆流時）の場合）
・バルブリフト量Ｌ（クランク角から算出）
（３）出力
吸気弁通過吸気流量（瞬時値）ｍｃ
・吸気弁通過吸気流量ｍｃＡＶＥ（平均値）
（Ｄ）シリンダモデル（図７、Ｍ４０）
（１）計算式
【０１０１】
【数２１】

【０１０２】
【数２２】

【０１０３】
（２）入力
・シリンダ容積Ｖｃ（クランク角から算出）
・吸気弁通過吸気量（瞬時値）ｍｃ
・吸気管内温度Ｔｍ
・排気弁通過吸気量ｍｅ
・排気温度Ｔｅ
（３）出力
・シリンダ圧力Ｐｃ
・シリンダ内温度Ｔｃ
・筒内充填吸気量Ｍｃ
（Ｅ）排気弁モデル（図７、Ｍ５０）
（１）計算式
【０１０４】
【数２３】

【０１０５】
【数２４】

【０１０６】
（２）入力
・排気管圧力Ｐｅ
・排気管温度Ｔｅ
・シリンダ圧力Ｐｃ
・シリンダ内温度Ｔｃ（数２４の場合）
・バルブリフト量Ｌ（クランク角から算出）
（３）出力
排気弁通過吸気流量（瞬時値）ｍｅ
上述のように、スロットル弁モデルと排気弁モデル以外のモデル、すなわち吸気管モデル、吸気弁モデル、シリンダモデルの入力はクランク角以外は全て他のモデルの出力値が使用できる。また、クランク角は機関回転数ＮＥの関数として算出できる。
【０１０７】
更に、スロットル弁モデルの入力も、スロットル弁開度θ（ｔ）、大気圧Ｐａ、大気温度Ｔａを除き、他のモデルの出力値が使用できる。そして、大気圧Ｐａ、大気温度Ｔａは短時間では変化しないため、略一定値と考えて良い。
従って、図７に示したエアモデルでは基本的にスロットル弁開度θ（ｔ）と機関回転数ＮＥ、及び排気圧力Ｐｅ、温度Ｔｅが定まれば筒内充填吸気量ｍｃを算出することができる。
【０１０８】
具体的には、例えば機関始動時には初期値としてＰｍ、Ｐｃを大気圧Ｐａに、Ｔｍ、Ｔｃを大気温度Ｔａに、また、流量ｍｔ、ｍｃは０にそれぞれ設定し、スロットル弁開度θ（ｔ）及びクランク角を用いて上述の数１４から２４の計算を行う（Ｐｅ、Ｔｅは実測値を使用する）。
そして、この結果得られたＰｍ、ＰｃＴｍ、Ｔｃ、ｍｔ、ｍｃ、ｍｅ等の値を用いて数１４から数２４の次回の計算を行う。そして、この計算を機関始動時から所定の間隔（例えば８ｍｓ程度）でその時点におけるθ（ｔ）、ＮＥの値と前回の計算で求めたＰｍ、ｔｍ、ｍｔ、ｍｃ、ｍｅ等の値を用いて繰り返す。これにより、Ｐｍ、ｔｍ、ｍｔ、ｍｃ等は計算開始後短時間で実際の値に収束し、実際の筒内充填吸気量に一致するようになるはずである。
【０１０９】
上記の従来のエアモデルでは、吸気弁モデル（図４）において、シリンダから吸気ポートへの逆流が生じる場合（数２０）をも考慮した吸気量推定が行われているため、例えばバルブオーバラップ量が大きくて多量の吹き返しが生じるような場合にも筒内充填吸気量は正確に推定されるはずである。
しかし、前述したように、実際には従来のエアモデルは吹き返し時の吸気ポートにおける局所的な圧力上昇を考慮していないために、吹き返し量が大きい場合には特に筒内充填吸気量の推定精度が低下する問題が生じる。
【０１１０】
本発明では、上記従来のエアモデルに対して、吸気管と吸気ポートとを分けてモデル化し、吸気ポートモデルを用いて吹き返し発生時の吸気ポート圧力を吸気管圧力とは別個の値として算出することにより、筒内充填吸気量の推定精度を向上させている。
図８は、本実施形態のエアモデルの全体構成を説明する図７と同様なブロック図である。
【０１１１】
図８に示すように、本実施形態では吸気管モデルＭ２０と吸気弁モデルＭ３０との間に、吸気ポートモデルＭ２５が設けられている点が相違する。
吸気ポートモデルＭ２５の詳細については後述するが、本実施形態ではこの吸気ポートモデルＭ２５を用いて吸気ポート圧力Ｐｐを算出し、吸気弁通過流量ｍｃを数２０の代わりに次に示す数２５、数２６を用いて算出するようにしている。ここで、数２５は吸気ポートへの吹き返しが生じていない場合（すなわち、Ｐｐ≧Ｐｃの場合）、数２６は吸気ポートに吹き返しが生じている場合（すなわちＰｐ＜Ｐｃの場合）である。
【０１１２】
【数２５】

【０１１３】
【数２６】

【０１１４】
ここで、後述するように本実施形態の吸気ポートモデルＭ２５ではシリンダからの吹き返しが生じている場合には、吸気管圧力Ｐｍとは別個に、Ｐｍより高い圧力として吸気ポート圧力Ｐｐを算出するが、吹き返しが生じていない場合には、吸気ポート圧力Ｐｐとして吸気管圧力Ｐｍと等しい値を使用し、吸気管圧力Ｐｐと別個に吸気ポート圧力Ｐｐを算出することは行わない。
【０１１５】
すなわち、図８のエアモデルは、シリンダからの吹き返しが生じていない場合には図７のモデルと同一になる。
これにより、吸気弁開弁開始時の吹き返し気流による吸気ポートの局所的な圧力上昇も吸気弁通過流量に正確に反映されるようになるため、バルブオーバラップが大きい運転状態においても正確に筒内充填吸気量が算出されるようになる。
【０１１６】
図９は、本実施形態の吸気ポートモデルＭ２５を用いた吸気弁通過流量算出操作の概略を示すフローチャートである。本操作は、ＥＣＵ８０により例えば１ｍｓ程度の比較的短い間隔で繰り返されるルーチンとして行われる。
図９の操作では、吸気ポートへの吹き返しが生じているか否かを判定し、吹き返しが生じている場合には、吸気ポートへの吹き返しによる吸気ポートの局所的な圧力上昇を加味した吸気ポート圧力を算出し、この吸気ポート圧力Ｐｐに基づいて吸気弁モデル計算式数２６を用いて吸気弁通過流量を算出する。また、吹き返しが生じていない場合には、Ｐｐ＝Ｐｍ、Ｔｐ＝Ｔｍとして、数２５を用いて吸気弁通過吸気流量を算出する。この場合には数２５は数２０と同じ式になる。
【０１１７】
すなわち、図９においてステップ９０１では前回操作実行時に算出した吸気ポート圧力Ｐｐを用いて吸気弁通過吸気量ｍｃを算出する。
そして、ステップ９０３では現在シリンダから吸気ポートへの吹き返しが生じているか否かを判定する。この判定は、前回算出した吸気ポート圧力ＰｐがＰｃより低くなっているか否かに基づいて判断しても良いが、単に圧力の比較のみで吹き返しの有無を判断したのでは吸気ポートの圧力脈動の影響による誤差が生じる場合がある。このため、本実施形態ではＰｐ≧Ｐｃの場合の吸気弁通過流量（数２５）を正、Ｐｐ＜Ｐｃの場合の吸気弁通過流量（数２６）を負の値として、数回分の吸気弁通過流量ｍｃを積算し、この積算値が負の値になったときにはシリンダから吸気ポートへの吹き返しが発生していると判断するようにしている。
【０１１８】
ステップ９０３で、吹き返しが発生している場合、すなわちステップ９０１で算出した吸気弁通過吸気量ｍｃの積算値が負の値であった場合には、ステップ９０７に進み、後述する吸気ポートモデル計算式に基づいて吸気ポート圧力Ｐｐを算出する。
また、ステップ９０３で吸気弁通過流量ｍｃの積算値が正であった場合、すなわち吹き返しが生じていない場合には、吸気ポート圧力Ｐｐとして前述の吸気管モデル計算式（数１６から数１８）で算出した吸気管圧力Ｐｍを設定して今回の操作を終了する。
【０１１９】
これにより、次回の操作実行時には吹き返しが生じている場合には吸気ポートモデル計算式に基づいて算出した吸気ポート圧力を用いて、また吹き返しが生じていない場合には吸気管圧力を用いて吸気弁通過流量が算出されるため、吹き返しの影響を考慮した正確な吸気弁通過流量が算出される。
従って、正負を考慮して吸気弁通過流量を吸気行程１サイクルにわたって積算することにより、筒内に充填される新気の量Ｍｃが正確に算出されるようになる。
【０１２０】
次に、上記ステップ９０７の吸気ポート圧力Ｐｐ算出のために使用する吸気ポートモデルについて説明する。
本実施形態で使用することができる吸気ポートモデルは１つに限られるわけではなく、いくつかの吸気ポートモデルが使用可能である。以下に本実施形態で使用可能な吸気ポートモデルの例について説明するが、図９の操作では以下の吸気ポートモデルのうちいずれかに基づいて吸気ポート圧力が算出される。
【０１２１】
（１）第１の実施形態
図１０は、本実施形態の吸気ポートモデルを模式的に示す図である。
本実施形態では、吹き返しが生じている場合には、吸気ポートを所定の仮想容積を有する閉空間と仮定し、この閉空間にシリンダからの吹き返し気体の全量が流入するとした場合の圧力上昇を求め、吸気管圧力にこの圧力上昇分を加えた圧力を吸気ポート圧力Ｐｐとして使用する。
吹き返しにより、吸気弁開弁直後の短時間にシリンダから高速の気流が吸気ポートに流入すると、吸気ポートには流入する気体により局所的に圧力の上昇が生じる。この局所的な圧力上昇は流入する気体が高速であるほど、また、流量が大きいほど大きくなると考えられる。
【０１２２】
そこで、本実施形態では吸気ポートが機関運転状態により定まる仮想の容積を有する閉空間（容器）であると仮定し、この閉空間に流入する吹き返し気流により生じる圧力上昇が吸気ポートの圧力上昇に等しいと仮定する。
今、吸気ポートの閉空間としての容積をＶｐ、単位時間当たりの吹き返し流量、すなわち吹き返し時の吸気弁通過流量をｍｃとすると、吹き返し気体の流入による吸気ポートＰｐの単位時間当たり圧力変化ｄＰｐ／ｄｔは、状態方程式より次の数２７で表される。
【０１２３】
【数２７】

