JP3979149B2 - Driving force control device - Google Patents

Description

【００１８】
そして、加減速度モデルマッチング補償器４０６により、前記フィードフォワード制御用目標変速機入力トルク指令値Ｔ^* _p0から変速機入力トルク指令値のフィードフォワード制御分Ｔ^* _p-ffが得られる。加減速度モデルマッチング補償器４０６の伝達関数Ｇ_FF-1(s) を下記２式に示す。この加減速度モデルマッチング補償器４０６は、加減速度に関する規範モデルを用いて規範化する（式中の分母）と共に、出力の先方にある一次遅れ系の変速機入力トルク制御系、つまりエンジントルク制御系の応答遅れの逆数（式中の分子）を乗じて位相合わせを行っている。なお、式中のｓはラプラス演算子、τ_eng はエンジントルク制御系の応答遅れ時定数、ω_r 、ζ_r は目標加減速度α^* _p に対する加減速度α_p の規範モデル応答（二次遅れモデル）のカットオフ周波数とダンピング定数である。
【００１９】
【数２】

【００２０】
但し、前述したマイクロコンピュータで演算処理を行うためには、例えばタスティン近似等で離散化して、下記２ａ式で示すような、ソフトウエアで実行可能な差分方程式を求めて変速機入力トルク指令値のフィードフォワード制御分Ｔ^* _p-ffを算出する。なお、式中のMTN0、MTN1、MTN2、MTD1、MTD2は、前記時定数τ_eng 、カットオフ周波数ω_r 、ダンピング定数ζ_r 、演算処理のサンプリング周期ΔＴから決まる定数である。また、(k) は今回値、(k-1) は前回値、(k-2) は前々回値を示す。
【００２１】
【数３】

【００２２】
一方、前記目標駆動力Ｆ^* _w と車輪速度Ｖ_w とを乗算器４０７で乗じると目標エンジンパワー（出力）Ｐ^* が得られるので、目標変速機入力回転速度設定部４０８では前述したエンジン運転拘束マップを用いて当該目標エンジンパワーＰ^* を達成し且つ最適な燃費が得られる目標変速機入力回転速度ω^* _p を算出設定する。目標エンジンパワーＰ^* の算出式を下記３式に示す。
【００２３】
【数４】

【００２４】
従って、除算器４０９で、この目標変速機入力回転速度ω^* _p を前記車輪速度Ｖ_w で除し、更に乗算器４１０で、タイヤ有効半径Ｒを乗じ且つ最終減速比Ｉ_f で除して、フィードフォワード制御用目標変速比Ｉ^* _p0が得られる。フィードフォワード制御用目標変速比Ｉ^* _p0の算出式を下記４式に示す。
【００２５】
【数５】

【００２６】
そして、変速比モデルマッチング補償器４１１により、前記フィードフォワード制御用目標変速比Ｉ^* _p0から変速比指令値のフィードフォワード制御分Ｉ^* _p-ffが得られる。変速比モデルマッチング補償器４１１の伝達関数Ｇ_FF-2(s) を下記５式に示す。この変速比モデルマッチング補償器４１１は、変速比に関する規範モデルを用いて規範化する（式中の分母）と共に、出力の先方にある一次遅れ系の変速比制御系の応答遅れの逆数（式中の分子）を乗じて位相合わせを行っている。なお、式中のτ_cvt は変速比制御系の応答遅れ時定数、τ_ref-wpは目標変速機入力軸回転速度ω^* _p に対する変速機入力軸回転速度ω_p の規範モデル応答（一次遅れモデル）の時定数である。
【００２７】
【数６】

【００２８】
但し、前述したマイクロコンピュータで演算処理を行うためには、例えばタスティン近似等で離散化して、下記５ａ式で示すような、ソフトウエアで実行可能な差分方程式を求めて変速比指令値のフィードフォワード制御分Ｉ^* _p-ffを算出する。なお、式中のMIN0、MIN1、MID1は、前記時定数τ_cvt 、τ_ref-wp、演算処理のサンプリング周期ΔＴから決まる定数である。
【００２９】
【数７】

【００３０】
次に、前記エンジン運転拘束マップについて図５を用いて説明する。例えば、図のように横軸にエンジン回転数ω_e （＝変速機入力回転速度ω_p ）をとり、縦軸にエンジントルクＴ_e （＝変速機入力トルクＴ_p ）をとると、同等のエンジンパワー（出力）を結んだ等出力線（図では破線）や、最適燃費点を中心とする等燃料消費線（図では一点鎖線）が描ける。等出力線上の最適燃費点を連続した曲線が最適燃費運転線となる。一般に、昨今のエンジンでは、アクセルオフの状態で燃料を噴射しないので、最適燃費点や等燃費線はエンジントルクＴ_e が正の領域にのみ存在する。従って、最適燃費運転線もエンジントルクＴ_e が正の領域にしか存在しない。逆に、エンジントルクＴ_e が負の領域では、エンジンブレーキトルクとエンジン回転速度との関係を示すエンジンブレーキ特性線が表れる。前述のように、エンジントルクＴ_e が負の領域では燃料を噴射しないので、エンジンブレーキトルクを制御するためにはエンジン回転速度を制御する必要がある。本実施形態では、変速機に無段変速機を用いているので、任意の走行速度で所望するエンジンブレーキトルクを得るためには、無段変速機の変速比を制御すればよい。これらの曲線の関係を、燃費を考慮してマップ化したものがエンジン運転拘束マップである。
【００３１】
次に、本実施形態でのフィードバック補償器３３の設計手法を簡潔に説明する。前述した図３の非線形制御対象モデルを、変速機入力トルク指令値Ｔ^* _p 、変速比指令値Ｉ^* _p の二入力、加減速度α_w 、変速機入力軸回転速度差Δω_p の二出力の非線形制御対象モデルであると仮定する。本実施形態では、フィードバック制御系の安定性を確保する目的で、前記検出部や一部制御部を車両モデルと組合せて制御対象モデルとしている。この非線形制御対象モデルを、特定の動作点で線形近似を行って、制御系設計用の線形近似制御対象モデルを導出する。「ロバスト制御理論」の一つである「μシンセシス」を用いてフィードバック補償器を設計するためには、更に変動要素をモデル化して一般化プラントモデルに拡張する必要があるが、ここではその詳細は割愛する。
【００３２】
前記フィードバック補償器３３への入力は、前述のように加減速度差Δα_w 、変速機入力軸回転速度差（−Δω_p ）であるから、出力変速機入力トルク指令値Ｔ^* _p 、変速比指令値Ｉ^* _p を当該フィードバック補償器３３の伝達関数Ｇ_FB(s) で示すと下記６式となり、当該伝達関数Ｇ_FB(s) の各要素は７式で表れる。
【００３３】
【数８】

