JP2004278359A - Apparatus for estimating intake air amount of internal combustion engine - Google Patents

Abstract

To accurately estimate an in-cylinder charged intake air amount of an engine.
An intake flow rate filled in a cylinder using a throttle valve model (M10), an intake pipe model (M20), an intake valve model (M30), a cylinder model (M40), an exhaust valve model (M50), and the like, modeling each element of an intake system of an internal combustion engine. When estimating mc, if a blowback occurs in which the burned gas in the cylinder flows back to the intake port when the intake valve is opened, the intake port pressure is determined separately from the intake pipe pressure using the intake port model M25. Is calculated. Thus, an accurate blowback gas amount is calculated in consideration of the influence of the local pressure rise of the intake port caused by the blowback, so that the accuracy of estimating the in-cylinder charged intake air amount is improved.
[Selection] Fig. 8

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／Ａ^２、Ｃ＝ｃ／Ａ^２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ｕ^２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−（ρｕ^２／２）−α（ｌ／ｄ）ｕ^２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、Ｃ_１からＣ_４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]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to an intake air amount estimating apparatus for an internal combustion engine, and more particularly, to an intake air amount estimating apparatus that calculates an intake air amount of an internal combustion engine by a calculation formula modeling an intake system of the internal combustion engine.
[0002]
[Prior art]
The intake system of the internal combustion engine is divided into elements such as a throttle valve, an intake pipe, an intake valve, a cylinder, and an exhaust valve, and each element is modeled and expressed by a mathematical expression. There is known an intake air amount estimating apparatus for an internal combustion engine using a so-called air model, in which an intake air amount (in-cylinder charged intake air amount) of an engine is calculated by calculating the relationship by using a so-called air model.
[0003]
In the intake air amount estimating apparatus using such an air model, it is usually possible to calculate the in-cylinder charged intake air amount based only on the engine speed and the throttle valve opening other than the atmospheric pressure and the atmospheric temperature. In the intake air amount estimating device using such an air model, since the in-cylinder intake air amount can be calculated by calculation using the model formula, the responsiveness can be improved even during a transient operation in which the throttle valve opening change rate is large. It is possible to calculate the intake air amount.
[0004]
An example of a control device for an internal combustion engine that performs this type of intake air amount estimation is disclosed in Patent Document 1, for example.
The device disclosed in Patent Document 1 divides an intake system of an engine into each element of a throttle valve, an intake pipe including a surge tank, and an intake valve, and represents these elements by a simulation model, and furthermore, a pressure of an intake flow in each model. , Temperature, and flow rate are calculated using physical laws such as the law of conservation of energy, the law of conservation of mass, and the law of conservation of momentum. The in-cylinder charged intake air amount is calculated based on the intake air flow passing through the intake valve calculated as described above.
[0005]
In the device of Patent Document 1, when the intake valve is closed, it is assumed that the in-cylinder pressure is equal to the intake pipe pressure (surge tank pressure) due to the intake air charged into the cylinder, and the in-cylinder charged intake air amount (that is, the intake air amount). The in-cylinder intake amount when the valve is closed) is determined as a value proportional to the pressure in the intake pipe (surge tank) when the intake valve is closed.
However, in an actual engine, when the intake valve is opened, gas in the cylinder may flow backward from the intake valve to the intake port due to the pressure in the cylinder depending on the operating conditions of the engine. Since the device does not consider this blowback, there is a problem that the calculation accuracy of the in-cylinder charged intake air amount is reduced.
[0006]
In order to solve this problem, the present applicant has already proposed, in Japanese Patent Application No. 2001-316350, a device for estimating the intake air amount of an internal combustion engine using an air model in which the blowback is considered.
In the device disclosed in Japanese Patent Application No. 2001-316350, the intake system of the engine is modeled by dividing it into elements of a throttle valve, an intake pipe including a surge tank, and an intake valve. The air model which modeled is used. In the device of Japanese Patent Application No. 2001-316350, the throttle valve and the intake pipe use the same models as those of Patent Document 1, but in addition to these, an exhaust valve, a cylinder, and an intake valve are modeled, and By calculating the temperature, pressure, and flow rate of the intake air using the above-described physical laws, the in-cylinder charged intake air amount is calculated in consideration of the blowback at the time of opening the intake valve.
[0007]
That is, in the device of Japanese Patent Application No. 2001-316350, the pressure in the cylinder is calculated based on the relationship between the models and physical laws such as the law of conservation of mass and the law of conservation of energy. The intake flow rate passing through the intake valve is calculated based on the intake flow rate. This flow rate is a positive flow rate from the intake pipe toward 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, a blowback occurs. In this case, it becomes a negative value. The in-cylinder charged intake air amount is obtained by integrating the intake air passing through the intake valve 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 Japanese Patent Application No. 2001-316350, the intake valve passing intake flow rate was calculated based only 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 is significantly reduced under operating conditions where the valve overlap is large, that is, under the condition where the amount of blowback is large.
[0010]
In the apparatus disclosed in Japanese Patent Application No. 2001-316350, when a blowback occurs, the calculated intake pipe pressure increases in accordance with the blowback flow rate. However, since the intake pipe volume including the surge tank is larger than the blowback flow rate, the calculated intake pipe pressure is generated by the blowback. The increase in the intake pipe pressure is relatively small. For this reason, the difference between the cylinder internal pressure and the intake pipe pressure at the time of blowback is a relatively large value, and the amount of gas flowing out of the cylinder due to blowback is relatively large. Thus, when a large amount of blown back gas flows into the intake pipe, fresh air flows into the cylinder even when the cylinder pressure decreases and the flow of intake air from the intake port into the cylinder is restarted. Since the timing of the operation is delayed, the amount of fresh air charged into the cylinder is reduced.
[0011]
However, in an actual engine, the amount of gas flowing backward by blowing back when the intake valve is opened is smaller than the amount calculated by the air model of Japanese Patent Application No. 2001-316350, and the in-cylinder charged intake air amount (fresh air amount) is Is larger than the value calculated by
As a result of the research, it is considered that the difference in the blowback amount occurs because a sudden local pressure rise occurs in the intake port portion immediately after the intake valve is opened in the actual engine.
[0012]
In the case where a blowback occurs, particularly when the valve overlap is large, immediately after the intake valve opens, a high-speed blowback airflow near the sonic velocity flows into the intake port from the intake valve. In the Japanese Patent Application No. 2001-316350, the high-speed airflow is calculated assuming that the high-speed airflow flows into the intake pipe (surge tank) and the pressure in the surge tank uniformly increases. For this reason, the pressure rise in the surge tank is relatively small, and the difference between the surge tank and the cylinder internal pressure remains relatively large, so that the blowback flow rate is relatively large.