【０１２４】
ここで、Ｒは吹き返しガスの気体定数、またＴｐは吸気ポート内の気体温度であり略シリンダ内の気体温度Ｔｃに等しいと考えられる。
従って、吹き返しによる吸気ポート圧力上昇幅ΔＰｐは吹き返し開始後の数２７の積分値に等しくなり、Ｔｃの変化を無視すると次の数２８で表される。
【０１２５】
【数２８】

【０１２６】
ここで、吸気弁通過流量ｍｃの積分値は数２６で算出される流量ｍｃ（吹き返し流量）を積算することにより求められ、吸気ポート圧力Ｐｐは、
Ｐｐ＝Ｐｍ＋ΔＰｐ
として算出される。
【０１２７】
図９ステップ９０７では、上記により吸気ポート圧力Ｐｐを算出する操作と、算出したＰｐの値を用いて次回の吸気弁通過流量ｍｃを算出する操作（図９、ステップ９０１）を逐次繰り替えすことにより、吸気弁通過吸気量ｍｃが正確に算出されるようになる。
なお、吸気ポートの仮想容積Ｖｐは予め実験結果に基づいて算出され、例えば機関回転数ＮＥと機関１の吸気バルブタイミングＶＴとを用いた数値マップとしてＥＣＵ８０のＲＯＭ８２に格納されており、図９の操作実行毎に実際の機関回転数とバルブタイミングとに基づいてＥＣＵ８０により読み出される。
【０１２８】
例えば機関の回転数が高くなるほど吹き返し気体の量は少なくなる。このため、仮想容積Ｖｐは回転数が増大するほど小さな値に設定される。また、バルブタイミングが進角するほど（バルブオーバラップが大きいほど）吹き返し気体の量は増大するが吸気ポートの圧力上昇は吹き返し気体の量に対応して増大しない。このため、このため、仮想容積Ｖｐは吸気弁のバルブタイミングが進角するほど大きな値に設定されている。
【０１２９】
（２）第２の実施形態
上記第１の実施形態では、吸気ポートの仮想容積Ｖｐは機関運転状態（例えば機関回転数とバルブタイミング）とに応じて定まる一定値とされていた。
しかし、実際の運転では吸気ポートは閉空間ではなくサージタンクに接続されており、吸気ポートに流入した吹き返し流量の全量が吸気ポートにとどまるわけではない。流入した気体は、吸気ポートや吸気管内に存在する気体を圧縮しながら圧力波となって吸気ポート内をサージタンクに向かって進行する。
【０１３０】
そこで、本実施形態では吸気ポートの仮想容積が内部の圧力に応じて膨張すると仮定する。また、図１１に示すように、吸気ポートの仮想容積の膨張の際には吸気ポート内の既存の気体との境界にばね定数ｋのスプリングＳと減衰定数ｃのダッシュポットＤで表される抵抗が生じ、この抵抗に抗して吹き返し気体が膨張すると仮定する。
【０１３１】
今、吸気ポートに流入する吹き返し気体の流量をｍｃ、温度をＴｃとすると、吸気ポートに流入する気体のエネルギー保存則は次の数２９で表される。
【０１３２】
【数２９】

【０１３３】
ここで、Ｗは吸気ポートに流入する吹き返し気体がする仕事であるが、この仕事を、図１１に示すようにスプリングＳとダッシュポットＤとで付勢された仮想の可動境界壁１１０を押動する仕事として表すと次の数３０を得る。
【０１３４】
【数３０】

【０１３５】
ここで、Ａはポート断面積（一定値）である。
数３０の右辺を展開し、ｋ／Ａ＝Ｋ、ｃ／Ａ＝Ｃと置くと数３１を得る。
【０１３６】
【数３１】

【０１３７】
従って、数３１を数２９に代入して数３２を得る。
【０１３８】
【数３２】

【０１３９】
また、状態方程式からＶｐ、Ｐｐ、ｍｃの間に次の数３３の関係が成立する。
【０１４０】
【数３３】

【０１４１】
更に、数３２を離散化して表すと次の数３４を得る。
【０１４２】
【数３４】

【０１４３】
吸気ポート圧力Ｐｐは、数２６で算出される吸気弁通過流量ｍｃ（吹き返し流量）と数３３とを用いて数３４の逐次計算を行うことにより算出される。数３４における定数Ｋ及びＣは予め実験により機関回転数ＮＥと吸気弁バルブタイミングＶＴとの関数として求められ、ＮＥとＶＴとを用いた数値マップとしてＥＣＵ８０のＲＯＭ８２に格納されており、数３４の演算毎に実際の機関回転数とバルブタイミングとに基づいてこのマップから読み出される。
これにより、更に正確に吹き返しによる吸気ポート圧力上昇を算出することが可能となり、筒内充填吸気量の算出精度が向上する。
【０１４４】
（３）第３の実施形態
上述の第２の実施形態では、図１１の仮想の可動境界壁１１０のする仕事と、吸気ポートに流入する気体のエネルギ保存則に基づいて吸気ポート内圧力Ｐｐを算出したが、本実施形態では計算を簡素化するために、図１１の可動境界壁１１０の両側に作用する力の釣り合いについて考える。
【０１４５】
図１１の可動壁１１０の左側（吸気管側）には吸気管圧力Ｐｍと、スプリングＳとダッシュポットＤとによる力Ｆが作用しており、可動壁１１０の右側（吸気ポート側）には吸気ポート圧力Ｐｐか作用しているのであるから、ΔＰｐ＝Ｐｐ−Ｐｍと置くとΔＰｐ×Ａ＝Ｆとなる。ここで、Ａは可動壁１１０の面積（吸気ポートの断面積）、ＦはスプリングＳとダッシュポットＤとによる付勢力の合計でありスプリングＳのばね定数ｋとダッシュポットＤの減衰定数ｃを用いて次の数３５で与えられる。
【０１４６】
【数３５】

【０１４７】
数３５と上記ΔＰｐ×Ａ＝Ｆとを用いて、Ｋ＝ｋ／Ａ²、Ｃ＝ｃ／Ａ²と置くと数３６を得る。
【０１４８】
【数３６】

【０１４９】
また、ＶｐとｄＶｐ／ｄｔとは、それぞれ数３７と数３８とで表される。
【０１５０】
【数３７】

【０１５１】
【数３８】

【０１５２】
ΔＰｐは吹き返し気体の流入による吸気ポートの局所的圧力上昇幅であるため、吸気ポート圧力Ｐｐは、となる。
従って、ΔＰｐはまず、吸気弁開弁時の初期値としてＰｐ＝Ｐｍ、Ｖｐ＝０とおいて、数２５で算出される吸気弁通過流量ｍｃ（吹き返し流量）を用いて数３８からを求め、これを数３６に代入してΔＰｐを求め、更にＰｐ＝Ｐｍ＋ΔＰｐからＰｐを求める。そして、求めたＰｐの値を用いて数２５でｍｃを求め、このｍｃ及びその積算値を用いて数３７、数３８とからＶｐとｄＶｐ／ｄｔとを算出し、数３６に代入することによりΔＰｐを求める操作を繰り返す逐次計算により算出することができる。
【０１５３】
（４）第４の実施形態
次に、吸気ポートモデルの第４の実施形態について説明する。
本実施形態では、吹き返しが発生した時の吸気ポート圧力Ｐｐは、吹き返し気流の吸気弁を含む仮想の管路通過時の圧損により定まると仮定する。
【０１５４】
ここで、管路通過時の気体流速をｕとすると圧損（Ｐｃ−Ｐｐ）はｕ²の関数となり、次の数３９で表される。
【０１５５】
【数３９】

【０１５６】
ここで、数３９右辺第一項のρは吹き返し気体の密度であり、シリンダ内圧力Ｐｃと温度Ｔｃとの関数となる。
【０１５７】
また、数３９右辺第二項は、数４０のように表される。
【０１５８】
【数４０】

【０１５９】
ここでαは定数、ｌ、ｄはそれぞれ吸気弁を含む仮想の流路の長さと直径とであり、吸気弁のリフトＬの関数となる。α、ｌ、ｄの値については、予め実際の機関を用いた実験により求めてあり、ＥＣＵ８０のＲＯＭ８２に格納してある。
【０１６０】
また、Ｐｃ、Ｔｃはシリンダモデル（図７、Ｍ４０）で算出した値を使用する。これにより、吸気ポート圧力Ｐｐは、次の数４１により算出される。なお、数３９から、吹き返しによる吸気ポート圧力上昇分ΔＰｐは、ΔＰｐ＝Ｐｐ−Ｐｍ＝Ｐｃ−Ｐｍ−（ρｕ²／２）−α（ｌ／ｄ）ｕ²と表すこともできる。
【０１６１】
【数４１】

【０１６２】
すなわち、本実施形態では、シリンダモデル（図７、Ｍ４０）、吸気弁モデル（図７、Ｍ３０）を用いて算出したｍｃ、Ｐｃ、Ｔｃを用いてρ、ｕを算出し、数４１を用いて吸気ポート圧力Ｐｐを算出し、次のｍｃ、Ｐｃ、Ｔｃを算出する操作を繰り返すことにより、吹き返しによる吸気ポート圧力の局所的上昇が吸気弁通過流量ｍｃに反映され、筒内充填吸気量の算出精度が向上する。
【０１６３】
（５）第５の実施形態
次に吸気ポートモデルの第５の実施形態について説明する。
図１２は、本実施形態の吸気ポートモデルを模式的に示す。図１２に示すように、本実施形態では吸気弁から吸気ポートに流入した吹き返し気体の一部が絞り１２０を通過して吸気管に流出すると仮定する。
【０１６４】
今、吸気管圧力Ｐｍ、吸気ポート圧力Ｐｐ、吸気ポート温度Ｔｐ、吸気ポート容積Ｖｐ、吸気弁通過流量（吹き返し流量）ｍｃ及び絞り１２０を通過して吸気ポートから吸気管に流入する気体流量ｍｐを用いて気体の状態方程式より数４２を得る。
【０１６５】
【数４２】

【０１６６】
また、絞り１２０を通過する流量ｍｐはＰｍ、Ｐｐ、Ｔｐ及び吹き返し気体の比熱比κとを用いて絞り通過流量の計算式数４３、数４４により算出される。
【０１６７】
【数４３】