【００３４】
実際の車両諸元、或いは要求する応答特性を代入し、前記「μシンセシス」によって各要素Ｇ₁₁(s) 〜Ｇ₂₂(s) を求めると、下記８式〜１１式のように表れる。
【００３５】
【数９】

【００３６】
これら各要素Ｇ₁₁(s) 〜Ｇ₂₂(s) を子細に考察すると、極が虚軸上又はその近傍にある部分が存在する。この虚軸上又はその近傍にある極は、応答の遅い極であり、目標値と実際値との差を蓄積する、換言すれば積分的特性を持つ部分であるといえる。そこで、前記７式のフィードバック補償器の伝達関数を、積分的特性を有する部分Ｇ_A (s) と、それ以外の部分Ｇ_B (s) とに分離し、下記１２式のように表す。
【００３７】
【数１０】

【００３８】
具体的な要素Ｇ_11-A(s) 〜Ｇ_22-A(s) 、Ｇ_11-B(s) 〜Ｇ_22-B(s) は下記１３式から２０式で表れる。
【００３９】
【数１１】

【００４０】
そして、前記加減速度α_w 及び変速機入力軸回転速度差（−Δω_p ）に前記積分的特性を有さない要素Ｇ_11-B(s) 〜Ｇ_22-B(s) を施した要素をｘ₁₁〜ｘ₂₂とし、これらの要素に前記積分的特性を有する要素Ｇ_11-A(s) 〜Ｇ_22-A(s) をｙ₁₁〜ｙ₂₂とすると、前記変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fb、変速比指令値のフィードバック制御分Ｉ^* _p-fbは、夫々、下記２１式、２２式で表れる。
【００４１】
【数１２】

【００４２】
次に、前記フィードフォワード補償器３１及びフィードバック補償器３３で行われる演算処理について図６のフローチャートを用いて説明する。この演算処理は、例えば１０msec. 程度の所定サンプリング周期ΔＴで行われる。なお、この演算処理では、通信のための全てのステップを記載していないが、必要な情報は随時他のコントローラ或いは記憶装置と授受されるし、演算処理で得られた情報は随時他のコントローラ或いは記憶装置と授受される。
【００４３】
この演算処理では、まずステップＳ１で前記アクセルセンサ２４で検出されたアクセル開度Ａ_p0を読込む。
次にステップＳ２に移行して、前記車輪速度センサ２５で検出された車輪速度Ｖ_w を読込む。
次にステップＳ３に移行して、前記変速機コントローラ５からプライマリ回転速度ω_p 、セカンダリ回転速度ω_s 、両者の比である変速比Ｉ_p を読込むと共に、前記エンジントルクコントローラ７からエンジン回転速度ω_e を読込む。
【００４４】
次にステップＳ４に移行して、図７に示す制御マップに従って、前記ステップＳ１で読込んだアクセル開度Ａ_p0、前記ステップＳ２で読込んだ車輪速度Ｖ_w に基づいて目標加減速度α^* _w を算出設定する。
次にステップＳ５に移行して、下記２３式で示す伝達関数Ｇ_bp(s) のバンドパスフィルタを用い、車輪速度Ｖ_w からノイズを除去した、所定周波数領域のみの車輪加減速度α_w を算出する。なお、式中のω_n は固有角周波数、ζ_n は減衰率であり、ω_n 、ζ_n は、検出される車輪速度のノイズレベルによって決定される。
【００４５】
【数１３】

【００４６】
但し、前述したマイクロコンピュータで演算処理を行うためには、例えばタスティン近似等で離散化して、下記２３ａ式で示すような、ソフトウエアで実行可能な差分方程式を求めて車輪加減速度α_w を算出する。なお、式中のBPN0、BPN1、BPN2は、前記固有角周波数ω_n 、減衰率ζ_n 、サンプリング周期ΔＴによって決まる定数である。
【００４７】
【数１４】

【００４８】
次にステップＳ６に移行して、例えば前回演算時の変速機入力トルク指令値Ｔ^* _p からエンジントルク応答遅れモデルによる変速機入力トルクＴ_p を算出する。このエンジントルク応答遅れモデルは、下記２５式の伝達関数で示す一次遅れ系である。
【００４９】
【数１５】

【００５０】
但し、前述したマイクロコンピュータで演算処理を行うためには、例えばタスティン近似等で離散化して、下記２５ａ式で示すような、ソフトウエアで実行可能な差分方程式を求めて変速機入力トルクＴ_p を算出する。なお、式中のTEN0、TEN1、TEN2は、前記時定数τ_eng 、サンプリング周期ΔＴによって決まる定数である。
【００５１】
【数１６】

【００５２】
次にステップＳ７に移行して、道路勾配抵抗力Ｆ_gra を算出する。具体的には、前述したように、前記ステップＳ６で算出した変速機入力トルクＴ_p 、変速比Ｉ_p 、最終減速比Ｉ_f 、タイヤ有効半径Ｒを用いて、下記２６式に従って駆動力Ｆ_d を算出する。次いで、下記２７式に従って、前記ステップＳ５で算出した加減速度α_w と車両質量Ｍとの積から実駆動力Ｆ_realを算出する。次いで、下記２８式に従って、前記駆動力Ｆ_d から実駆動力Ｆ_realを減じて全走行抵抗力Ｆ_all を算出する。次に、下記２９式に従って、前記全走行抵抗力Ｆ_all から、前述のように車輪速度Ｖ_w に応じて得られる平坦路走行抵抗力Ｆ_r を減じてノイズ除去前道路勾配抵抗力Ｆ_gralを算出する。
【００５３】
【数１７】

【００５４】
次に、下記３０式伝達関数Ｆ₂(s)で示すローパスフィルタで不必要な高域ノイズ成分を除去して道路勾配抵抗力Ｆ_gra を算出する。なお、式中のτ_gra はローパスフィルタのカットオフ周波数である。
【００５５】
【数１８】

【００５６】
但し、前述したマイクロコンピュータで演算処理を行うためには、例えばタスティン近似等で離散化して、下記３１式で示すような、ソフトウエアで実行可能な差分方程式を求めて道路勾配抵抗力Ｆ_gra を算出する。なお、式中のSLPNO 、SLPN1 、SLPD1 は、前記カットオフ周波数τ_gra 、サンプリング周期ΔＴによって決まる定数である。
【００５７】
【数１９】