[0013]
However, since the flow velocity of the blowback airflow immediately after the opening of the intake valve is high, a sudden and local pressure rise 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 opening of the intake valve, the pressure difference between the intake port and the inside of the cylinder is reduced due to an increase in the pressure inside the intake port, whereby the blowback from the cylinder to the intake port is suppressed.
[0014]
That is, the blowback flow rate is actually reduced due to a local pressure increase in the intake port immediately after the intake valve is opened.
For this reason, when the blowback flow rate is calculated based on the intake pipe pressure (surge tank pressure) and the cylinder pressure as in the device of Japanese Patent Application No. 2001-316350, the valve overlap in which the blowback flow rate becomes large is particularly large. When the operating condition is large, there arises 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 calculates the flow rate of blow-back from the cylinder to the intake port accurately by taking into account the pressure of the intake port, and improves the calculation accuracy of the in-cylinder charged intake air amount. It is intended to provide an air amount estimating device.
[0016]
[Means for Solving the Problems]
According to the first aspect of the present invention, the intake valve passing 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 passing intake flow rate is calculated. An intake air amount estimating device for an internal combustion engine that calculates an in-cylinder charged intake air amount using an intake port pressure that calculates a cylinder intake port pressure based on a predetermined intake port model calculation formula using the intake valve passing intake air flow rate. An intake air amount estimating apparatus for an internal combustion engine, comprising estimating means for estimating an in-cylinder charged intake air amount based on the estimated intake port pressure and the in-cylinder pressure is provided.
[0017]
That is, in the first aspect of the present invention, the intake port pressure estimating means for estimating the intake port pressure separately from the intake pipe pressure using the intake flow rate passing through the intake valve is provided. Also 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 pressure. For this reason, it is possible to accurately estimate the in-cylinder charged intake air amount even during an operation with a large valve overlap where the amount of the blowback gas is large.
[0018]
According to the second aspect of the present invention, the intake port pressure estimating means is equal to the intake pipe pressure when the residual gas in the cylinder does not flow back 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 of the intake air pressure is calculated as an intake port pressure, and when a blowback occurs, a pressure obtained by adding a pressure increase due to the blowback to the intake pipe pressure is calculated as the intake port pressure. A quantity estimating device is provided.
[0019]
In other words, according to the second aspect of the present 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 rise due to the blowback when the blowback occurs. As a result, even in the case where a blowback occurs, the amount of blowback is suppressed as compared with the conventional case where the intake pipe pressure and the intake port pressure are assumed to be equal, and the in-cylinder intake air amount is accurately estimated. It is possible to do.
[0020]
Further, in a normal intake flow in which no blowback occurs, the speed of the intake flow flowing into the cylinder through the intake port is much lower than that of the blowback airflow. Therefore, a large pressure difference does not occur between the intake pipe (surge tank) and the intake port. Therefore, when no blowback occurs, the in-cylinder charged intake air amount can be calculated using the same value as the intake pipe pressure as the intake port pressure without having to separately calculate the intake port pressure.
[0021]
According to the invention described in claim 3, the intake port pressure estimating means uses a virtual volume determined according to an engine operating state based on a predetermined relationship when the blowback occurs, The intake port pressure is calculated by assuming that a pressure rise width in the space when the gas flowing back to the intake port by the blowback flows into the space of the virtual volume is a pressure rise amount due to the blowback. 2. An intake air amount estimating device for an internal combustion engine according to item 2 is provided.
[0022]
In other words, in the third aspect of the invention, the intake port is assumed to be a closed volume having a volume determined according to the operating state of the engine, and when a blow-back occurs, the blow-back gas flows into the closed volume to increase the pressure of the intake port. Is assumed to occur. This makes it possible to calculate the pressure rise in the intake port in accordance with the amount of gas that is blown back, and to accurately calculate the in-cylinder charged intake air amount.
[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 due to the blowback flows into the space of the virtual volume. The intake air amount estimating apparatus for an internal combustion engine according to claim 3, wherein the apparatus changes based on a predetermined relationship according to a pressure rise width in the space.
[0024]
That is, in the fourth aspect of the invention, as in the case of the third aspect, when a blowback occurs, the intake port is assumed to be a closed volume portion, and the pressure of the intake port is changed according to the amount of the blowback gas flowing into this volume portion. Calculate the rise. However, in the invention of claim 4, it is assumed that the volume of the closed volume is not constant but can be expanded according to the gas pressure inside, that is, the virtual volume of the intake port is changed in accordance with the rise in the pressure of the intake port. The intake port pressure is calculated assuming that the pressure changes. For example, the expansion of a volume can be assumed to mean that the gas inside the volume expands the volume against resistance such as elastic resistance (spring) and frictional resistance (dashpot). In this way, by modeling the intake port as a variable volume part, it is possible to more accurately calculate the blown-back gas amount, so that it is possible to more accurately calculate the in-cylinder charged intake air amount. Become.
[0025]
According to the invention described in claim 5, the intake port pressure estimating means, when the blowback occurs, performs the blowback based on the pressure loss in the intake valve calculated based on the intake valve passing intake flow rate. An intake air amount estimating device for an internal combustion engine according to claim 2, which calculates a pressure rise.
[0026]
That is, in the fifth aspect of the present invention, the pressure drop at the intake valve (the virtual flow path including the intake valve and the intake port portion near the intake valve) based on the intake flow rate passing through the intake valve at the time of blowback (that is, the backflow flow rate due to blowback). Pressure loss) is calculated, and the intake port pressure rise 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 the invention described in claim 6, 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 blowback occurs. 3. The intake air amount estimating device for an internal combustion engine according to claim 2, wherein an intake port pressure rise due to the blowback is calculated based on a gas flow rate and the intake valve passing intake flow rate.
[0028]