【０１６８】
【数４４】

【０１６９】
ここで数４３のＣｐは絞り部の流量係数、Ａｐは仮想絞り１２０の開口面積であり、機関の運転状態（機関回転数ＮＥと吸気弁バルブタイミングＶＴ）に応じて定められる。
【０１７０】
ここで、吹き返し気体温度Ｔｐは略シリンダ内気体温度Ｔｃに等しいと仮定して数４２を離散形式に書きなおすと数４５を得る。
【０１７１】
【数４５】

【０１７２】
本実施形態では、吹き返し発生時にＰｐ＝Ｐｍ、ｍｐ＝０を初期値として、数２６を用いて吸気弁モデルから吸気弁通過流量ｍｃを算出し、このｍｃを用いて数４５から吸気ポート圧力Ｐｐを算出する。更に、次回からはこの算出したＰｐを用いて数２６からｍｃを、数４３からｍｐをそれぞれ算出し、これらを用いて数４５から吸気ポート圧力Ｐｐを算出する操作を繰り返す。これにより、吸気ポートへの吹き返しによる圧力の局所的上昇が吸気ポート圧力Ｐｐに正確に反映され、筒内充填吸気量の算出精度が向上する。
【０１７３】
（６）第６の実施形態
本実施形態では、吸気弁開弁後の吸気ポート圧力上昇を正弦関数で近似する。実際の運転においては、吸気弁開弁後の吸気ポート圧力上昇は当然ながら必ずしも正弦波にはなっていない。しかし、吹き返しが問題になるのは吸気弁開弁開始時後の極めて短い時間であるため、圧力上昇を正弦関数で近似しても大きな誤差は生じない。
【０１７４】
本実施形態では、吸気弁開弁後の圧力上昇ΔＰｐ＝Ｐｐ−Ｐｍをクランク角ＣＡの正弦関数として数４６で表す。
【０１７５】
【数４６】

【０１７６】
ここで、ＣＡioは吸気弁が開弁を開始するクランク角である。
【０１７７】
数４６における正弦関数の振幅ａ（最大圧力上昇幅）及び周期ｂ（圧力上昇期間）は、ともに吹き返し気体の流量に依存するため、バルブオーバラップ量ＯＬと吸気圧Ｐｍと排気圧Ｐｅとの比（Ｐｍ／Ｐｅ）との関数になり、数４７、数４８のように表すことができる。
【０１７８】
【数４７】