【００５８】
次にステップＳ８に移行して、図８に示す制御マップに従って、前記ステップＳ７で算出した道路勾配抵抗力Ｆ_gra から目標加減速度補正値α_gra を算出する。この制御マップは、上り坂で作用する道路勾配抵抗力Ｆ_gra を正値、下り坂で作用する道路勾配抵抗力Ｆ_gra を負値で表したとき、道路勾配抵抗力の絶対値｜Ｆ_gra ｜が比較的小さな第１所定値Ｆ_a1以下の領域では目標加減速度補正値α_gra は“０”であり、道路勾配抵抗力Ｆ_gra が比較的大きな第２所定値Ｆ_a2以上の領域では、目標加減速度補正値α_gra は、道路勾配による加減速度（ｇ・sin θ）と同等の傾きで且つそれより所定値α_Vmaxだけ小さくなるように設定されており、道路勾配抵抗力Ｆ_gra が絶対値の比較的大きな負値の第２所定値（−Ｆ_a2）以下の領域では、目標加減速度補正値α_gra は、道路勾配による加減速度（ｇ・sin θ）と同等の傾きで且つそれより所定値α_Vmaxだけ大きくなるように設定されており、所定値（−Ｆ_a2）から所定値（−Ｆ_a1）の間、或いは所定値Ｆ_a1から所定値Ｆ_a2の間では、目標加減速度補正値α_gra は、比較的小さな傾きで、道路勾配抵抗力のＦ_gra の増減と共にリニアに増減するように設定されている。なお、前記所定値α_Vmaxは、車両の最大駆動力を車両質量で除した値とした。
【００５９】
次にステップＳ９に移行して、前記目標加減速度α^* _w に前記ステップＳ８で算出された目標加減速度補正値α_gra を加算した値を、新たな目標加減速度α^* _w として補正する。
次にステップＳ１０に移行して、前述したエンジン運転拘束条件による規範変速機入力軸回転速度ω_p-ref と変速機入力軸回転速度ω_p との変速機入力軸回転数差Δω_p を算出する。具体的には、まず前記ステップＳ６で算出した変速機入力トルクＴ_p から前記最適燃費運転線又はエンジンブレーキ特性線上の変速機入力軸回転速度を目標変速機入力軸回転速度ω^* _p とし、これを下記３２式の伝達関数Ｇ_ref-wp(s) で示す規範モデル応答特性を用いて規範化し、規範変速機入力回転速度ω_p-ref を算出する。
【００６０】
【数２０】

【００６１】
但し、前述したマイクロコンピュータで演算処理を行うためには、例えばタスティン近似等で離散化して、下記３２ａ式で示すような、ソフトウエアで実行可能な差分方程式を求めて規範変速機入力回転速度ω_p-ref を算出する。なお、式中のPRN0、PRN1、PRD2は、前記時定数τ_ref-wp、サンプリング周期ΔＴによって決まる定数である。
【００６２】
【数２１】

【００６３】
そして、下記３３式で示すように、求めた規範変速機入力回転速度ω_p-ref から前記変速機入力軸回転速度ω_p を減じて変速機入力軸回転速度差Δω_p を算出する。
【００６４】
【数２２】

【００６５】
次にステップＳ１１に移行して、前記フィードフォワード補償器３１により、目標変速機入力トルク指令値のフィードフォワード制御分Ｔ^* _p-ff及び目標変速比指令値のフィードフォワード制御分Ｉ^* _p-ffを算出する。
次にステップＳ１２に移行して、フィードバック制御用規範加減速度α_w-ref と加減速度α_w との加減速度差Δα_w を算出する。具体的には、下記３４式で示す伝達関数Ｇ_ref-a からなる加減速度の規範モデル応答に相当する遅れ補償（二次遅れモデル）と、同じく伝達関数Ｇ_bp(s) からなる前記加減速度算出用バンドパスフィルタに相当する遅れ補償（二次遅れモデル）とを目標加減速度α^* _w に施して規範加減速度α_w-ref を算出する。
【００６６】
【数２３】

【００６７】
但し、前述したマイクロコンピュータで演算処理を行うためには、例えばタスティン近似等で離散化して、下記３４ａ式で示すような、ソフトウエアで実行可能な差分方程式を求めて規範加減速度α_w-ref を算出する。なお、式中のREN0、REN1、REN2、REN3、REN4、RED1、RED2、RED3、RED4は、前記カットオフ周波数ω_r 、ダンピング定数ζ_r 、固有角周波数ω_n 、減衰率ζ_n 、サンプリング周期ΔＴから決まる定数である。また、(k-4) は前々々回値を示す。
【００６８】
【数２４】

【００６９】
次にステップＳ１３に移行して、前述したように加減速度差Δα_w 、変速機入力軸回転速度差（−Δω_p ）に対し、前記積分的特性を除いたフィードバック補償器Ｇ_11-B(s) 〜Ｇ_22-B(s) による変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素ｘ₁₁、ｘ₁₂、及び変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素ｘ₂₁、ｘ₂₂を算出する。
【００７０】
次にステップＳ１４に移行して、後述する図１１の演算処理に従って、前記ステップＳ１３で算出された変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素ｘ₁₁、ｘ₁₂、及び変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素ｘ₂₁、ｘ₂₂に対し、前記積分的特性を有するフィードバック補償器Ｇ_11-A(s) 〜Ｇ_22-A(s) による変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素ｙ₁₁、ｙ₁₂、及び変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素ｙ₂₁、ｙ₂₂を算出する。
【００７１】
次にステップＳ１５に移行して、前記２１式、２２式に従って、前記ステップＳ１４で算出した要素ｙ₁₁、ｙ₁₂の加算値から変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbを算出すると共に、前記要素ｙ₂₁、ｙ₂₂の加算値から変速比指令値のフィードバック制御分Ｉ^* _p-fbを算出する。
次にステップＳ１６に移行して、下記３５式、３６式に従って、前記変速機入力トルク指令値のフィードフォワード制御分Ｔ^* _p-ffとフィードバック制御分Ｔ^* _p-fbとの加算値から変速機入力トルク指令値Ｔ^* _p を算出すると共に、変速比指令値のフィードフォワード制御分Ｉ^* _p-ffとフィードバック制御分Ｉ^* _p-fbとの加算値から変速比指令値Ｉ^* _p を算出する。
【００７２】
【数２５】