That is, according to the sixth aspect of the present invention, the amount of gas flowing into the intake port when the intake port pressure is generated (that is, the flow rate of the return flow that flows back through the intake valve to the intake port) and the amount of gas that flows out of the intake port when the blowback occurs. (The amount of gas flowing backward from the intake port to the surge tank) and the intake port pressure is calculated based on the law of conservation of mass. The amount of gas flowing into the intake port and the amount of gas flowing out of the intake port can be obtained, for example, as a gas flow passing through a throttle, based on the intake port pressure and the cylinder pressure, and the intake port pressure and the intake pipe pressure, respectively. it can. Thus, the intake port pressure is accurately calculated, and the calculation accuracy of the in-cylinder charged intake air amount is improved.
[0029]
According to the invention described in claim 7, the intake port pressure estimating means determines in advance the intake port pressure increase after opening the intake valve in accordance with the engine operating state when the blowback occurs. The intake air amount estimating device for an internal combustion engine according to claim 2, which has an amplitude and a period determined based on the relationship, and calculates the intake air amount of the internal combustion engine 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 temporal change of the intake port pressure rise after the opening of the intake valve when the blowback occurs is a sine function, and the amplitude and the period of this function are determined based on, for example, experimental values. 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 actually measured value, an accurate blowback flow rate closer to the actual value is calculated, and the calculation accuracy of the in-cylinder charged intake air amount is improved.
[0031]
BEST MODE FOR CARRYING OUT 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 in which the intake air amount estimating apparatus of the present invention is applied to a spark ignition type engine.
[0032]
In FIG. 1, reference numeral 10 denotes a spark ignition type internal combustion engine main 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. A combustion chamber 25 is formed above the piston 22 in the cylinder 21.
The cylinder head above the cylinder 21 has an intake port 31 communicating with the combustion chamber 25, an intake valve 32 for opening and closing the intake port 31, a phase angle of an intake camshaft for driving 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 the pressure. Reference numeral 33a in FIG. 1 denotes an actuator of the intake valve control device 33. The intake port 31 is provided with a fuel injection valve 39 for injecting fuel into the intake port.
[0034]
An exhaust port 34 that communicates with the combustion chamber 25, an exhaust valve 35 that opens and closes the exhaust port, and an exhaust camshaft 36 that opens and closes the exhaust valve 35 are provided in the cylinder head above the cylinder 21 in the same manner as 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]
An 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 an inlet of the intake pipe 41. , A throttle valve 43 disposed upstream of the surge tank 44 of the intake pipe 41.
[0036]
In the present embodiment, the throttle valve 43 is provided with an actuator 43a of an appropriate type such as a stepper motor or the like, and is a so-called electronic throttle valve which can take an opening independent of the operation of the accelerator pedal 82 by the driver. I have.
[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 air temperature sensor 62 for detecting an air temperature and an atmospheric pressure sensor 63 for detecting an atmospheric pressure are provided near the air flow meter 61.
[0038]
The exhaust system 50 of the engine 10 includes an exhaust pipe 52 connected to an exhaust port 34 of each cylinder via an exhaust manifold 51. Reference numeral 53 in FIG. 1 denotes a catalytic converter containing an exhaust purification catalyst.
Reference numeral 80 in FIG. 1 denotes an electronic control unit (ECU) that controls the engine 10. The ECU 80 is a microcomputer having a well-known configuration in which a CPU 81, a ROM 82, a RAM 83, an input / output interface 85, a backup RAM 84 capable of holding 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, an accurate intake air amount estimation taking into account blowback is performed based on an air model described later.
[0039]
For these controls, a signal corresponding to the amount of intake air (weight flow rate) flowing from the hot wire air flow meter 61 to the throttle valve 43 is sent to the input / output interface 85 of the ECU 80, and a signal from the intake air temperature sensor 62 and the atmospheric pressure sensor 63. And a signal corresponding to the intake air temperature (atmospheric temperature) and the atmospheric pressure, respectively. The outputs of 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 according to the air temperature and the atmospheric pressure.
Although not described in detail, in the present embodiment, the intake air amount detected by the air flow meter 61 is used to correct an intake air amount calculated using an air model described later.
[0040]
An input / output interface 85 of the ECU 80 outputs a voltage signal corresponding to the opening of the throttle valve 43 from a throttle valve opening sensor 64 provided near the throttle valve 43 to 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 arranged in the surge tank 44.
[0041]
Further, in the present embodiment, the intake valve lift sensor 67 is provided in the intake valve control device 33 and the O ₂ A sensor 70 is provided in the cooling water passage of the engine 10 with a cooling water temperature sensor 69, and a lift amount of the intake valve 32, an exhaust oxygen concentration, and an engine cooling water temperature are input to an input / output interface 85 of the ECU 80. are 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 a fixed angle (for example, 10 °). At the same time, each time the crankshaft 24 takes the reference rotational position (for example, every time the first cylinder reaches the top dead center), it outputs a reference crank position pulse signal having a pulse width wider than the crank rotation angle pulse signal.
[0043]
These signals are input to the interface 85 of the ECU 80. The ECU 80 calculates, for example, the engine rotation speed (rotation speed) 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]
Further, in this embodiment, an exhaust temperature sensor and an exhaust pressure sensor for detecting the exhaust pressure Pe and the temperature Te of the exhaust port are provided at the exhaust port of the cylinder (not shown), 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 air amount Mc charged into 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 by using an air model calculation formula that models the intake system 40 of the engine 10.
[0047]
In the present embodiment, basically, a method of calculating the in-cylinder charged intake air amount based on the air model proposed by the present applicant in Japanese Patent Application No. 2001-316350 is used. Therefore, in order to describe the method for calculating the cylinder charge intake amount Mc of the present embodiment, first, the method for calculating the cylinder charge intake amount Mc in Japanese Patent Application No. 2001-316350 will be briefly described.
[0048]
In the present embodiment, the intake system 40 of the engine 10 is disassembled into five elements, namely, a throttle valve, an intake pipe including a surge tank, an intake valve, a cylinder, and an exhaust valve, and each element is represented by a simulation model. By formulating the pressure, temperature, and flow rate of the flowing intake air, the intake air amount (in-cylinder charged intake air amount) Mc finally charged into the cylinder is calculated.
[0049]
Hereinafter, each model will be described.
(1) Throttle valve model
The throttle valve model is basically given as a flow of gas passing through a throttle (a 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, the gas flow rate mt passing through the throttle valve 43 is, as is well known, the gas pressure and temperature before and after the throttle valve, and the opening area of the throttle valve 43. And is expressed by the following formula 1 or 2 of the throttle valve model calculation formula.
[0051]
(Equation 1)