【０１７９】
【数４８】

【０１８０】
ここで、Ｃ₁からＣ₄は定数であり、実際の機関を用いた実験により予め定められる。
【０１８１】
図１３は、数４６から４８を用いて算出した吸気ポート圧力上昇幅ΔＰｐの一例を示す図である。図２においてカーブＶＴ４０、ＶＴ２０、ＶＴ０は、それぞれ吸気弁開弁クランク角（数４６のＣＡio）が上死点前（ＢＴＤＣ）４０度、２０度、０度の場合を示している。
なお、数４６の関数の振幅ａ及び周期ｂは数４７及び数４８のように数式を用いて表す代わりにＯＬと（Ｐｍ／Ｐｅ）とを引数とする数値マップの形で予めＥＣＵ８０のＲＯＭ８２に格納しておくようにしても良い。
【０１８２】
本実施形態では、吸気弁開弁後の吸気ポート圧力上昇幅ΔＰｐを正弦関数と仮定して数４６の形に表すことにより、簡易かつ正確に吹き返し発生時の吸気ポート圧力上昇を算出することが可能となり、筒内充填吸気量の算出精度を向上させることが可能となっている。
【０１８３】
【発明の効果】
各請求項に記載の発明によれば、吸気弁モデル計算式に基づいて吸気弁通過吸気流量を算出し、算出した吸気弁通過吸気流量を用いて筒内充填吸気量を算出する場合に、吸気ポートの圧力を考慮することにより正確にシリンダから吸気ポートへの吹き返し流量を算出することができ、筒内充填吸気量の算出精度を向上させることが可能となる共通の効果を奏する。
【図面の簡単な説明】
【図１】本発明を火花点火式機関に適用した場合の実施形態の概略構成を示す図である。
【図２】スロットル弁モデルを説明する図である。
【図３】吸気管モデルを説明する図である。
【図４】吸気弁モデルを説明する図である。
【図５】シリンダモデルを説明する図である。
【図６】排気弁モデルを説明する図である。
【図７】従来技術におけるエアモデルの全体を示すブロック図である。
【図８】吸気ポートモデルを持ちたいエアモデル全体を示すブロック図である。
【図９】図８のエアモデルにおける吸気弁通過流量算出操作を説明するフローチャートである。
【図１０】吸気ポートモデルの第１の実施形態を説明する図である。
【図１１】吸気ポートモデルの第２の実施形態を説明する図である。
【図１２】吸気ポートモデルの第５の実施形態を説明する図である。
【図１３】吸気ポートモデルの第６の実施形態を用いて算出した吸気ポート圧力上昇の変化の一例を示す図である。
【符号の説明】
１０…内燃機関全体
８０…電子制御ユニット（ＥＣＵ）
Ｍ１０…スロットル弁モデル
Ｍ２０…吸気管モデル
Ｍ２５…吸気ポートモデル
Ｍ３０…吸気弁モデル
Ｍ４０…シリンダモデル
Ｍ５０…排気弁モデル[0001]
BACKGROUND OF THE INVENTION
The present invention relates to an intake air amount estimation device for an internal combustion engine, and more particularly to an intake air amount estimation device that calculates an intake air amount of an internal combustion engine by a calculation formula that models an intake system of the internal combustion engine.
[0002]
[Prior art]
The intake system of an internal combustion engine is divided into elements such as throttle valves, intake pipes, intake valves, cylinders, and exhaust valves, and each element is modeled and expressed as a mathematical formula. Pressure, temperature, flow rate, etc. between each model There is known an intake air amount estimation device for an internal combustion engine using a so-called air model that obtains an intake air amount (in-cylinder charged intake amount) of an engine by calculation using the so-called air model.
[0003]
In an intake air amount estimation device using such an air model, it is usually possible to calculate the in-cylinder charged intake amount only by the engine speed and the throttle valve opening in addition to the atmospheric pressure and the atmospheric temperature. In the intake air amount estimation device using such an air model, the in-cylinder charged intake amount can be calculated by calculation using a model formula, so that the responsiveness is good even during transient operation where the change rate of the throttle valve opening is large. The intake air amount can be calculated.
[0004]
As an example of a control device for an internal combustion engine that performs this kind of intake air amount estimation, there is one disclosed in Patent Document 1, for example.
In the apparatus of Patent Document 1, the engine intake system is divided into each element of a throttle valve, an intake pipe including a surge tank, and an intake valve, and these elements are represented by a simulation model, and the pressure of the intake air flow in each model The temperature and flow rate are calculated using physical laws such as energy conservation law, mass conservation law, and momentum conservation law. The in-cylinder charged intake air amount is calculated based on the intake air flow rate that passes through the intake valve calculated as described above.
[0005]
In the device of Patent Document 1, it is assumed that the cylinder pressure is equal to the intake pipe pressure (surge tank pressure) due to the intake gas filled in the cylinder when the intake valve is closed, and the cylinder intake air amount (ie, intake air) The in-cylinder intake amount when the valve is closed) is obtained as a value proportional to the pressure of the intake pipe (surge tank) when the intake valve is closed.
However, in an actual engine, depending on the operating conditions of the engine, when the intake valve is opened, the gas in the cylinder may flow backward from the intake valve to the intake port due to the pressure in the cylinder. Since the device does not consider this blowback, there is a problem that the calculation accuracy of the in-cylinder charged intake air amount decreases.
[0006]
In order to solve this problem, the present applicant has already proposed an intake air amount estimation device for an internal combustion engine using an air model in consideration of the blow-back in Japanese Patent Application No. 2001-316350.
In the device disclosed in Japanese Patent Application No. 2001-316350, the engine intake system is modeled by dividing it into a throttle valve, an intake pipe including a surge tank, and intake valve elements, and in addition to these, cylinder and exhaust valve elements are modeled. The air model that models is used. In the device of this Japanese Patent Application No. 2001-316350, the throttle valve and the intake pipe use the same models as in Patent Document 1, but in addition to these models, exhaust valves, cylinders, and intake valves are modeled, By calculating the temperature, pressure, and flow rate of the intake air using the above physical laws, the in-cylinder charged intake air amount that takes into account the blowback when the intake valve is opened is calculated.
[0007]
That is, in the device of the above Japanese Patent Application No. 2001-316350, the pressure in the cylinder is calculated based on the relation between the models and the physical law such as the law of conservation of mass and the law of conservation of energy, and the intake pipe pressure and the pressure in the cylinder are calculated. Based on this, the intake flow rate passing through the intake valve is calculated. This flow rate is a positive flow rate from the intake pipe into the cylinder when the intake pipe pressure is higher than the pressure in the cylinder, but when the pressure in the cylinder is higher than the intake pipe pressure, that is, blowback occurs. In some cases it is negative. The in-cylinder charged intake air amount is obtained by integrating the intake valve passing intake air amount for one cycle of the intake valve opening period.
[0008]
[Patent Document 1]
JP 2001-41095 A
[0009]
[Problems to be solved by the invention]
However, according to the subsequent research, as in the device of the above Japanese Patent Application No. 2001-316350, the intake valve passage intake flow rate was calculated only based on the intake pipe pressure including the surge tank and the cylinder internal pressure. In particular, it has been found that the calculation accuracy of the in-cylinder charged intake air amount greatly decreases under operating conditions where the valve overlap is large, that is, under conditions where the amount of blowback is large.
[0010]
In the device of the above Japanese Patent Application No. 2001-316350, when blow-back occurs, the calculated intake pipe pressure rises according to the blow-back flow rate. However, since the intake pipe volume including the surge tank is larger than the blow-back flow rate, it is caused by blow-back. The rise of the intake pipe pressure is relatively small. For this reason, the difference between the cylinder internal pressure and the intake pipe pressure when the blowback occurs is a relatively large value, and the amount of gas flowing out of the cylinder due to the blowback is relatively large. When a large amount of blown back gas flows into the intake pipe in this way, fresh air flows into the cylinder even when the cylinder internal pressure decreases and the intake flow from the intake port to the cylinder resumes. Since the timing to perform is delayed, the amount of fresh air filled in the cylinder is reduced.
[0011]
However, in an actual engine, the amount of gas flowing backward due to the blowback when the intake valve is opened is smaller than the amount calculated by the air model of the above Japanese Patent Application No. 2001-316350, and the in-cylinder charged intake amount (new air amount) is the above model. It becomes larger than the value calculated by. As a result of the research, the difference in the blowback amount is considered to be caused by a sudden local pressure increase in the intake port portion immediately after the intake valve is opened in an actual engine.
[0012]
When blowback occurs, particularly when the valve overlap is large, immediately after the intake valve is opened, a high-speed blown airflow almost close to the sonic velocity flows into the intake port from the intake valve. In Japanese Patent Application No. 2001-316350, the intake valve passage intake flow rate (return flow rate) is calculated on the assumption that this high-speed airflow flows into the intake pipe (surge tank) and the pressure in the surge tank rises uniformly. For this reason, the pressure increase in the surge tank is relatively small, and the difference between the surge tank and the cylinder internal pressure remains relatively large, so the blowback flow rate is relatively large.
[0013]
However, since the flow velocity of the blowback airflow immediately after the intake valve is opened is actually high, an abrupt and local pressure increase occurs at the intake port into which this airflow flows, and the intake port pressure becomes considerably higher than the surge tank pressure. . For this reason, immediately after the intake valve is opened, the pressure difference between the intake port and the cylinder is reduced due to an increase in the intake port pressure, thereby suppressing the blowback from the cylinder to the intake port.
[0014]
In other words, the blow-back flow rate actually decreases due to a local pressure increase in the intake port immediately after the intake valve is opened.
Therefore, when the blowback flow rate is calculated based on the intake pipe pressure (surge tank pressure) and the cylinder internal pressure as in the device of the above Japanese Patent Application No. 2001-316350, the valve overlap particularly increases the blowback flow rate. When the operating condition is large, there is a problem that the calculation accuracy of the in-cylinder charged intake air amount is greatly reduced.
[0015]
In view of the above problems, the present invention accurately calculates the blowback flow rate from the cylinder to the intake port by taking the intake port pressure into account, and can improve the calculation accuracy of the in-cylinder charged intake air amount. An object of the present invention is to provide an air amount estimation device.
[0016]
[Means for Solving the Problems]
According to the first aspect of the invention, the intake valve passage intake flow rate is calculated based on a predetermined intake valve model calculation formula using the intake pipe pressure and the in-cylinder pressure, and the calculated intake valve passage intake flow rate is calculated. In the intake air amount estimation device for an internal combustion engine that uses the intake valve passage intake flow rate to calculate a cylinder intake port pressure based on a predetermined intake port model calculation formula using the intake valve passage intake flow rate An intake air amount estimation device for an internal combustion engine is provided that includes an estimation means and estimates a cylinder intake charge amount based on the estimated intake port pressure and cylinder internal pressure.
[0017]
In other words, in the first aspect of the present invention, the intake port pressure estimating means for estimating the intake port pressure is provided separately from the intake pipe pressure using the intake air flow rate passing through the intake valve. Even in this case, the intake port pressure is accurately estimated, and the blowback flow rate is accurately calculated based on the estimated intake port pressure and the cylinder internal pressure. For this reason, it is possible to accurately estimate the in-cylinder charged intake amount even during operation with a large valve overlap and a large amount of blown-back gas.
[0018]
According to a second aspect of the present invention, the intake port pressure estimating means is the same as the intake pipe pressure when there is no blowback in which the residual gas in the cylinder passes through the intake valve and flows back to the intake port. The intake air of the internal combustion engine according to claim 1, wherein a value obtained by adding the pressure increase due to the blow-back to the intake pipe pressure is calculated as the intake port pressure when the blow-back occurs. A quantity estimation device is provided.
[0019]
That is, in the second aspect of the invention, the intake port pressure estimating means calculates the intake port pressure as a pressure higher than the intake pipe pressure in consideration of the pressure increase due to the blowback when the blowback occurs. As a result, the amount of blow-back is suppressed compared to the conventional case where the intake pipe pressure and the intake port pressure are assumed to be equal even when blow-back occurs, and the cylinder intake charge amount is accurately estimated. It becomes possible to do.
[0020]
Further, in a normal intake flow in which no blowback occurs, the speed of the intake flow that flows into the cylinder through the intake port is much smaller than that of the blowback flow. For this reason, a large pressure difference does not occur between the intake pipe (surge tank) and the intake port. Therefore, when there is no blowback, the intake port pressure can be calculated using the same value as the intake pipe pressure as the intake port pressure without calculating the intake port pressure separately.
[0021]
According to a third aspect of the present invention, when the blow-back occurs, the intake port pressure estimating means uses a virtual volume determined according to the engine operating state based on a predetermined relationship, and The intake port pressure is calculated on the assumption that the pressure increase width in the space when the gas flowing back to the intake port due to the reflow flows into the space of the virtual volume is the pressure increase due to the reflow. An intake air amount estimation device for an internal combustion engine according to 2, is provided.
[0022]
That is, in the invention of claim 3, it is assumed that the intake port is a closed volume portion having a volume determined according to the engine operating state, and when the blow-back occurs, the blow-back gas flows into the closed volume portion, thereby increasing the pressure of the intake port. Is assumed to occur. As a result, the pressure increase of the intake port can be calculated according to the amount of gas blown back, and the in-cylinder charged intake amount can be accurately calculated.
[0023]
According to the invention described in claim 4, the virtual volume further includes an amount of gas flowing into the space when the gas flowing backward to the intake port by the blow-back flows into the space of the virtual volume. The intake air amount estimation device for an internal combustion engine according to claim 3, wherein the intake air amount estimation device changes based on a predetermined relationship in accordance with a pressure increase width in the space.
[0024]
That is, in the invention of claim 4, as in claim 3, when blowback occurs, the intake port is assumed to be a closed volume portion, and the pressure of the intake port is determined according to the amount of blowback gas flowing into the volume portion. Calculate the rise. However, in the invention of claim 4, it is assumed that the volume of the closed volume portion is not constant and can be expanded according to the internal gas pressure, that is, the virtual volume of the intake port is changed according to the pressure increase of the intake port. The intake port pressure is calculated assuming that it changes. For example, the expansion of the volume can be assumed that the gas inside the volume expands the volume against resistances such as elastic resistance (spring), frictional resistance (dashpot) and the like. In this way, by modeling the intake port as a volume part of a variable volume, it becomes possible to calculate the amount of blown back gas more accurately, so it is possible to calculate the in-cylinder charged intake amount more accurately. Become.
[0025]
According to the fifth aspect of the present invention, the intake port pressure estimating means is configured to perform the blowback based on the pressure loss in the intake valve calculated based on the intake valve passage intake flow rate when the blowback occurs. The intake air amount estimation device for an internal combustion engine according to claim 2, which calculates an increase in pressure.
[0026]
That is, in the fifth aspect of the present invention, the pressure drop in the intake valve based on the intake flow rate that passes through the intake valve at the time of blowback (that is, the reverse flow rate due to blowback) Pressure loss), and an intake port pressure increase is calculated based on the pressure loss. Thereby, since the intake port pressure is calculated according to the blowback flow rate at the time of blowback, the blowback flow rate is accurately calculated, and the calculation accuracy of the in-cylinder charged intake air amount is improved.
[0027]
According to a sixth aspect of the present invention, the intake port pressure estimating means flows into the intake pipe from the intake port calculated based on the intake port pressure and the intake pipe pressure when the blow-back occurs. The intake air amount estimation device for an internal combustion engine according to claim 2, wherein an intake port pressure increase due to the blowback is calculated based on a gas flow rate and an intake valve passage intake flow rate.
[0028]
That is, in the invention of claim 6, the amount of gas flowing into the intake port when the blow-back occurs (that is, the blow-back flow rate flowing back to the intake port through the intake valve) and the amount of gas flowing out from the intake port The intake port pressure is calculated based on the law of conservation of mass using (the amount of gas flowing back from the intake port to the surge tank). The amount of gas flowing into the intake port and the amount of gas flowing out of the intake port can be obtained as, for example, the gas flow rate passing through the throttle, based on the intake port pressure and cylinder pressure, and the intake port pressure and intake pipe pressure, respectively. it can. Thereby, since the intake port pressure is accurately calculated, the calculation accuracy of the in-cylinder charged intake air amount is improved.
[0029]
According to the seventh aspect of the present invention, the intake port pressure estimating means predetermines an increase in the intake port pressure after opening the intake valve according to the engine operating state when the blow-back occurs. An intake air amount estimation device for an internal combustion engine according to claim 2, wherein the intake air amount estimation device has an amplitude and a period determined based on the relationship, and is calculated as a sine function starting from the start of intake valve opening.
[0030]
That is, in the invention of claim 7, it is assumed that the time change of the intake port pressure rise after the intake valve opening when the blowback occurs is a sine function, and the amplitude and period of this function based on, for example, experimental values. Is obtained in advance as a function of the engine operating state (for example, engine speed, load, valve overlap amount, etc.). By calculating the intake port pressure using such a function based on the actual measurement value, an accurate blow-back flow rate that is closer to the actual value is calculated, so that the calculation accuracy of the in-cylinder charged intake air amount is improved.
[0031]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the accompanying drawings.
FIG. 1 is a diagram showing a schematic configuration of an embodiment when the intake air amount estimation device of the present invention is applied to a spark ignition engine.
[0032]
In FIG. 1, reference numeral 10 indicates a spark ignition type internal combustion engine body. In the present embodiment, a four-cylinder gasoline engine is used as the internal combustion engine 10, but the present invention is not limited to a four-cylinder gasoline engine and can be applied to other types of internal combustion engines.
[0033]
The engine 10 includes a cylinder 21, a piston 22, and a crankshaft 24, and a combustion chamber 25 is formed in the upper portion of the piston 22 in the cylinder 21.
The cylinder head above the cylinder 21 includes an intake port 31 communicating with the combustion chamber 25, an intake valve 32 that opens and closes the intake port 31, a phase angle of the intake camshaft that drives the intake valve 32, and a valve lift amount of the intake valve 32. Is provided with an intake valve control device 33 capable of continuously changing. An actuator 33 of the intake valve control device 33 is indicated by 33a in FIG. The intake port 31 is provided with a fuel injection valve 39 for injecting fuel into the intake port.
[0034]
The cylinder head above the cylinder 21 is provided with an exhaust port 34 communicating with the combustion chamber 25, an exhaust valve 35 for opening and closing the exhaust port, and an exhaust camshaft 36 for opening and closing the exhaust valve 35, as with the intake port 31. ing.
Further, the cylinder head above the cylinder 21 is provided with an ignition plug 37 facing the combustion chamber 25 and an ignition circuit (igniter) 38 for supplying a high voltage to the ignition plug.
[0035]
The intake system 40 of the engine 10 includes an intake pipe 41 connected to an intake port 31 of each cylinder of the engine via an intake manifold, a surge tank 44 provided in the intake pipe 41, and an air cleaner 42 provided at the inlet of the intake pipe 41. A throttle valve 43 is provided on the upstream side of the surge tank 44 of the intake pipe 41.
[0036]
In this embodiment, the throttle valve 43 is provided with an actuator 43a of an appropriate type such as a stepper motor, and is a so-called electronic throttle valve that can take an opening degree independent of the driver's accelerator pedal 82 operation. Yes.
[0037]
The intake pipe 41 upstream of the throttle valve 43 is provided with an air flow meter 61 that detects the amount of intake air that is drawn into the intake pipe and flows through the throttle valve 43. In the present embodiment, the air flow meter 61 is a hot-wire flow meter, and an intake air temperature sensor 62 that detects the intake air temperature and an atmospheric pressure sensor 63 that detects the atmospheric pressure are provided in the vicinity of the air flow meter 61.
[0038]
Further, the exhaust system 50 of the engine 10 includes an exhaust pipe 52 connected to the exhaust port 34 of each cylinder via an exhaust manifold 51. In FIG. 1, reference numeral 53 denotes a catalytic converter that houses an exhaust purification catalyst.
An electronic control unit (ECU) for controlling the engine 10 is indicated by 80 in FIG. The ECU 80 is a microcomputer having a known configuration in which a CPU 81, a ROM 82, a RAM 83, an input / output interface 85, a backup RAM 84 that can retain stored contents even when the power is turned off, and the like are connected by a bidirectional bus. In addition to performing basic control of the engine such as injection control, in the present embodiment, accurate intake air amount estimation considering blowback is performed based on an air model described later.
[0039]
For these controls, the input / output interface 85 of the ECU 80 receives a signal corresponding to the amount of intake air (weight flow rate) flowing into the throttle valve 43 from the hot-wire air flow meter 61, and from the intake air temperature sensor 62 and the atmospheric pressure sensor 63. , The intake air temperature (atmospheric temperature) and the signal corresponding to the atmospheric pressure are input. Outputs from the intake air temperature sensor 62 and the atmospheric pressure sensor 63 are used to correct the intake air amount detected by the air flow meter 61 in accordance with the temperature and the atmospheric pressure.
Although not described in detail, in this embodiment, the intake air amount detected by the air flow meter 61 is used to correct the intake air amount calculated using an air model described later.
[0040]
An input / output interface 85 of the ECU 80 receives a voltage signal corresponding to the throttle valve 43 opening from a throttle valve opening sensor 64 provided near the throttle valve 43, and an accelerator opening sensor provided near the accelerator pedal 82 of the engine. A voltage signal corresponding to the accelerator pedal depression amount (accelerator opening) of the driver is input from 71, and an intake pipe pressure is input from an intake pressure sensor 65 disposed in the surge tank 44.
[0041]
Further, in the present embodiment, the intake valve control device 33 includes an intake valve lift sensor 67, and the exhaust pipe 52 is disposed upstream of the catalytic converter 53 with O. ₂ The sensor 70 is provided with a cooling water temperature sensor 69 in the cooling water passage of the engine 10, and inputs the lift amount of the intake valve 32, the exhaust oxygen concentration, and the engine cooling water temperature to the input / output interface 85 of the ECU 80, respectively. is doing.
[0042]
In FIG. 1, reference numeral 66 denotes a cam angle sensor disposed near the intake camshaft, and reference numeral 68 denotes a crank angle sensor disposed near the crankshaft.
The cam angle sensor 66 outputs a cam position signal every time the crankshaft 24 rotates 180 °, and the crank angle sensor 68 outputs a crank rotation angle pulse signal every time the crankshaft 24 rotates by a certain angle (for example, 10 °). At the same time, every time the crankshaft 24 takes the reference rotation position (for example, every time the first cylinder reaches the top dead center), a reference crank position pulse signal having a wider pulse width than the crank rotation angle pulse signal is output.
[0043]
These signals are input to the interface 85 of the ECU 80. For example, the ECU 80 calculates the engine rotation speed (number of rotations) from the interval between the crank rotation angle pulse signals, and calculates the rotation position (crank angle) of the crankshaft from the number of crank rotation angle pulse signals after the input of the reference crank position pulse signal. calculate. The cam position signal input from the cam angle sensor 66 is used to calculate the valve timing of the intake valve together with the crank angle.
[0044]
In the present embodiment, the exhaust port of the cylinder is provided with an exhaust temperature sensor and an exhaust pressure sensor (not shown) for detecting the exhaust pressure Pe and the temperature Te of the exhaust port, respectively. A signal corresponding to the exhaust pressure Te is input to the interface 85 of the ECU 80.
[0045]
In the present embodiment, the ECU 80 calculates the fuel injection amount fc in the cylinder using the intake amount Mc charged in the cylinder 20 of the engine 10 as fc = KI × Mc. KI is a conversion coefficient determined according to the engine air-fuel ratio.
[0046]
Further, the ECU 80 calculates an intake air amount (in-cylinder charged intake air amount) Mc charged into the cylinder 20 using an air model calculation formula that models the intake system 40 of the engine 10.
[0047]
In the present embodiment, basically, the in-cylinder charged intake air amount calculation method based on the air model proposed by the applicant of the present application in Japanese Patent Application No. 2001-316350 is used. Therefore, in order to explain the method for calculating the in-cylinder charged intake air amount Mc of the present embodiment, first, the in-cylinder charged intake air amount Mc calculating method in Japanese Patent Application No. 2001-316350 will be briefly described.
[0048]
In this embodiment, the intake system 40 of the engine 10 is disassembled into five elements: a throttle valve, an intake pipe including a surge tank, an intake valve, a cylinder, and an exhaust valve. The intake air amount (in-cylinder charged intake air amount) Mc that is finally filled into the cylinder is calculated by formulating the pressure, temperature, and flow rate of the flowing intake air.
[0049]
Each model will be described below.
(1) Throttle valve model
The throttle valve model is basically given as a gas flow passing through a throttle (throttle valve opening) as shown in FIG.
[0050]
That is, when the throttle valve 43 is considered as a throttle as shown in FIG. 2, as is well known, the gas flow rate mt passing through the throttle valve 43 is the gas pressure and temperature before and after the throttle valve, and the throttle valve 43 opening area. And is expressed by the following formula for calculating the throttle valve model.
[0051]
[Expression 1]