【００７３】
次にステップＳ１７に移行して、図９、図１０に示す制御マップから、前記変速機入力トルク指令値Ｔ^* _p 、変速比指令値Ｉ^* _p に制限処理を施す。つまり、目標値を、実際に発生可能な制御量の上下限値で制限する。
次にステップＳ１８に移行して、前記変速機入力トルク指令値Ｔ^* _p 、変速比指令値Ｉ^* _p を、夫々、前記エンジントルクコントローラ７、変速機コントローラ５に向けて出力してからメインプラグラムに復帰する。
【００７４】
次に、前記図６の演算処理のステップＳ１４で行われるマイナプログラムについて図１１のフローチャートに従って説明する。この演算処理では、まずステップＳ２１で、変速機入力トルク指令値の前回値Ｔ^* _p (k-1) が上限値で且つ前記加減速度差Δα_w が“０”以上であるか、又は変速機入力トルク指令値の前回値Ｔ^* _p (k-1) が下限値で且つ前記加減速度差Δα_w が“０”以下であるか否かを判定し、何れかの条件が満足される場合にはステップＳ２２に移行し、そうでない場合にはステップＳ２３に移行する。
【００７５】
前記ステップＳ２３では、前記変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素ｘ₁₁に対し、前記積分的特性を有するフィードバック補償器Ｇ_11-A(s) による変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素ｙ₁₁を算出してからステップＳ２４に移行する。具体的には、前述と同様に離散化して求めた差分方程式に基づいて要素の今回値ｙ₁₁(k) を更新する。
【００７６】
前記ステップＳ２２では、前記変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素の前回値ｙ₁₁(k-1) を今回値ｙ₁₁(k) として出力してから前記ステップＳ２４に移行する。実質的には、前記ステップＳ２３で説明した差分方程式の要素の今回値ｙ₁₁(k) を更新せず、前回値ｙ₁₁(k-1) のまま保存する。
前記ステップＳ２４では、変速機入力トルク指令値の前回値Ｔ^* _p (k-1) が上限値で且つ前記変速機入力軸回転速度差Δω_p が“０”以上であるか、又は変速機入力トルク指令値の前回値Ｔ^* _p (k-1) が下限値で且つ前記変速機入力軸回転速度差Δω_p が“０”以下であるか否かを判定し、何れかの条件が満足される場合にはステップＳ２５に移行し、そうでない場合にはステップＳ２６に移行する。
【００７７】
前記ステップＳ２６では、前記変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素ｘ₁₂に対し、前記積分的特性を有するフィードバック補償器Ｇ_12-A(s) による変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素ｙ₁₂を算出してからステップＳ２７に移行する。具体的には、前述と同様に離散化して求めた差分方程式に基づいて要素の今回値ｙ₁₂(k) を更新する。
【００７８】
前記ステップＳ２５では、前記変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fbの要素の前回値ｙ₁₂(k-1) を今回値ｙ₁₂(k) として出力してから前記ステップＳ２７に移行する。実質的には、前記ステップＳ２６で説明した差分方程式の要素の今回値ｙ₁₂(k) を更新せず、前回値ｙ₁₂(k-1) のまま保存する。
前記ステップＳ２７では、変速比指令値の前回値Ｉ^* _p (k-1) が上限値で且つ前記加減速度差Δα_w が所定値Δα_w1以上であるか、又は変速比指令値の前回値Ｉ^* _p (k-1) が下限値で且つ前記加減速度差Δα_w が所定値Δα_w2以下であるか否かを判定し、何れかの条件が満足される場合にはステップＳ２８に移行し、そうでない場合にはステップＳ２９に移行する。なお、前記所定値Δα_w1は正値であって加速指令を意味し、前記所定値Δα_w2は負値であって減速指令を意味する。
【００７９】
前記ステップＳ２９では、前記変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素ｘ₂₁に対し、前記積分的特性を有するフィードバック補償器Ｇ_21-A(s) による変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素ｙ₂₁を算出してからステップＳ３０に移行する。具体的には、前述と同様に離散化して求めた差分方程式に基づいて要素の今回値ｙ₂₁(k) を更新する。
【００８０】
前記ステップＳ２８では、前記変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素の前回値ｙ₂₁(k-1) を今回値ｙ₂₁(k) として出力してから前記ステップＳ３０に移行する。実質的には、前記ステップＳ２９で説明した差分方程式の要素の今回値ｙ₂₁(k) を更新せず、前回値ｙ₂₁(k-1) のまま保存する。
前記ステップＳ３０では、変速比指令値の前回値Ｉ^* _p (k-1) が上限値で且つ前記変速機入力軸回転速度差Δω_p が“０”以上であるか、又は変速比指令値の前回値Ｉ^* _p (k-1) が下限値で且つ変速機入力軸回転速度差Δω_p が“０”以下であるか否かを判定し、何れかの条件が満足される場合にはステップＳ３１に移行し、そうでない場合にはステップＳ３２に移行する。
【００８１】
前記ステップＳ３２では、前記変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素ｘ₂₂に対し、前記積分的特性を有するフィードバック補償器Ｇ_22-A(s) による変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素ｙ₂₂を算出してから前記図６の演算処理のステップＳ１５に移行する。具体的には、前述と同様に離散化して求めた差分方程式に基づいて要素の今回値ｙ₂₂(k) を更新する。
【００８２】
前記ステップＳ３１では、前記変速比指令値のフィードバック制御分Ｉ^* _p-fbの要素の前回値ｙ₂₂(k-1) を今回値ｙ₂₂(k) として出力してから前記図６の演算処理のステップＳ１５に移行する。実質的には、前記ステップＳ３２で説明した差分方程式の要素の今回値ｙ₂₂(k) を更新せず、前回値ｙ₂₂(k-1) のまま保存する。
【００８３】
これらの演算処理によれば、目標加減速度α^* _w に応じた目標変速機入力トルク指令値のフィードフォワード制御分Ｔ^* _p-ff及び目標変速比指令値のフィードフォワード制御分Ｉ^* _p-ffが算出設定され、同じく加減速度差Δα_w 及び変速機入力軸回転速度差（−Δω_p ）に応じた変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fb及び変速比指令値のフィードバック制御分Ｉ^* _p-fbが算出設定され、両者の加算値から変速機入力トルク指令値Ｔ^* _p 及び変速比指令値Ｉ^* _p が算出設定される。しかしながら、操作量である変速機入力トルク指令値Ｔ^* _p や変速比指令値Ｉ^* _p が飽和しているときには、変速機入力トルク指令値のフィードバック制御分Ｔ^* _p-fb及び変速比指令値のフィードバック制御分Ｉ^* _p-fbを算出設定するときの積分的特性を有するフィードバック補償器だけが停止される。前述のように積分的特性を有するフィードバック補償器は、目標値と実際値との差を蓄積する特性があるので、操作量が飽和しているときに、この積分的特性を有するフィードバック補償器を停止すれば、目標値と実際値との差は蓄積されず、操作量が飽和しなくなったときに制御量のオーバシュートが抑制防止される。
【００８４】
また、前記積分的特性を有するフィードバック補償器の停止条件については、変速機入力トルク指令値Ｔ^* _p が上限値であるのに、加減速度差Δα_w が正値である、即ち更なる加速が要求されるときや、或いは変速機入力トルク指令値Ｔ^* _p が下限値であるのに、加減速度差Δα_w が負値である、即ち更なる減速が要求されるときには、当該加減速度差Δα_w から変速機入力トルク指令値Ｔ^* _p を算出する要素ｙ₁₁のみの演算が停止される。また、変速機入力トルク指令値Ｔ^* _p が上限値であるのに、変速機入力軸回転速度差Δω_p が正値である、即ち更なる増速が要求されるときや、或いは変速機入力トルク指令値Ｔ^* _p が下限値であるのに、変速機入力軸回転速度差Δω_p が負値である、即ち更なる減速が要求されるときには、当該変速機入力軸回転速度差Δω_p から変速機入力トルク指令値Ｔ^* _p を算出する要素ｙ₁₂のみの演算が停止される。また、変速比指令値Ｉ^* _p が上限値であるのに、加減速度差Δα_w が所定値Δα_w1以上である、即ち更なる加速が要求されているときや、或いは変速比指令値Ｉ^* _p が下限値であるのに、加減速度差Δα_w が所定値Δα_w2以下である、即ち更なる減速が要求されているときには、当該加減速度差Δα_w から変速比指令値Ｉ^* _p を算出する要素ｙ₂₁のみの演算が停止される。また、変速比指令値Ｉ^* _p が上限値であるのに、変速機入力軸回転速度差Δω_p が正値である、即ち更なる増速が要求されるときや、或いは変速比指令値Ｉ^* _p が下限値であるのに、変速機入力軸回転速度差Δω_p が負値である、即ち更なる減速が要求されるときには、当該変速機入力軸回転速度差Δω_p から変速比指令値Ｉ^* _p を算出する要素ｙ₂₂のみの演算が停止される。従って、操作量が飽和していない要素の演算は継続されることとなり、その分だけ、より一層頑健なフィードバック制御が可能となる。
【００８５】
また、前述したように操作量が飽和したとき、積分的特性を有するフィードバック補償器のみの演算が停止され、それ以外のフィードバック補償器の演算は継続される。前述のように、現代制御理論に基づくフィードバック補償器を分離して、積分的特性を除いたフィードバック補償器は、逆に言えば、微分的特性、即ち位相進みの特性を有してもおり、従ってフィードバック補償器全体を停止することは、フィードバック補償器演算再開時に、この微分的特性を有するフィードバック補償器が過敏に反応してしまう可能性がある。本実施形態では、このように積分的特性を除いたフィードバック補償器の演算を継続することにより、制御量の過敏な応答を抑制防止することができるのである。
【００８６】
図１１は、本実施形態の加減速度フィードバック制御装置において、発進加速後、所定の間隔でアクセルペダルをオンオフしたときの加減速度、変速機入力軸回転速度、変速機入力トルク、変速比、走行速度、スロットル開度の経時変化を示したものである。このシミュレーションは加減速度の目標値追従を優先している。このシミュレーションでは、時刻ｔ₀₁以後と、時刻ｔ₀₂以後の夫々で、変速機入力トルクが上限値に飽和しており、本実施形態では、それ以後、前記積分的特性を有するフィードバック補償器の演算を停止している。そのため、その後の加減速度のオーバシュートが抑制されている。これに対し、図１２は、フィードバック補償器の演算を一切停止していない。そのため、時刻ｔ₀₁の後、及び時刻ｔ₀₂の後、夫々、加減速度がオーバシュートしている。
【００８７】
また、本実施形態では、道路勾配抵抗力Ｆ_gra を検出して目標加減速度α^* _w を補正する。このとき、道路勾配抵抗力の絶対値｜Ｆ_gra ｜が前記第１所定値Ｆ_a1以下の領域、即ち平坦路又は勾配の小さいときには目標加減速度補正値α_gra を“０”、つまり目標加減速度α^* _w を実質的に補正しないので、例えば道路勾配抵抗力、つまり道路勾配を誤検出したときに運転者に違和感を与えることがない。また、道路勾配抵抗力の絶対値｜Ｆ_gra ｜が前記第２所定値Ｆ_a1以上の領域、即ち勾配の大きいときには目標加減速度補正値α_gra を重力加速度の道路勾配方向成分（ｇ・sin θ）と同等の傾きとし、且つそれより前記所定値α_Vmax分だけ絶対値で小さくなるように設定し、且つ当該所定値α_Vmaxを、最大駆動力を車両質量で除した最大加速度に設定するため、車両の最大駆動力を越える駆動力を加減速度フィードバック制御装置が要求することがなく、駆動力の飽和を防止することができる。更に、道路勾配抵抗力の絶対値｜Ｆ_gra ｜が大きいときには、小さいときに比較して、目標加減速度補正値α_gra の変化率の絶対値を大きくしているため、早期の駆動力飽和を防止でき、運転者はアクセルペダル操作量に対する加減速度のリニアリティを広い領域で保つことができる。
【００８８】
なお、前記実施形態では各コントローラとしてマイクロコンピュータを適用した場合について説明したが、これに代えてカウンタ、比較器等の電子回路を組み合わせて構成することもできる。
【図面の簡単な説明】
【図１】本発明の駆動力制御装置の一実施形態を示す加減速度フィードバック制御装置の概略構成図である。
【図２】図１の加減速度フィードバック制御装置のシステム構成図である。
【図３】図２のプラントモデルを示す構成図である。
【図４】図２のフィードフォワード補償器の構成図である。
【図５】エンジン運転拘束マップである。
【図６】図２のフィードバック補償器及びフィードフォワード補償器で行われる演算処理のフローチャートである。
【図７】図６の演算処理に用いられる制御マップである。
【図８】図６の演算処理に用いられる制御マップである。
【図９】図６の演算処理に用いられる制御マップである。
【図１０】図６の演算処理に用いられる制御マップである。
【図１１】図６の演算処理で行われるマイナプログラムのフローチャートである。
【図１２】本実施形態の作用の説明図である。
【図１３】従来の駆動力制御装置の作用の説明図である。
【符号の説明】
１はエンジン
２は無段変速機
３はトルクコンバータ
４は車輪
５は変速機コントローラ
６は加減速度コントローラ
７はエンジントルクコントローラ
１１はスロットルアクチュエータ
１２はスロットルバルブ
１３はプライマリプーリ
１４はセカンダリプーリ
１６はロックアップクラッチ
３１はフィードフォワード補償器
３３はフィードバック補償器[0001]
[Industrial application fields]
The present invention relates to a driving force control device that controls a driving force so that an acceleration / deceleration of a vehicle matches a target acceleration / deceleration.
[0002]
[Prior art]
An example of such a driving force control device is described in Japanese Patent Application Laid-Open No. 2001-173474. In this driving force control device, a road gradient resistance force proportional to the road gradient is detected, and the target acceleration / deceleration is proportionally corrected according to the road gradient resistance force. In other words, the target acceleration / deceleration is proportionally decreased on the uphill where the road gradient resistance increases, and the target acceleration / deceleration is proportionally increased on the downhill where the road gradient resistance decreases, and the vehicle is adjusted to the target acceleration / deceleration. The engine torque and the gear ratio are feedback controlled so that the speeds coincide. According to this driving force control device, the driver feels uncomfortable by matching the driver's acceleration / deceleration feeling from the road gradient with the actual road gradient.
[0003]
[Problems to be solved by the invention]