[0052]
(Equation 2)

[0053]
Here, Equation 1 represents the flow rate of intake air passing through the throttle valve 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 flow rates of the air passing through the throttle valve 43 when the pressure on the upstream side of the throttle valve 43 is lower than the pressure on the downstream side and backflow occurs are 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 in the present embodiment, and the atmospheric pressure, and Pm is the pressure on the downstream side of the throttle valve, and represents the intake pipe pressure in the present embodiment. Further, Ta and Tm are the intake air temperature on the upstream side and the downstream side of the throttle valve, respectively. In this embodiment, the atmospheric temperature and the intake pipe temperature are respectively shown, 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 Expressions 1 and 2, the throttle valve model of the present embodiment receives 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 as inputs, and passes through the throttle valve. It is a model calculation formula that uses 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 based on the mass flow rate and the energy conservation law based on the flow rate of the intake air flowing into the space having a constant volume and the flow rate of the intake air flowing out.
[0057]
Assuming that the volume of the intake pipe 41 including the surge tank 44 from the throttle valve 43 to the intake valve 32 is Vm, the intake air flowing into the volume Vm is the intake air passing through the throttle valve as shown in FIG. The flow rate is mt, and the intake air temperature is Ta.
[0058]
The flow rate of the intake air flowing out of the volume Vm is the flow rate mc of the intake air that passes through the intake valve and fills the cylinder, and the temperature of the intake flow is the intake pipe temperature Tm.
[0059]
Therefore, the following equation 3 is derived from the law of conservation of mass, and the following equation 4 is derived from the law of conservation of energy.
[0060]
[Equation 3]

[0061]
(Equation 4)

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

[0064]
(Equation 6)

[0065]
Here, the subscript k represents the value at the time of executing the current calculation, k-1 represents the value at the time of executing the previous calculation, and Δt represents the 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 of the in-cylinder intake flow rate mc described later over one cycle of the intake stroke.
[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 an intake flow passing through the intake valve 32 as shown in FIG. 4, and is basically expressed as an intake flow passing through a throttle (intake valve opening area) similarly to the throttle valve model. You.
[0067]
Therefore, the intake valve model is expressed by the following equations 7 and 8 similarly to the above-described throttle valve model (equations 1 and 2).
[0068]
(Equation 7)

[0069]
(Equation 8)