[0052]
[Expression 2]

[0053]
Here, Equation (1) indicates the intake valve passage intake flow rate in the normal intake flow (forward flow) direction when the pressure on the upstream side (atmosphere side) of the throttle valve 43 is higher than the pressure on the downstream side (intake pipe side). The throttle valve passing air flow rate when the pressure on the upstream side of the throttle valve 43 is lower than the pressure on the downstream side and a reverse flow is generated is shown.
[0054]
In Equations 1 and 2, Ct (θt) is a flow coefficient, At (θt) is a throttle valve opening area, and Ct (θt) and At (θt) are functions of the throttle valve opening θt, respectively. Pa is the pressure on the upstream side of the throttle valve, and in this embodiment, it is atmospheric pressure, and Pm is the pressure on the downstream side of the throttle valve, and in this embodiment, it represents the intake pipe pressure. Ta and Tm are the intake air temperatures on the upstream and downstream sides of the throttle valve, respectively, and in this embodiment, the atmospheric temperature and the intake pipe temperature, R is the gas constant, and κ is the specific heat ratio of the intake air. In the present embodiment, R and κ are treated as constant values.
[0055]
In the present embodiment, the flow coefficient Ct (θt) and the throttle valve opening area At (θt) are calculated in advance for each throttle valve opening θ and stored in the ROM 82 of the ECU 80 in the form of a numerical map.
As shown in the equations (1) and (2), the throttle valve model of the present embodiment is configured to input the atmospheric pressure Pa and the intake pipe pressure Pm, the atmospheric temperature Ta or the intake pipe temperature Tm, and the throttle valve opening θt and input the throttle valve. It is a model calculation formula using the intake flow rate mt as an output.
[0056]
(2) Intake pipe model
FIG. 3 shows an intake pipe model. The intake pipe model calculates the intake pipe pressure Pm and the temperature Tm using the mass conservation law and the energy conservation law based on the intake flow rate flowing into and out of the fixed volume space.
[0057]
If the volume of the intake pipe 41 including the surge tank 44 extending from the throttle valve 43 to the intake valve 32 is Vm, the intake flow flowing into the volume Vm is a throttle valve passing intake flow as shown in FIG. The flow rate is mt and the intake air temperature is Ta.
[0058]
The intake flow rate flowing out from the volume Vm is the intake flow rate mc that passes through the intake valve and fills the cylinder, and the temperature of the intake flow is the intake pipe temperature Tm.
[0059]
Accordingly, the following formula 3 is derived from the mass conservation law, and the following formula 4 is derived from the energy conservation law.
[0060]
[Equation 3]

[0061]
[Expression 4]

[0062]
In the present embodiment, in order to perform arithmetic processing in the ECU 80, the following formulas 5 and 6 obtained by integrating and discretizing the formulas 3 and 4 are used.
[0063]
[Equation 5]

[0064]
[Formula 6]

[0065]
Here, the subscript k represents a value at the time of the current calculation execution, k−1 represents a value at the time of the previous calculation execution, and Δt represents a difference time (for example, a mesh of about 8 ms) at the time of discretization. Further, mcAVE in the equations (5) and (6) is a time average value over one cycle of the intake stroke of a cylinder intake air flow rate mc described later.
[0066]
The intake pipe temperature Tm is calculated as Tm = Pm / (Pm / Tm) using Pm / Tm obtained from Equation 5 and Pm obtained from Equation 6.
(3) Intake valve model
The intake valve model is a model representing the intake flow passing through the intake valve 32 as shown in FIG. 4, and is basically represented as an intake flow passing through the throttle (intake valve opening area), as in the throttle valve model. The
[0067]
Accordingly, the intake valve model is expressed by the following equations (7) and (8), similar to the throttle valve model (Equations 1 and 2).
[0068]
[Expression 7]

[0069]
[Equation 8]