However, according to the conventional driving force control device, the target acceleration / deceleration is simply corrected proportionally according to the road gradient. For example, when the flat road or the road gradient is small, compared to when the road gradient is large. The driver's sensitivity to acceleration / deceleration is high. For example, if there is a detection error in the road gradient, there may be a sense of incongruity by correcting the target acceleration / deceleration. In order to avoid this problem, if the correction value of the target acceleration / deceleration is set small, that is, the correction gain is small, the driving force is likely to be saturated when the road gradient is large.
[0004]
The present invention has been developed in view of these various problems, and an object of the present invention is to provide a driving force control device that suppresses and prevents the driver from feeling uncomfortable and that is difficult to saturate the driving force.
[0005]
[Means for Solving the Problems]
  In order to solve the above problem, a driving force control apparatus according to claim 1 of the present invention detects a road gradient in a driving force control apparatus that performs feedback control so that the acceleration / deceleration of the vehicle matches the target acceleration / deceleration. A target acceleration / deceleration correction value is set based on the road gradient detection means and the road gradient detected by the road gradient detection means, and the target acceleration / deceleration correction value isModerateTarget acceleration / deceleration correction means for correcting the target acceleration / deceleration, wherein the target acceleration / deceleration correction means is configured such that when the road gradient detected by the road gradient detection means is large, the target acceleration / deceleration is compared with when the road gradient is small. Change rate of correction valueStep by stepIt is characterized by being enlarged.
[0006]
According to a second aspect of the present invention, there is provided the driving force control apparatus according to the first aspect, wherein the target acceleration / deceleration correction unit is configured to start from a state in which the road gradient detected by the road gradient detection unit is flat. The target acceleration / deceleration correction value is set to zero until a predetermined value is reached.
According to a third aspect of the present invention, in the driving force control device according to the first or second aspect, the target acceleration / deceleration correcting means is configured such that the road gradient detected by the road gradient detecting means is a second road gradient. When the value is equal to or greater than a predetermined value, the target acceleration / deceleration correction value is equal to or close to the road gradient direction component of gravity acceleration.
[0007]
【The invention's effect】
  Thus, according to the driving force control device of the present invention, the road gradient is large.When compared to when smallSince the rate of change of the target acceleration / deceleration correction value is increased, the uncomfortable feeling of the target acceleration / deceleration correction when the road or the road gradient is small is suppressed and the driving force is saturated when the road gradient is large. Suppression can be prevented.
[0008]
Further, according to the driving force control apparatus according to claim 2 of the present invention, the target acceleration / deceleration correction value is set to zero between the state where the road gradient is flat and the first predetermined value. The uncomfortable feeling of the target acceleration / deceleration correction when the road or road gradient is small can be prevented.
In the driving force control apparatus according to claim 3 of the present invention, when the road gradient is equal to or greater than the second predetermined value, the target acceleration / deceleration correction value is equal to or near the road gradient direction component of gravity acceleration. Since the configuration is a value, the driving force is further less likely to be saturated.
[0009]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, an embodiment in which a driving force control device of the present invention is applied to a vehicle acceleration / deceleration control device will be described with reference to the accompanying drawings.
FIG. 1 is a schematic configuration diagram showing an embodiment of an acceleration / deceleration control apparatus according to this embodiment. In the figure, reference numeral 1 denotes an engine, reference numeral 2 denotes a continuously variable transmission, reference numeral 3 denotes a torque converter with a lockup mechanism interposed between the engine 1 and the continuously variable transmission 2, and reference numeral 4 denotes a drive wheel. The engine 1 is configured such that the engine torque can be controlled by adjusting the opening of the throttle valve 12 by a throttle actuator 11 and controlling the intake air amount. The continuously variable transmission 2 is a so-called belt-type continuously variable transmission, and the gear ratio is controlled by controlling the belt contact radii of the primary pulley (input pulley) 13 and the secondary pulley (output pulley) 14. It is comprised so that it can control. The secondary pulley 14 of the continuously variable transmission 2 is connected to the drive wheels 4 via the final reduction gear 15. The torque converter 3 includes a lockup clutch 16.
[0010]
The engine 1 is controlled by an engine controller 7. Therefore, a crank angle sensor 21 for detecting the rotational speed of the engine 1 is provided, and the operating state of the engine 1 is controlled based on the detected value. The continuously variable transmission 2 and the lockup clutch 16 of the torque converter 3 are controlled by the transmission controller 5. Therefore, the primary speed sensor 22 for detecting the rotational speed of the primary pulley 13, that is, the transmission input shaft rotational speed, and the secondary output for detecting the rotational speed of the secondary pulley 14, which is the rotational speed of the transmission output shaft and also the traveling speed of the vehicle. A speed sensor 23 is provided, and the gear ratio of the continuously variable transmission 2 and the engagement state of the lockup clutch 16 are controlled based on the detected value. Incidentally, in this embodiment, the lock-up clutch 16 is released only in the extremely low speed range and is engaged in almost all speed ranges except that it can be stopped and started.
[0011]