[0070]
Here, mc is the flow rate of intake air passing through the intake valve, that is, the amount of intake air charged into the cylinder in a unit time, Pm and Tm are intake pipe pressure and temperature, and Pc and Tc are cylinder pressure and temperature. 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, which is also a function of the lift amount L.
[0071]
In the present embodiment, the value of the product value Civ (L) · Aiv (L) is calculated for each value of L and stored in the ROM 82 of the ECU 80 in the form of a numerical table.
Similar to the throttle valve model (Equations 1 and 2), Equation 7 shows the case where the intake pipe pressure Pm is larger than the cylinder pressure Pc and the intake air flows into the cylinder, and Equation 8 shows that Pm is smaller than Pc. , Respectively, represents a case where the intake air flows backward to the intake pipe.
[0072]
As shown in Expressions 7 and 8, the intake valve model of the present embodiment uses the intake pipe pressure Pm, the cylinder pressure Pc, the intake pipe temperature Tm or the cylinder temperature Tc, and the intake valve lift amount L as inputs, and The valve passage intake flow rate mc is output.
The average intake air flow rate mcAVE used in the aforementioned intake pipe model calculation formulas (5) and (6) is obtained by integrating the value Mc integrated over the entire intake stroke calculated by the formulas (7) and (8) into the crank angle. This is a value obtained by dividing by a time T (180 CA) corresponding to 180 °.
[0073]
That is, mcAVE = Mc / T (180 CA)
(4) Cylinder model
In the cylinder model according to the present embodiment, as shown in FIG. 5, basically, intake air having a flow rate mc and a temperature Tm per unit time flows from an intake valve into a closed space having a volume Vc (variable) (Vc is a function of a crank angle). In consideration of the flow of exhaust gas flowing from the exhaust port through the exhaust valve and flowing back into the cylinder during the opening of the intake valve, the flow rate me and the temperature Te other than the intake flow from the intake valve are considered. It is given as a model into which the exhaust flows.
[0074]
That is, in the cylinder model of the present embodiment, mc, me, Tm, Te (and crank angle) are input, and the cylinder pressure Pc, the temperature Tc, and the total amount Mc of the intake air charged into the cylinder during the intake stroke are output. I do.
[0075]
The cylinder model is basically expressed by the following equation 9 based on the law of conservation of energy.
[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, and the like.
[0078]
The cylinder temperature Tc is expressed by the following equation 10 from the equation of state of gas.
[0079]
(Equation 10)

[0080]
Here, Mc1 is a charged intake air amount from the opening of the intake valve and is given as an integral value of the flow rate (mc + me) of the inflowing gas from the intake valve and the exhaust valve from the opening of the intake valve to the present time (at the time of calculating Tc). Can be
[0081]
Since the equation (9) cannot be operated by the ECU 80 in this state, it is converted into a discretized form and used in the following equation (11).
[0082]
[Equation 11]

[0083]
It should be noted that Δt in Equation 11 is a difference time when discretizing Equation 9 (Δt is set shorter in the cylinder model than in the case of the intake pipe model, and is set to, for example, 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)
The exhaust valve model is, like the intake valve model, a model that outputs the exhaust valve passing flow rate me with the input of the exhaust port pressure and temperature, Pe, Te, the cylinder pressure, the temperature Pc, 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]
(Equation 12)

[0087]
(Equation 13)

Equation 12 is for the case where exhaust gas flows backward from the exhaust port into the cylinder (ie, when the cylinder pressure Pc is lower than the exhaust port pressure Pe), and Equation 13 is for the case where exhaust gas flows from the cylinder to the exhaust port (where Pc is higher than Pe) Case) are shown. Note that Cev (L) and Aev (L) are the flow coefficient and the opening area of the exhaust valve 35, respectively, and both are functions of the lift amount L of the exhaust valve.
[0088]
In the present 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 in the exhaust port. Instead of detecting the exhaust pressure with a pressure sensor, it is also possible to perform the calculation on the assumption that Pe is approximately equal to the atmospheric pressure.
Similarly, since the exhaust temperature Te is substantially equal to the gas temperature in the cylinder, the exhaust temperature Te may be substituted by the cylinder temperature Tc without providing an exhaust temperature sensor.
[0089]
FIG. 7 is a block diagram showing an entire air model including the above-described throttle valve model, intake pipe model, intake valve model, cylinder model, and exhaust valve model. With reference to FIG. 6, the model calculation formula of each model element, the input and output, the relationship between each model element, 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) Formula
[0091]
[Equation 14]

[0092]
(Equation 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 (when number 15 (at the time of backflow))
(3) Output
・ Intake flow rate mt passing through throttle valve
(B) Intake pipe model (Fig. 7, M20)
(1) Formula
[0094]
(Equation 16)

[0095]
[Equation 17]

[0096]
(Equation 18)

[0097]
(2) Input
・ Throttle valve passing intake flow rate mt (k-1)
-Intake valve passing 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) Formula
[0098]
[Equation 19]

[0099]
(Equation 20)

[0100]
(2) Input
・ Intake pipe pressure Pm
・ Intake pipe temperature Tm
・ Cylinder pressure Pc
・ Cylinder temperature Tc (in the case of Equation 20 (at the time of backflow))
-Valve lift L (calculated from crank angle)
(3) Output
Intake valve passing intake flow rate (instantaneous value) mc
・ Intake valve passage intake flow rate mcAVE (average value)
(D) Cylinder model (Fig. 7, M40)
(1) Formula
[0101]
(Equation 21)

[0102]
(Equation 22)

[0103]
(2) Input
・ Cylinder volume Vc (calculated from crank angle)
・ Amount of intake air passing through intake valve (instantaneous value) mc
・ Tm in the intake pipe
・ Amount of intake air passing through exhaust valve me
・ Exhaust temperature Te
(3) Output
・ Cylinder pressure Pc
・ Cylinder temperature Tc
-In-cylinder charged intake air amount Mc
(E) Exhaust valve model (Fig. 7, M50)
(1) Formula
[0104]
[Equation 23]

[0105]
(Equation 24)