[0070]
Here, mc is an intake air flow rate passing through the intake valve, that is, an intake air amount filled in the cylinder within a unit time, Pm and Tm are intake pipe pressure and temperature, and Pc and Tc are in-cylinder pressure and temperature. Further, Civ (L) is a flow coefficient, a function of the lift amount L of the intake valve, and Aiv (L) is an opening area of the intake valve 32, and is also a function of the lift amount L.
[0071]
In this embodiment, the product values Civ (L) · Aiv (L) are calculated for each L value and stored in the ROM 82 of the ECU 80 in the form of a numerical table.
Like the throttle valve model (Equation 1 and Equation 2), Equation 7 shows the case where the intake pipe pressure Pm is larger than the cylinder internal pressure Pc and the intake air is flowing into the cylinder. The case where the intake air flows back to the intake pipe from the left is shown.
[0072]
As shown in the equations (7) and (8), the intake valve model of the present embodiment receives the intake pipe pressure Pm, the cylinder internal pressure Pc, the intake pipe internal temperature Tm or the cylinder internal temperature Tc, and the intake valve lift amount L as inputs. The valve passage intake flow rate mc is output.
The average intake air flow rate mcAVE used in the above-described intake pipe model calculation formulas (5) and (6) is obtained by multiplying the value Mc obtained by integrating the intake valve passing air flow rate mc calculated in the equations (7) and (8) over the entire intake stroke with a crank angle. It is a value divided by time T (180 CA) corresponding to 180 °.
[0073]
That is, mcAVE = Mc / T (180 CA)
(4) Cylinder model
In the cylinder model of the present embodiment, as shown in FIG. 5, basically, intake air having a flow rate mc and temperature Tm per unit time flows from an intake valve into a closed space (Vc is a function of a crank angle) having a volume Vc (variable). Considering the flow of exhaust gas that passes through the exhaust valve from the exhaust port and flows back into the cylinder while the intake valve is open, the flow rate me and the temperature Te are set in addition to the intake flow from the intake valve. It is given as a model for exhaust flow.
[0074]
That is, in the cylinder model of the present embodiment, mc, me, Tm, Te (and crank angle) are input, and the cylinder internal pressure Pc, temperature Tc, and the total amount Mc of the intake air filled in the cylinder during the intake stroke are output. To do.
[0075]
The cylinder model is basically expressed by the following formula 9 based on the energy conservation law.
[0076]
[Equation 9]

[0077]
Here, Qw is the amount of heat given to the intake air from the cylinder wall surface, the intake port wall surface, or the like.
[0078]
Further, the in-cylinder temperature Tc is expressed by the following equation 10 from the gas equation of state.
[0079]
[Expression 10]

[0080]
Here, Mc1 is a charging intake air amount from the time when the intake valve is opened, and is given as an integral value of the inflow gas flow rate (mc + me) from the intake valve and the exhaust valve from the time when the intake valve is opened until the present time (Tc calculation time). It is done.
[0081]
Since the above formula 9 cannot be calculated by the ECU 80 in this form, it is converted into a discretized form and used in the form of the following formula 11.
[0082]
## EQU11 ##

[0083]
In addition, Δt in Equation 11 is a difference time when Discretizing Equation 9 (Δt is set shorter than that in the intake pipe model in the cylinder model, for example, set to a mesh of about 1 ms).
[0084]
Further, the exhaust valve passage flow rate me is obtained by an exhaust valve model described below.
(5) Exhaust valve model (Fig. 7, M50)
Similar to the intake valve model, the exhaust valve model is a model that outputs the exhaust valve passage flow rate me by inputting the pressure and temperature of the exhaust port, Pe and Te, and the cylinder pressure, temperature Pc and Tc, and the crank angle.
[0085]
FIG. 6 shows an exhaust valve model. As shown in FIG. 6, the exhaust valve model is basically the same as the intake valve model, and the exhaust valve passage flow rate is given by Expression 12 or Expression 13.
[0086]
[Expression 12]

[0087]
[Formula 13]

Equation 12 represents the case where exhaust flows backward from the exhaust port into the cylinder (that is, when the cylinder internal pressure Pc is lower than the exhaust port pressure Pe), and Equation 13 represents the case where exhaust flows from the cylinder to the exhaust port (Pc is higher than Pe). Case) respectively. Cev (L) and Aev (L) are a flow coefficient and an opening area of the exhaust valve 35, respectively, and both are functions of the lift amount L of the exhaust valve.
[0088]
In this embodiment, the exhaust port pressure Pe and the exhaust temperature Te are detected by an exhaust pressure sensor and an exhaust temperature sensor (not shown) provided at the exhaust port. However, since the pressure loss of the exhaust system is relatively small, Instead of detecting the exhaust pressure with the atmospheric pressure sensor, it is also possible to perform the calculation assuming that Pe is approximately equal to the atmospheric pressure.
Similarly, since the exhaust temperature Te is substantially equal to the in-cylinder gas temperature, the in-cylinder temperature Tc may be substituted for the exhaust temperature Te without providing an exhaust temperature sensor.
[0089]
FIG. 7 is a block diagram showing the entire air model composed of the throttle valve model, the intake pipe model, the intake valve model, the cylinder model and the exhaust valve model described above. Using FIG. 6, the model calculation formula of each model element, the input and output, the relationship between the model elements, the calculation process of the in-cylinder charged intake air amount Mc, and the like are summarized again.
[0090]
(A) Throttle valve model (Fig. 7, M10)
(1) Calculation formula
[0091]
[Expression 14]

[0092]
[Expression 15]

[0093]
(2) Input
・ Throttle valve opening θt
・ Atmospheric pressure Pa (Atmospheric pressure sensor 63 output)
・ Atmospheric temperature Ta (intake air temperature sensor 62 output)
・ Intake pipe pressure Pm
・ Intake pipe temperature Tm (in the case of formula 15 (back flow))
(3) Output
・ Throttle valve passage intake flow rate mt
(B) Intake pipe model (Fig. 7, M20)
(1) Calculation formula
[0094]
[Expression 16]

[0095]
[Expression 17]

[0096]
[Expression 18]

[0097]
(2) Input
・ Intake flow rate through throttle valve mt (k-1)
・ Intake valve passage intake flow rate (average value) mcAVE (k-1)
・ Pm (k-1), Tm (k-1), (Pm / Tm) (k-1)
(3) Output
・ Intake pipe pressure Pm (k)
・ Intake pipe temperature Tm (k)
(C) Intake valve model (Fig. 7, M30)
(1) Calculation formula
[0098]
[Equation 19]

[0099]
[Expression 20]

[0100]
(2) Input
・ Intake pipe pressure Pm
・ Intake pipe temperature Tm
・ Cylinder pressure Pc
・ In-cylinder temperature Tc (in case of number 20 (back flow))
・ Valve lift amount L (calculated from crank angle)
(3) Output
Intake valve passage intake flow rate (instantaneous value) mc
・ Intake valve passage intake flow rate mcAVE (average value)
(D) Cylinder model (Fig. 7, M40)
(1) Calculation formula
[0101]
[Expression 21]

[0102]
[Expression 22]

[0103]
(2) Input
・ Cylinder volume Vc (calculated from crank angle)
・ Intake valve passage intake volume (instantaneous value) mc
・ Intake pipe temperature Tm
Exhaust valve passage intake air amount me
・ Exhaust temperature Te
(3) Output
・ Cylinder pressure Pc
・ In-cylinder temperature Tc
・ In-cylinder charged intake air amount Mc
(E) Exhaust valve model (Fig. 7, M50)
(1) Calculation formula
[0104]
[Expression 23]

[0105]
[Expression 24]

[0106]
(2) Input
・ Exhaust pipe pressure Pe
・ Exhaust pipe temperature Te
・ Cylinder pressure Pc
・ In-cylinder temperature Tc (in case of formula 24)
・ Valve lift amount L (calculated from crank angle)
(3) Output
Exhaust valve passage intake flow rate (instantaneous value) me
As described above, the models other than the throttle valve model and the exhaust valve model, that is, the input values of the intake pipe model, the intake valve model, and the cylinder model can use output values of other models except for the crank angle. The crank angle can be calculated as a function of the engine speed NE.
[0107]
Further, the throttle valve model can be input using output values of other models except for the throttle valve opening θ (t), the atmospheric pressure Pa, and the atmospheric temperature Ta. Since the atmospheric pressure Pa and the atmospheric temperature Ta do not change in a short time, they can be considered to be substantially constant values.
Accordingly, in the air model shown in FIG. 7, the cylinder intake charge amount mc can be calculated basically when the throttle valve opening θ (t), the engine speed NE, the exhaust pressure Pe, and the temperature Te are determined.
[0108]
Specifically, for example, when the engine is started, Pm and Pc are set to the atmospheric pressure Pa as initial values, Tm and Tc are set to the atmospheric temperature Ta, and the flow rates mt and mc are set to 0, respectively, and the throttle valve opening θ (t ) And the crank angle are used to calculate the above formulas 14 to 24 (Pe and Te use actual measured values).
Then, using the values such as Pm, PcTm, Tc, mt, mc, and me obtained as a result, the next calculation of Expressions 14 to 24 is performed. This calculation is performed at a predetermined interval (for example, about 8 ms) from the start of the engine using values of θ (t) and NE at that time and values such as Pm, tm, mt, mc, and me obtained in the previous calculation. And repeat. As a result, Pm, tm, mt, mc, etc. should converge to actual values in a short time after the start of calculation, and coincide with the actual in-cylinder charged intake air amount.
[0109]
In the above-described conventional air model, the intake air amount estimation is performed in consideration of the case where the backflow from the cylinder to the intake port occurs in the intake valve model (FIG. 4) (Equation 20). Even in the case where a large amount of blowback occurs, the in-cylinder charged intake amount should be accurately estimated.
However, as described above, since the conventional air model does not take into account the local pressure increase at the intake port at the time of blowback, the estimation accuracy of the in-cylinder charged intake amount is particularly large when the blowback amount is large. The problem of deteriorating arises.
[0110]
In the present invention, an intake pipe and an intake port are separately modeled with respect to the conventional air model, and the intake port pressure at the time of occurrence of blowback is calculated as a value different from the intake pipe pressure using the intake port model. Thus, the estimation accuracy of the in-cylinder charged intake air amount is improved.
FIG. 8 is a block diagram similar to FIG. 7 for explaining the overall configuration of the air model of the present embodiment.
[0111]
As shown in FIG. 8, the present embodiment is different in that an intake port model M25 is provided between the intake pipe model M20 and the intake valve model M30.
Although details of the intake port model M25 will be described later, in this embodiment, the intake port pressure Pp is calculated using the intake port model M25, and the intake valve passage flow rate mc is expressed by the following formulas 25 and 26 is used for calculation. Here, Formula 25 is a case where blowback to the intake port does not occur (that is, when Pp ≧ Pc), and Formula 26 is a case where blowback occurs to the intake port (that is, Pp <Pc).
[0112]
[Expression 25]

[0113]
[Equation 26]