The vehicle further includes an acceleration / deceleration controller 6 for controlling the acceleration / deceleration of the host vehicle. The acceleration / deceleration controller 6 is connected to the engine torque controller 7 and the transmission controller 5 through a high-speed communication line, and the information, the accelerator opening detected by the accelerator sensor 24 and the wheel speed detected by the wheel speed sensor 23. The acceleration / deceleration of the host vehicle is controlled based on the above. Specifically, based on the difference between the target acceleration / deceleration and gear ratio and the actual acceleration / deceleration and gear ratio, target values of the transmission input torque and gear ratio are set, and these are set as the engine controller 4 and the gear shift, respectively. It outputs toward the machine controller 5 to control the acceleration / deceleration of the host vehicle. The conventional engine torque controller controls the engine torque according to the accelerator opening and the engine speed, and the conventional transmission controller controls the gear ratio based on the accelerator opening, the engine speed and the traveling speed. The acceleration / deceleration required by the driver and the fuel consumption were compatible to some extent, but in order to further improve the acceleration / deceleration and the fuel consumption, an acceleration / deceleration controller that takes the entire vehicle into account is provided. They are controlled in accordance with the transmission input torque calculated, that is, the engine torque and the gear ratio. Each controller includes an arithmetic processing unit such as a microcomputer.
[0012]
In this vehicle, the acceleration / deceleration control system is configured as shown in FIG. The plant model 34 in the figure is the host vehicle. The output of the host vehicle is acceleration / deceleration α_wAnd transmission input shaft rotation speed ω_pIt is. For example, accelerator opening A_p0And traveling speed, that is, wheel speed V_wAnd target acceleration / deceleration α^* _wAnd the target transmission input shaft rotational speed ω from the engine rotational speed, that is, the transmission input shaft rotational speed and the engine torque, that is, the transmission input torque.^* _pIs determined, the feedforward compensator 31 uses the transfer function G_FFIn accordance with (s), the target acceleration / deceleration α^* _pTo the target transmission input torque command value for feedforward control T^* _p-ffAnd target forward gear ratio command value feedforward control I^* _p-ffIs calculated and set. On the other hand, in the normative model unit 32, a predetermined normative model G_MAccording to (s), standard acceleration / deceleration α_w-refAnd reference transmission input shaft rotational speed ω_p-refIs calculated and set by the adder / subtractor 35, 36 from the acceleration / deceleration α._wAnd transmission input shaft rotation speed ω_pAcceleration / deceleration difference Δα_wAnd transmission input shaft rotation speed difference (−Δω_p) Is calculated. In the feedback compensator 33, this acceleration / deceleration difference Δα_wAnd transmission input shaft rotation speed difference (−Δω_p) For a given transfer function G_FBAccording to (s), the feedback control amount T of the target transmission input torque command value^* _p-fbAnd target feedback ratio command value feedback control I^* _p-fbIs calculated and set. Then, the feedforward control amount T of the target transmission input torque command value^* _p-ffAnd target transmission input torque command value feedback control T^* _p-fbIs added by the adder 37 and the target transmission input torque command value T^* _pIs calculated, and the target speed ratio command value of the feedforward control I^* _p-ffAnd target speed ratio command value feedback control I^* _p-fbIs added by an adder 38 to obtain a target gear ratio command value I^* _pIs calculated.
[0013]
FIG. 3 shows the target transmission input torque command value T of the vehicle model that is the plant model 34 and the reference model unit 32.^* _pFrom reference transmission input shaft rotation speed ω_p-refThe reference transmission input shaft rotation speed calculation unit 32a and the adder / subtractor 36 are shown. First, in the vehicle that is the plant model 34, the transmission input shaft rotational speed ω is set by the upper / lower limiter 301._pIn accordance with the target transmission input torque command value T^* _p(Substantially performed in the feedback compensator 33), and the value is input to the transmission input torque T via the engine torque control system 302 of the first-order lag system._pIt becomes. On the other hand, with another upper / lower limiter 303, the wheel speed V_wDepending on the target gear ratio command value I^* _p(Substantially performed in the feedback compensator 33), the value of which is changed through the first-order delay system gear ratio control system 304._pAnd gear ratio change rate I ′_pIt becomes. Wheel speed V_wIs divided by the effective tire radius by the divider 305, and the wheel angular velocity ω_wWheel angular velocity ω_wAnd the gear ratio change rate I ′_pAre multiplied by a multiplier 306, and further a drive system inertia J is multiplied by a multiplier 307.₁And final reduction ratio I_fInner shuttle T_ineIt becomes.
[0014]
Therefore, the transmission input torque T_pTo Inner Shuttle T_ineThe value obtained by subtracting the value by the adder / subtracter 308 is the drive torque T_wIt becomes. This driving torque T_wIn the multiplier 309, the gear ratio I_pIs multiplied by a multiplier 310 and a final reduction ratio I_fAnd the driving force F by dividing by the effective tire radius R_wIt becomes. In the running resistance system 311, the wheel speed V_wDriving resistance F according to_rThe driving force F_WFrom the driving resistance F by the adder / subtractor 312_rIs the wheel drive force F_dThe wheel acceleration α is obtained by dividing this by the vehicle mass M by the divider 313._wFurther, the wheel speed V is integrated by the integrator 314._wIt becomes. Further, the gear ratio I_pIn the multiplier 316, the wheel angular velocity ω_wIs multiplied by a multiplier 317 and the final reduction ratio I_fMultiplied by the transmission input shaft rotation speed ω_pIt becomes. In the present embodiment, the wheel speed V_wThrough the bandpass filter 315 to accelerate the wheel acceleration / deceleration α_wfIs calculated.
[0015]
On the other hand, in the reference transmission input shaft rotational speed calculation unit 32a, the target transmission input torque command value T is set via the engine torque control system 302 as described above.^* _pIs the transmission input torque T_pTherefore, this transmission input torque T_pFrom the target transmission input shaft rotational speed setting unit 318 according to the engine operation restriction map.^* _pCalculate and set the target transmission input shaft rotational speed ω^* _pIs normalized by a reference model unit 319 of a first-order lag system, and the reference transmission input shaft rotational speed ω_p-refIs obtained. The engine operation restriction map will be described in detail later.
[0016]
FIG. 4 shows the feedforward compensator 31. In the feedforward compensator 31, first, the target acceleration / deceleration α^* _wIs multiplied by the vehicle mass M by the multiplier 401 to obtain the target wheel driving force F.^* _dIs obtained. On the other hand, according to the flat road running resistance map in the flat road running resistance calculating unit 402, the wheel speed V_wFlat road running resistance F according to_rTo calculate the flat road running resistance F_rAnd the target wheel driving force F^* _dIs added by the adder 403 and the target driving force F is added.^* _wIs obtained. This target driving force F^* _wOn the other hand, the multiplier 404 multiplies the effective tire radius R and the final reduction ratio I._fAnd subtractor 405 to change gear ratio I._pBy dividing by the target transmission input torque command value T for feedforward control^* _p0Is obtained. Target transmission input torque command value T for feedforward control^* _p0The calculation formula is shown in the following one formula.
[0017]
[Expression 1]