[0106]
(2) Input
・ Exhaust pipe pressure Pe
・ Exhaust pipe temperature Te
・ Cylinder pressure Pc
・ Cylinder temperature Tc (Equation 24)
-Valve lift L (calculated from crank angle)
(3) Output
Exhaust valve passing intake air flow (instantaneous value) me
As described above, the input values of the models other than the throttle valve model and the exhaust valve model, that is, the intake pipe model, the intake valve model, and the cylinder model can use the output values of all other models except the crank angle. Further, the crank angle can be calculated as a function of the engine speed NE.
[0107]
Further, as the input of the throttle valve model, output values of other models can be used 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.
Therefore, in the air model shown in FIG. 7, it is basically possible to calculate the in-cylinder charged intake air amount mc if 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, Tm and Tc are set to the atmospheric temperature Ta, 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 equations 14 to 24 (Pe and Te use actual measurement values).
Then, using the values of Pm, PcTm, Tc, mt, mc, me, etc. obtained as a result, the next calculation of Expression 14 to Expression 24 is performed. Then, this calculation is performed at predetermined intervals (for example, about 8 ms) from the start of the engine using the values of θ (t) and NE at that time and the values of Pm, tm, mt, mc, me, etc. obtained in the previous calculation. Repeat. As a result, Pm, tm, mt, mc, etc. should converge to the actual values in a short time after the start of the calculation, and should match the actual in-cylinder charged intake air amount.
[0109]
In the above-described conventional air model, in the intake valve model (FIG. 4), the intake air amount is estimated in consideration of the case where the reverse flow from the cylinder to the intake port occurs (Equation 20). Even in the case where a large amount of blowback occurs, the in-cylinder charged intake air amount should be accurately estimated.
However, as described above, since the conventional air model does not actually consider a local pressure increase at the intake port at the time of blowback, the estimation accuracy of the in-cylinder charged intake air amount is particularly high when the blowback amount is large. There is a problem of degradation.
[0110]
In the present invention, for the conventional air model, the intake pipe and the intake port are separately modeled, and the intake port pressure at the time of blowback is calculated as a value different from the intake pipe pressure using the intake port model. As a result, the accuracy of estimating the in-cylinder charged intake air amount is improved.
FIG. 8 is a block diagram similar to FIG. 7 illustrating the overall configuration of the air model of the present embodiment.
[0111]
As shown in FIG. 8, the present embodiment is different from the first embodiment in that an intake port model M25 is provided between an intake pipe model M20 and an intake valve model M30.
Although details of the intake port model M25 will be described later, in the present embodiment, the intake port pressure Pp is calculated using the intake port model M25, and the intake valve passage flow rate mc is replaced with the following equation 25 instead of equation 20. 26. Here, Equation 25 indicates the case where the blow-back to the intake port has not occurred (that is, when Pp ≧ Pc), and Equation 26 indicates the case where the blow-back has occurred for the intake port (that is, Pp <Pc).
[0112]
(Equation 25)

[0113]
(Equation 26)

[0114]
Here, as described later, in the intake port model M25 of the present embodiment, when a blow-back from the cylinder occurs, the intake port pressure Pp is calculated as a pressure higher than Pm separately from the intake pipe pressure Pm. If no blowback 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 no blow-back from the cylinder occurs.
As a result, the local pressure rise of the intake port due to the blowback airflow at the start of opening of the intake valve is also accurately reflected on the flow rate passing through the intake valve. The charged 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 relatively short intervals of, for example, about 1 ms.
In the operation of FIG. 9, it is determined whether or not a blowback to the intake port has occurred. If the blowback has occurred, the intake port pressure in consideration of a local pressure increase of the intake port due to the blowback to the intake port. Is calculated based on the intake port pressure Pp, and the intake valve passage flow rate is calculated using the intake valve model calculation formula (26). If no blowback has occurred, the intake valve passing intake flow rate is calculated using Equation 25 with Pp = Pm and Tp = Tm. In this case, Equation 25 becomes the same equation as Equation 20.
[0117]
That is, in FIG. 9, in step 901, the intake valve passing intake air amount mc is calculated using the intake port pressure Pp calculated at the time of performing the previous operation.
Then, in step 903, it is determined whether or not blow-back 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 / absence of blowback is determined only by comparing the pressures, the pressure pulsation of the intake port may be reduced. An error due to the influence may occur. Therefore, in the present embodiment, the intake valve passage flow rate (Equation 25) when Pp ≧ Pc is positive, and the intake valve passage flow rate (Equation 26) when Pp <Pc is a negative value, the intake valve passage flow for several times is set. The flow rate mc is integrated, and when the integrated value becomes a negative value, it is determined that blowback from the cylinder to the intake port has occurred.
[0118]
In step 903, if a blowback has occurred, that is, if the integrated value of the intake valve passing 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 , The intake port pressure Pp is calculated.
If the integrated value of the intake valve passage flow rate mc is positive in step 903, that is, if no blowback occurs, the intake port pressure Pp is calculated using the above-described intake pipe model calculation formula (Formula 16 to Formula 18). The calculated intake pipe pressure Pm is set, and the current operation ends.
[0119]
Thus, at the next operation execution, the intake valve uses the intake port pressure calculated based on the intake port model calculation formula when blowback occurs, and uses the intake pipe pressure when blowback does not occur. Since the passing flow rate is calculated, an accurate intake valve passing flow rate in consideration of the effect of the blowback is calculated.
Therefore, by integrating the intake valve passage flow rate over one cycle of the intake stroke in consideration of the sign, the amount Mc of fresh air charged into the cylinder can be accurately calculated.
[0120]
Next, an 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 the present embodiment, when a blowback occurs, the intake port is assumed to be a closed space having a predetermined virtual volume, and the pressure rise when the entire amount of the blowback gas from the cylinder flows into this closed space is determined. The pressure obtained by adding the pressure increase to the intake pipe pressure is used as the intake port pressure Pp.
When a high-speed airflow from the cylinder flows into the intake port in a short time immediately after the opening of the intake valve due to the blowback, the pressure of the gas flowing into the intake port locally increases in the intake port. This local pressure increase is considered to increase as the flow rate of the flowing gas increases and as the flow rate increases.
[0122]
Therefore, in the present embodiment, it is assumed that the intake port is a closed space (container) having a virtual volume determined by the operation state of the engine, and the pressure rise caused by the blowback airflow flowing into this closed space is equal to the pressure rise of the intake port. Assume that
Now, assuming that the volume of the closed space of the intake port is Vp and the blow-back flow rate per unit time, that is, the intake valve passage flow rate at the time of blow-back is mc, the pressure change per unit time dPp / dt of the intake port Pp due to the inflow of the blow-back gas. Is expressed by the following equation 27 from the state equation.
[0123]
[Equation 27]