[0114]
Here, as will be described later, in the intake port model M25 of the present embodiment, when the blowback from the cylinder occurs, the intake port pressure Pp is calculated as a pressure higher than Pm separately from the intake pipe pressure Pm. When no blow-back occurs, a value equal to the intake pipe pressure Pm is used as the intake port pressure Pp, and the intake port pressure Pp is not calculated separately from the intake pipe pressure Pp.
[0115]
That is, the air model in FIG. 8 is the same as the model in FIG. 7 when there is no blowback from the cylinder.
As a result, the local pressure increase in the intake port due to the blow back air flow at the start of the intake valve opening is accurately reflected in the intake valve passage flow rate. The charging intake air amount is calculated.
[0116]
FIG. 9 is a flowchart showing an outline of an intake valve passage flow rate calculation operation using the intake port model M25 of the present embodiment. This operation is performed as a routine that is repeated by the ECU 80 at a relatively short interval of about 1 ms, for example.
In the operation of FIG. 9, it is determined whether or not the blowback to the intake port has occurred. If the blowback has occurred, the intake port pressure taking into account the local pressure increase of the intake port due to the blowback to the intake port And the intake valve passage flow rate is calculated based on the intake port pressure Pp using the intake valve model calculation formula (26). Further, when the blow-back does not occur, the intake valve passage intake flow rate is calculated using Equation 25 assuming Pp = Pm and Tp = Tm. In this case, Equation 25 is the same as Equation 20.
[0117]
That is, in FIG. 9, at step 901, the intake valve passage intake air amount mc is calculated using the intake port pressure Pp calculated at the time of the previous operation execution.
In step 903, it is determined whether or not a blowback from the cylinder to the intake port is currently occurring. This determination may be made based on whether or not the previously calculated intake port pressure Pp is lower than Pc. However, if the presence or absence of blow-back is determined only by comparing the pressure, the pressure pulsation of the intake port is determined. Error due to influence may occur. Therefore, in this embodiment, the intake valve passage flow rate (Equation 25) in the case of Pp ≧ Pc is positive, and the intake valve passage flow rate (Equation 26) in the case of Pp <Pc is a negative value. The flow rate mc is integrated, and when this integrated value becomes a negative value, it is determined that blowback from the cylinder to the intake port has occurred.
[0118]
If blowback has occurred in step 903, that is, if the integrated value of the intake valve passage intake air amount mc calculated in step 901 is a negative value, the process proceeds to step 907, and an intake port model calculation formula described later is performed. Based on the above, the intake port pressure Pp is calculated.
Further, when the integrated value of the intake valve passage flow rate mc is positive in step 903, that is, when the blow-back does not occur, the intake port pressure Pp is calculated by the above-described intake pipe model calculation formula (Expression 16 to Expression 18). The calculated intake pipe pressure Pm is set, and the current operation is terminated.
[0119]
As a result, the intake valve pressure calculated based on the intake port model calculation formula is used when blowback occurs during the next operation, and the intake valve pressure is used when blowback does not occur. Since the passage flow rate is calculated, an accurate intake valve passage flow rate considering the influence of blowback is calculated.
Accordingly, by adding the intake valve passage flow rate over one cycle of the intake stroke in consideration of positive and negative, the amount Mc of fresh air filled in the cylinder can be accurately calculated.
[0120]
Next, the intake port model used for calculating the intake port pressure Pp in step 907 will be described.
The number of intake port models that can be used in the present embodiment is not limited to one, and several intake port models can be used. An example of an intake port model that can be used in the present embodiment will be described below. In the operation of FIG. 9, the intake port pressure is calculated based on one of the following intake port models.
[0121]
(1) First embodiment
FIG. 10 is a diagram schematically illustrating the intake port model of the present embodiment.
In this embodiment, when blow-back occurs, the intake port is assumed to be a closed space having a predetermined virtual volume, and the pressure increase when the entire amount of blow-back gas from the cylinder flows into this closed space is obtained. The pressure obtained by adding this pressure increase to the intake pipe pressure is used as the intake port pressure Pp.
When a high-speed airflow flows from the cylinder into the intake port in a short time immediately after the intake valve is opened due to the blow-back, a pressure rise locally occurs in the intake port due to the inflowing gas. This local pressure increase is considered to increase as the inflowing gas becomes faster and the flow rate increases.
[0122]
Therefore, in this embodiment, it is assumed that the intake port is a closed space (container) having a virtual volume determined by the engine operating state, and the pressure increase caused by the blown airflow flowing into the closed space is equal to the pressure increase of the intake port. Assume that
Now, assuming that the volume of the intake port as a closed space is Vp and the blowback flow rate per unit time, that is, the intake valve passage flow rate at the time of blowback is mc, the pressure change per unit time of the intake port Pp due to the inflow of blowback gas dPp / dt Is expressed by the following equation (27) from the equation of state.
[0123]
[Expression 27]

[0124]
Here, R is a gas constant of the blown-back gas, and Tp is a gas temperature in the intake port, which is considered to be substantially equal to the gas temperature Tc in the cylinder.
Therefore, the intake port pressure increase width ΔPp due to the blow-back is equal to the integral value of Formula 27 after the start of blow-back, and is expressed by the following Formula 28 when the change in Tc is ignored.
[0125]
[Expression 28]

[0126]
Here, the integral value of the intake valve passage flow rate mc is obtained by integrating the flow rate mc (return flow rate) calculated by Equation 26, and the intake port pressure Pp is
Pp = Pm + ΔPp
Is calculated as
[0127]
In step 907 of FIG. 9, the operation of calculating the intake port pressure Pp as described above and the operation of calculating the next intake valve passage flow rate mc using the calculated value of Pp (FIG. 9, step 901) are sequentially repeated. Thus, the intake valve passing intake air amount mc is accurately calculated.
The virtual volume Vp of the intake port is calculated in advance based on experimental results, and is stored in the ROM 82 of the ECU 80 as a numerical map using, for example, the engine speed NE and the intake valve timing VT of the engine 1, as shown in FIG. It is read out by the ECU 80 based on the actual engine speed and valve timing every time the operation is executed.
[0128]
For example, as the engine speed increases, the amount of blown-back gas decreases. For this reason, the virtual volume Vp is set to a smaller value as the rotational speed increases. Further, the amount of blown-back gas increases as the valve timing advances (the valve overlap increases), but the pressure increase in the intake port does not increase corresponding to the amount of blow-back gas. For this reason, the virtual volume Vp is set to a larger value as the valve timing of the intake valve advances.
[0129]
(2) Second embodiment
In the first embodiment, the virtual volume Vp of the intake port is a constant value determined according to the engine operating state (for example, the engine speed and the valve timing).
However, in actual operation, the intake port is connected to a surge tank, not a closed space, and the entire amount of the blowback flow rate flowing into the intake port does not remain in the intake port. The inflowing gas becomes a pressure wave while compressing the gas existing in the intake port and the intake pipe, and proceeds in the intake port toward the surge tank.
[0130]
Therefore, in this embodiment, it is assumed that the virtual volume of the intake port expands according to the internal pressure. Further, as shown in FIG. 11, when the virtual volume of the intake port is expanded, a resistance represented by a spring S having a spring constant k and a dashpot D having a damping constant c at the boundary with the existing gas in the intake port. Suppose that the gas blows back against this resistance.
[0131]
Now, assuming that the flow rate of the blown-back gas flowing into the intake port is mc and the temperature is Tc, the energy conservation law of the gas flowing into the intake port is expressed by the following equation (29).
[0132]
[Expression 29]

[0133]
Here, W is a work performed by the blowback gas flowing into the intake port. This work is pushed by the virtual movable boundary wall 110 biased by the spring S and the dashpot D as shown in FIG. When expressed as work to be performed, the following number 30 is obtained.
[0134]
[30]

[0135]
Here, A is a port cross-sectional area (a constant value).
When the right side of Expression 30 is expanded and k / A = K and c / A = C are set, Expression 31 is obtained.
[0136]
[31]

[0137]
Therefore, Expression 31 is obtained by substituting Expression 31 into Expression 29.
[0138]
[Expression 32]

[0139]
Further, the following equation 33 is established among Vp, Pp, and mc from the state equation.
[0140]
[Expression 33]

[0141]
Further, when the expression 32 is discretized, the following expression 34 is obtained.
[0142]
[Expression 34]

[0143]
The intake port pressure Pp is calculated by performing sequential calculation of Formula 34 using the intake valve passage flow rate mc (return flow rate) calculated by Formula 26 and Formula 33. The constants K and C in Equation 34 are obtained in advance as a function of the engine speed NE and the intake valve timing VT by experiment, and are stored in the ROM 82 of the ECU 80 as a numerical map using NE and VT. Each calculation is read from this map based on the actual engine speed and valve timing.
As a result, it is possible to calculate the intake port pressure increase due to blow-back more accurately, and the calculation accuracy of the cylinder intake charge amount is improved.
[0144]
(3) Third embodiment
In the second embodiment described above, the intake port pressure Pp is calculated based on the work of the virtual movable boundary wall 110 in FIG. 11 and the energy conservation law of the gas flowing into the intake port. In the present embodiment, In order to simplify the calculation, a balance of forces acting on both sides of the movable boundary wall 110 in FIG. 11 will be considered.
[0145]
The intake pipe pressure Pm and the force F generated by the spring S and the dashpot D are applied to the left side (intake pipe side) of the movable wall 110 in FIG. 11, and the intake air is applied to the right side (intake port side) of the movable wall 110. Since the port pressure Pp is acting, if ΔPp = Pp−Pm, then ΔPp × A = F. Here, A is the area of the movable wall 110 (cross-sectional area of the intake port), F is the total urging force of the spring S and the dashpot D, and the spring constant k of the spring S and the damping constant c of the dashpot D are used. Is given by the following equation (35).
[0146]
[Expression 35]

[0147]
Using Equation 35 and ΔPp × A = F, K = k / A ² , C = c / A ² If it puts, number 36 will be obtained.
[0148]
[Expression 36]

[0149]
Further, Vp and dVp / dt are expressed by Equation 37 and Equation 38, respectively.
[0150]
[Expression 37]

[0151]
[Formula 38]