[0018]
Then, by the acceleration / deceleration model matching compensator 406, the feedforward control target transmission input torque command value T^* _p0From feedforward control for transmission input torque command value T^* _p-ffIs obtained. Transfer function G of acceleration / deceleration model matching compensator 406_FF-1(s) is shown in the following two equations. This acceleration / deceleration model matching compensator 406 standardizes using a normative model relating to acceleration / deceleration (denominator in the equation), and at the same time, a first-order delay transmission input torque control system, that is, an engine torque control system Phase matching is performed by multiplying the reciprocal of the response delay (numerator in the equation). In the formula, s is a Laplace operator, τ_engIs the response delay time constant of the engine torque control system, ω_r, Ζ_rIs the target acceleration / deceleration α^* _pAcceleration / deceleration with respect to_pIs the cutoff frequency and damping constant of the reference model response (second-order lag model).
[0019]
[Expression 2]

[0020]
However, in order to perform arithmetic processing by the microcomputer described above, for example, it is discretized by Tustin approximation or the like, and a differential equation that can be executed by software is obtained as shown by the following equation 2a to obtain the transmission input torque command value. Feed forward control T^* _p-ffIs calculated. In the formula, MTN0, MTN1, MTN2, MTD1, and MTD2 are the time constant τ._eng, Cutoff frequency ω_r, Damping constant ζ_r, A constant determined from the sampling period ΔT of the arithmetic processing. (K) is the current value, (k-1) is the previous value, and (k-2) is the previous value.
[0021]
[Equation 3]