[0124]
Here, R is the gas constant of the blowback gas, and Tp is the 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 blowback is equal to the integral value of Equation 27 after the start of the blowback, and is expressed by the following Equation 28 if the change in Tc is ignored.
[0125]
[Equation 28]

[0126]
Here, the integral value of the intake valve passage flow rate mc is obtained by integrating the flow rate mc (blow-back 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. Each time the operation is performed, the ECU 80 reads the value based on the actual engine speed and the valve timing.
[0128]
For example, the higher the number of revolutions of the engine, the smaller the amount of gas blown back. Therefore, the virtual volume Vp is set to a smaller value as the rotation speed increases. Further, as the valve timing is advanced (the valve overlap is larger), the amount of the blowback gas increases, but the pressure rise of the intake port does not increase corresponding to the amount of the blowback gas.
Therefore, 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 fixed value determined according to the engine operating state (for example, the engine speed and the valve timing).
However, in an actual operation, the intake port is connected to the surge tank instead of the closed space, and the entire amount of the blowback flow flowing into the intake port does not always remain at the intake port. The inflowing gas becomes a pressure wave while compressing the gas present in the intake port and the intake pipe, and travels through the intake port toward the surge tank.
[0130]
Therefore, in the present 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 expands, a resistance represented by a spring S having a spring constant k and a dashpot D having a damping constant c is present at a boundary between the existing gas in the intake port and the existing gas. And the blowback gas expands against this resistance.
[0131]
Now, assuming that the flow rate of the blowback 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]
(Equation 29)

[0133]
Here, W is the work performed by the blowback gas flowing into the intake port, and this work is pushed against the virtual movable boundary wall 110 urged by the spring S and the dashpot D as shown in FIG. The following equation (30) is obtained when expressed as work to be performed.
[0134]
[Equation 30]

[0135]
Here, A is a port cross-sectional area (constant value).
By expanding the right side of Equation 30 and setting k / A = K and c / A = C, Equation 31 is obtained.
[0136]
[Equation 31]

[0137]
Therefore, Equation 32 is obtained by substituting Equation 31 into Equation 29.
[0138]
(Equation 32)

[0139]
Further, the following equation 33 holds between Vp, Pp, and mc from the state equation.
[0140]
[Equation 33]

[0141]
Further, the following Expression 34 is obtained by expressing Expression 32 in a discretized manner.
[0142]
(Equation 34)

[0143]
The intake port pressure Pp is calculated by sequentially calculating Equation 34 using the intake valve passage flow rate mc (blow-back flow rate) calculated in Equation 26 and Equation 33. The constants K and C in Expression 34 are obtained in advance as a function of the engine speed NE and the intake valve timing VT by experiments, and are stored in the ROM 82 of the ECU 80 as a numerical map using NE and VT. Each map is read from this map based on the actual engine speed and valve timing.
This makes it possible to more accurately calculate the intake port pressure rise due to the blowback, thereby improving the calculation accuracy of the in-cylinder charged intake air amount.
[0144]
(3) Third embodiment
In the above-described second embodiment, the intake port pressure Pp is calculated based on the work performed by the virtual movable boundary wall 110 in FIG. 11 and the law of conservation of energy of the gas flowing into the intake port. To simplify the calculation, consider the balance of the forces acting on both sides of the movable boundary wall 110 in FIG.
[0145]
An intake pipe pressure Pm and a force F by a spring S and a dashpot D are acting on the left side (intake pipe side) of the movable wall 110 in FIG. Since the port pressure Pp is acting, if ΔPp = Pp−Pm, ΔPp × A = F. Here, A is the area of the movable wall 110 (cross-sectional area of the intake port), F is the sum of the urging forces of the spring S and the dashpot D, and uses the spring constant k of the spring S and the damping constant c of the dashpot D. Is given by the following equation 35.
[0146]
(Equation 35)

[0147]
Using equation 35 and the above ΔPp × A = F, K = k / A ² , C = c / A ² And the number 36 is obtained.
[0148]
[Equation 36]

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

[0151]
[Equation 38]

[0152]
Since ΔPp is a local pressure increase width of the intake port due to the inflow of the blowback gas, the intake port pressure Pp is as follows.
Therefore, ΔPp is first obtained from Equation 38 using Pp = Pm and Vp = 0 as initial values when the intake valve is opened, and using the intake valve passage flow rate mc (blow-back flow rate) calculated in Equation 25, and Is substituted into Expression 36 to determine ΔPp, and further, Pp is determined from Pp = Pm + ΔPp. Then, using the value of Pp thus obtained, mc is calculated by Expression 25, and using this mc and its integrated value, Vp and dVp / dt are calculated from Expressions 37 and 38, and 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 at the time of blowback is determined by a pressure loss at the time of passage of a blowback airflow through a virtual pipeline including an intake valve.
[0154]
Here, assuming that the gas flow velocity when passing through the pipeline is u, the pressure loss (Pc-Pp) is u ² And is represented by the following equation 39.
[0155]
[Equation 39]