[0152]
Since ΔPp is a local pressure increase width of the intake port due to the inflow of blown-back gas, the intake port pressure Pp is as follows.
Therefore, ΔPp is first obtained from Equation 38 using the intake valve passage flow rate mc (return flow rate) calculated by Equation 25 with Pp = Pm and Vp = 0 as initial values when the intake valve is opened. Is substituted into Equation 36 to obtain ΔPp, and Pp is further obtained from Pp = Pm + ΔPp. Then, mc is calculated by using the obtained Pp value in Expression 25, Vp and dVp / dt are calculated from Expression 37 and Expression 38 using this mc and its integrated value, and are substituted into Expression 36. It can be calculated by sequential calculation in which the operation for obtaining ΔPp is repeated.
[0153]
(4) Fourth embodiment
Next, a fourth embodiment of the intake port model will be described.
In the present embodiment, it is assumed that the intake port pressure Pp when blowback occurs is determined by the pressure loss when the blown airflow passes through the virtual pipeline including the intake valve.
[0154]
Here, when the gas flow velocity when passing through the pipe is u, the pressure loss (Pc−Pp) is u. ² Which is expressed by the following equation (39).
[0155]
[39]

[0156]
Here, ρ in the first term on the right side of Equation 39 is the density of the blown-back gas and is a function of the cylinder pressure Pc and the temperature Tc.
[0157]
The second term on the right side of Equation 39 is expressed as Equation 40.
[0158]
[Formula 40]

[0159]
Here, α is a constant, and l and d are the length and diameter of a virtual flow path including the intake valve, respectively, and are functions of the lift L of the intake valve. The values of α, l, and d are obtained in advance by an experiment using an actual engine and stored in the ROM 82 of the ECU 80.
[0160]
For Pc and Tc, values calculated by a cylinder model (FIG. 7, M40) are used. Thereby, the intake port pressure Pp is calculated by the following equation (41). Note that from equation 39, the intake port pressure increase ΔPp due to blow-back is ΔPp = Pp−Pm = Pc−Pm− (ρu ² / 2) -α (l / d) u ² Can also be expressed.
[0161]
[Expression 41]

[0162]
That is, in the present embodiment, ρ and u are calculated using mc, Pc, and Tc calculated using the cylinder model (FIG. 7, M40) and the intake valve model (FIG. 7, M30), and using Equation 41. By calculating the intake port pressure Pp and repeating the operation of calculating the next mc, Pc, and Tc, the local increase in the intake port pressure due to the blowback is reflected in the intake valve passage flow rate mc, and the in-cylinder charged intake amount is calculated. Accuracy is improved.
[0163]
(5) Fifth embodiment
Next, a fifth embodiment of the intake port model will be described.
FIG. 12 schematically shows an intake port model of the present embodiment. As shown in FIG. 12, in this embodiment, it is assumed that a part of the blown-back gas that has flowed into the intake port from the intake valve passes through the throttle 120 and flows out to the intake pipe.
[0164]
Now, the intake pipe pressure Pm, the intake port pressure Pp, the intake port temperature Tp, the intake port volume Vp, the intake valve passage flow rate (blow-back flow rate) mc, and the gas flow rate mp flowing through the throttle 120 and flowing into the intake pipe from the intake port. By using the equation of state of gas, Equation 42 is obtained.
[0165]
[Expression 42]

[0166]
Further, the flow rate mp passing through the throttle 120 is calculated by the formulas 43 and 44 of the throttle passing flow rate using Pm, Pp, Tp and the specific heat ratio κ of the blown-back gas.
[0167]
[Equation 43]

[0168]
(44)

[0169]
Here, Cp in Expression 43 is the flow coefficient of the throttle, and Ap is the opening area of the virtual throttle 120, which is determined according to the engine operating state (engine speed NE and intake valve timing VT).
[0170]
Here, assuming that the blown-back gas temperature Tp is substantially equal to the in-cylinder gas temperature Tc, the formula 42 is obtained by rewriting the formula 42 into a discrete form.
[0171]
[Equation 45]

[0172]
In the present embodiment, Pb = Pm and mp = 0 are set as initial values at the time of blowback, and the intake valve passage flow rate mc is calculated from the intake valve model using Equation 26, and the intake port pressure Pp is calculated from Equation 45 using this mc. Is calculated. Further, from the next time, using this calculated Pp, mc is calculated from Equation 26 and mp is calculated from Equation 43, and the operation of calculating the intake port pressure Pp from Equation 45 using these is repeated. As a result, the local increase in pressure due to the blowback to the intake port is accurately reflected in the intake port pressure Pp, and the calculation accuracy of the in-cylinder charged intake air amount is improved.
[0173]
(6) Sixth embodiment
In this embodiment, the increase in the intake port pressure after the intake valve is opened is approximated by a sine function. In actual operation, the intake port pressure rise after the intake valve is opened is not necessarily a sine wave. However, since the blowback becomes a problem for a very short time after the start of the intake valve opening, even if the pressure rise is approximated by a sine function, a large error does not occur.
[0174]
In the present embodiment, the pressure increase ΔPp = Pp−Pm after the intake valve is opened is expressed by Equation 46 as a sine function of the crank angle CA.
[0175]
[Equation 46]

[0176]
Here, CAio is a crank angle at which the intake valve starts to open.
[0177]
Since the amplitude a (maximum pressure increase width) and period b (pressure increase period) of the sine function in Equation 46 both depend on the flow rate of the blowback gas, the ratio of the valve overlap amount OL, the intake pressure Pm, and the exhaust pressure Pe. It is a function of (Pm / Pe), and can be expressed as Equations 47 and 48.
[0178]
[Equation 47]

[0179]
[Formula 48]

[0180]
Where C ₁ To C _Four Is a constant, and is predetermined by an experiment using an actual engine.
[0181]
FIG. 13 is a diagram illustrating an example of the intake port pressure increase width ΔPp calculated using Equations 46 to 48. In FIG. In FIG. 2, curves VT40, VT20, and VT0 indicate cases where the intake valve opening crank angle (CAio in Expression 46) is 40 degrees, 20 degrees, and 0 degrees before top dead center (BTDC), respectively.
It should be noted that the amplitude a and period b of the function of Equation 46 are stored in advance in the ROM 82 of the ECU 80 in the form of a numerical map having OL and (Pm / Pe) as arguments instead of using mathematical expressions as in Equations 47 and 48. You may make it store.
[0182]
In the present embodiment, the intake port pressure increase ΔPp after the intake valve is opened is expressed in the form of Equation 46 assuming a sine function, whereby the intake port pressure increase at the time of occurrence of blowback can be calculated easily and accurately. This makes it possible to improve the calculation accuracy of the in-cylinder charged intake air amount.
[0183]
【The invention's effect】
According to the invention described in each claim, when the intake valve passage intake flow rate is calculated based on the intake valve model calculation formula and the in-cylinder charged intake amount is calculated using the calculated intake valve passage intake flow rate, By considering the port pressure, it is possible to accurately calculate the blowback flow rate from the cylinder to the intake port, and there is a common effect that the calculation accuracy of the in-cylinder charged intake air amount can be improved.
[Brief description of the drawings]
FIG. 1 is a diagram showing a schematic configuration of an embodiment when the present invention is applied to a spark ignition engine.
FIG. 2 is a diagram illustrating a throttle valve model.
FIG. 3 is a diagram illustrating an intake pipe model.
FIG. 4 is a diagram illustrating an intake valve model.
FIG. 5 is a diagram illustrating a cylinder model.
FIG. 6 is a diagram illustrating an exhaust valve model.
FIG. 7 is a block diagram showing an entire air model in the prior art.
FIG. 8 is a block diagram showing an entire air model that wants to have an intake port model.
9 is a flowchart for explaining an intake valve passage flow rate calculation operation in the air model of FIG. 8;
FIG. 10 is a diagram illustrating a first embodiment of an intake port model.
FIG. 11 is a diagram illustrating a second embodiment of an intake port model.
FIG. 12 is a diagram illustrating a fifth embodiment of an intake port model.
FIG. 13 is a diagram showing an example of a change in intake port pressure rise calculated using the sixth embodiment of the intake port model.
[Explanation of symbols]
10 ... The whole internal combustion engine
80 ... Electronic control unit (ECU)
M10 ... Throttle valve model
M20 ... Intake pipe model
M25 ... Intake port model
M30 ... Intake valve model
M40 ... Cylinder model
M50 ... Exhaust valve model

Claims (7)

An internal combustion engine that calculates an intake valve passage intake flow rate based on a predetermined intake valve model calculation formula using an intake pipe pressure and an in-cylinder pressure, and calculates an in-cylinder charge intake amount using the calculated intake valve passage intake flow rate In an intake air amount estimation device for an engine,
Intake port pressure estimating means for calculating a cylinder intake port pressure based on a predetermined intake port model calculation formula using the intake valve passage intake flow rate,
An intake air amount estimation device for an internal combustion engine that estimates an in-cylinder charged intake amount based on the estimated intake port pressure and in-cylinder pressure. The intake port pressure estimating means calculates the same value as the intake pipe pressure as the intake port pressure when the residual gas in the cylinder passes through the intake valve and does not flow back to the intake port. 2. The intake air amount estimation device for an internal combustion engine according to claim 1, wherein when intake air pressure occurs, a pressure obtained by adding a pressure increase due to blow-back to the intake pipe pressure is calculated as an intake port pressure. When the blowback occurs, the intake port pressure estimating means uses a virtual volume determined according to the engine operating state based on a predetermined relationship, and the gas that flows backward to the intake port by the blowback is the virtual The intake air pressure estimation for an internal combustion engine according to claim 2, wherein the intake port pressure is calculated on the assumption that the pressure increase width in the space when flowing into the volume space is the pressure increase due to the blowback. apparatus. The virtual volume is further preliminarily determined in accordance with the amount of gas flowing into the space when the gas flowing back to the intake port due to the blow-back flows into the space of the virtual volume and the pressure increase width in the space. The intake air amount estimation device for an internal combustion engine according to claim 3, wherein the intake air amount estimation device changes based on a predetermined relationship. The said intake port pressure estimation means calculates the pressure increase part by the said blowback based on the pressure loss in the intake valve calculated based on the said intake valve passage intake flow volume, when the said blowback has arisen. An intake air amount estimation device for an internal combustion engine as described. The intake port pressure estimating means, when the blow-back occurs, a gas flow rate flowing into the intake pipe from the intake port calculated based on the intake port pressure and the intake pipe pressure, the intake valve passing intake flow rate, The intake air amount estimation device for an internal combustion engine according to claim 2, wherein an intake port pressure increase due to the blowback is calculated based on The intake port pressure estimating means, when the blow-back occurs, an amplitude and a period determined based on a predetermined relationship according to an engine operating state for the intake port pressure increase after the intake valve is opened. The intake air amount estimation device for an internal combustion engine according to claim 2, wherein the intake air amount is calculated as a sine function starting at the start of intake valve opening.

Images