[0156]
Here, ρ in the first term on the right side of Expression 39 is the density of the blown-back gas and is a function of the cylinder pressure Pc and the temperature Tc.
[0157]
In addition, the second term on the right side of Expression 39 is expressed as Expression 40.
[0158]
(Equation 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 experiments using an actual engine, and are stored in the ROM 82 of the ECU 80.
[0160]
Pc and Tc use values calculated by the cylinder model (M40 in FIG. 7). Thus, the intake port pressure Pp is calculated by the following equation 41. From Equation 39, the intake port pressure rise ΔPp due to the blowback is ΔPp = Pp−Pm = Pc−Pm− (ρu ² / 2) -α (l / d) u ² It can also be expressed as
[0161]
(Equation 41)

[0162]
That is, in the present embodiment, ρ and u are calculated using mc, Pc, and Tc calculated using the cylinder model (M40 in FIG. 7) and the intake valve model (M30 in FIG. 7), and By repeating the operation of calculating the intake port pressure Pp and calculating the next mc, Pc, and Tc, a local increase in the intake port pressure due to the blowback is reflected on the intake valve passage flow rate mc, and the in-cylinder charged intake air amount is calculated. The accuracy is improved.
[0163]
(5) Fifth embodiment
Next, a fifth embodiment of the intake port model will be described.
FIG. 12 schematically shows the intake port model of the present embodiment. As shown in FIG. 12, in the present embodiment, it is assumed that a part of the blow-back gas flowing 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 from the intake port into the intake pipe through the throttle 120 are shown in FIG. Equation 42 is obtained from the equation of state of gas.
[0165]
(Equation 42)

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

[0168]
[Equation 44]

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

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

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

[0179]
[Equation 48]

[0180]
Where C ₁ To C ₄ Is a constant and is determined in advance 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. 2, curves VT40, VT20, and VT0 show the case where the intake valve opening crank angle (CAio in Equation 46) is 40 degrees, 20 degrees, and 0 degrees before the top dead center (BTDC), respectively.
Note that the amplitude a and the 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 being represented by equations as in Equations 47 and 48. It may be stored.
[0182]
In the present embodiment, it is possible to easily and accurately calculate the intake port pressure rise at the time of blowback by expressing the intake port pressure rise width ΔPp after opening the intake valve as a sine function in the form of Equation 46. 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 calculating the intake valve passing intake flow rate based on the intake valve model calculation formula and calculating the in-cylinder charged intake air amount using the calculated intake valve passing intake flow rate, By taking the port pressure into account, it is possible to accurately calculate the blow-back flow rate from the cylinder to the intake port, and to achieve the common effect of improving the calculation accuracy of the in-cylinder charged intake air amount.
[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 type 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 related art.
FIG. 8 is a block diagram illustrating an entire air model that the user wants to have an intake port model.
FIG. 9 is a flowchart illustrating 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 the intake port model.
FIG. 12 is a diagram illustrating a fifth embodiment of the intake port model.
FIG. 13 is a diagram showing an example of a change in intake port pressure rise calculated using a sixth embodiment of the intake port model.
[Explanation of symbols]
10: 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 that calculates an intake valve passing intake air flow rate based on a predetermined intake valve model calculation equation using an intake pipe pressure and a cylinder pressure, and calculates an in-cylinder charged intake air amount using the calculated intake valve passing intake flow rate. In the engine intake air amount estimation device,
An intake port pressure estimating means for calculating a cylinder intake port pressure based on an intake port model calculation formula predetermined using the intake valve passing intake flow rate,
An intake air amount estimating device for an internal combustion engine, which estimates an in-cylinder charged intake air amount based on the estimated intake port pressure and the in-cylinder pressure. The intake port pressure estimating means calculates the same value as the intake pipe pressure as the intake port 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. 2. The intake air amount estimating device for an internal combustion engine according to claim 1, wherein when the pressure is generated, a pressure obtained by adding the pressure increase due to the blowback to the intake pipe pressure is calculated as the intake port pressure. The intake port pressure estimating means uses a virtual volume determined according to an engine operating state based on a predetermined relationship when the blowback occurs, and the gas that flows back to the intake port due to the blowback is virtual. 3. An intake air amount estimation for an internal combustion engine according to claim 2, wherein an intake port pressure is calculated by assuming a pressure rise width in the space when it is assumed to have flowed into the space having a volume as a pressure rise due to the blowback. apparatus. The virtual volume is further determined in advance according to the amount of gas flowing into the space and the pressure increase width in the space when the gas that flows back to the intake port due to the blowback flows into the space of the virtual volume. 4. The intake air amount estimating device for an internal combustion engine according to claim 3, wherein the device varies based on a determined relationship. The intake port pressure estimating means calculates a pressure increase due to the blowback based on a pressure loss in an intake valve calculated based on the intake valve passing intake flow rate when the blowback occurs. An apparatus for estimating an intake air amount of an internal combustion engine according to the above. The intake port pressure estimating means, when the blowback occurs, the gas flow rate flowing into the intake pipe from the intake port calculated based on the intake port pressure and the intake pipe pressure, and the intake valve passing intake flow rate The intake air amount estimating device for an internal combustion engine according to claim 2, wherein an intake port pressure rise due to the blowback is calculated based on the following equation. The intake port pressure estimating means, when the blowback is occurring, the intake port pressure rise after opening the intake valve, the amplitude and cycle determined based on a predetermined relationship according to the engine operating state. 3. The intake air amount estimating device for an internal combustion engine according to claim 2, wherein the calculating unit calculates the sine function starting from the start of the intake valve opening.

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