[0022]
On the other hand, the target driving force F^* _wAnd wheel speed V_wIs multiplied by the multiplier 407 and the target engine power (output) P^*Therefore, the target transmission input rotational speed setting unit 408 uses the engine operation restriction map described above to generate the target engine power P^*Target transmission input rotational speed ω that achieves optimal fuel efficiency^* _pIs calculated and set. Target engine power P^*The calculation formula is shown in the following three formulas.
[0023]
[Expression 4]

[0024]
Therefore, in the divider 409, this target transmission input rotational speed ω^* _pThe wheel speed V_wIn addition, the multiplier 410 multiplies the effective tire radius R and the final reduction ratio I._fThe target gear ratio I for feedforward control is divided by^* _p0Is obtained. Target gear ratio I for feedforward control^* _p0The calculation formula is shown in the following four formulas.
[0025]
[Equation 5]

[0026]
Then, the gear ratio model matching compensator 411 provides the target gear ratio I for feedforward control.^* _p0To the feed-forward control for the gear ratio command value I^* _p-ffIs obtained. Transfer function G of gear ratio model matching compensator 411_FF-2(s) is shown in the following five equations. The transmission ratio model matching compensator 411 normalizes using a reference model relating to the transmission ratio (denominator in the equation), and at the same time, the reciprocal of the response delay of the transmission ratio control system of the first-order lag system ahead of the output (in the equation) Phase alignment is performed by multiplying by (numerator). In the formula, τ_cvtIs the response delay time constant of the gear ratio control system, τ_ref-wpIs the target transmission input shaft rotational speed ω^* _pTransmission input shaft rotational speed ω_pIs the time constant of the reference model response (first-order lag model).
[0027]
[Formula 6]

[0028]
However, in order to perform arithmetic processing by the above-described microcomputer, for example, it is discretized by Tustin approximation or the like, and a difference equation that can be executed by software is obtained as shown by the following equation 5a to feed forward the transmission ratio command value. Control minute I^* _p-ffIs calculated. In the formula, MIN0, MIN1, and MID1 are the time constant τ_cvt, Τ_ref-wp, A constant determined from the sampling period ΔT of the arithmetic processing.
[0029]
[Expression 7]

[0030]
Next, the engine operation restriction map will be described with reference to FIG. For example, the engine speed ω on the horizontal axis as shown in the figure_e(= Transmission input rotational speed ω_p) And the vertical axis represents engine torque T_e(= Transmission input torque T_p), An equal output line (broken line in the figure) connecting the same engine power (output) and an equal fuel consumption line (dashed line in the figure) centering on the optimum fuel efficiency point can be drawn. A curve in which the optimum fuel consumption points on the iso-output line are continuous becomes the optimum fuel consumption driving line. Generally, in modern engines, fuel is not injected when the accelerator is off, so the optimum fuel consumption point and isofuel consumption line are the engine torque T_eExists only in the positive region. Therefore, the optimum fuel consumption driving line is also the engine torque T_eExists only in the positive region. Conversely, engine torque T_eIn the negative region, an engine brake characteristic line indicating the relationship between the engine brake torque and the engine speed appears. As mentioned above, the engine torque T_eHowever, since fuel is not injected in the negative region, it is necessary to control the engine speed in order to control the engine brake torque. In this embodiment, since a continuously variable transmission is used as the transmission, in order to obtain a desired engine brake torque at an arbitrary traveling speed, the speed ratio of the continuously variable transmission may be controlled. An engine operation restriction map is obtained by mapping the relationship between these curves in consideration of fuel consumption.
[0031]
Next, a design method of the feedback compensator 33 in this embodiment will be briefly described. The above-described nonlinear control target model of FIG.^* _p, Gear ratio command value I^* _p2 inputs, acceleration / deceleration α_w, Transmission input shaft rotational speed difference Δω_pIt is assumed that the model is a two-output nonlinear controlled object model. In the present embodiment, for the purpose of ensuring the stability of the feedback control system, the detection unit and the partial control unit are combined with a vehicle model as a controlled object model. This nonlinear control target model is linearly approximated at a specific operating point to derive a linear approximate control target model for control system design. In order to design a feedback compensator using “μ synthesis”, which is one of “robust control theory”, it is necessary to further model variable elements and extend them to a generalized plant model. Will be omitted.
[0032]
The input to the feedback compensator 33 is the acceleration / deceleration difference Δα as described above._w, Transmission input shaft rotational speed difference (−Δω_pTherefore, the output transmission input torque command value T^* _p, Gear ratio command value I^* _pThe transfer function G of the feedback compensator 33_FBWhen represented by (s), the following six equations are obtained, and the transfer function G_FBEach element of (s) is expressed by Equation 7.
[0033]
[Equation 8]

[0034]
Substituting actual vehicle specifications or required response characteristics, and each element G by the “μ synthesis”.₁₁(s) ~ G_{twenty two}When (s) is obtained, it is expressed as the following formulas 8 to 11.
[0035]
[Equation 9]

[0036]
Each of these elements G₁₁(s) ~ G_{twenty two}Considering (s) in detail, there is a part where the pole is on or near the imaginary axis. The pole on or in the vicinity of the imaginary axis is a pole with a slow response and accumulates the difference between the target value and the actual value, in other words, it is a part having an integral characteristic. Therefore, the transfer function of the feedback compensator of equation (7) is expressed as a portion G having an integral characteristic._A(s) and other parts G_B(s) and is expressed as shown in the following equation (12).
[0037]
[Expression 10]

[0038]
Specific element G_11-A(s) ~ G_22-A(s), G_11-B(s) ~ G_22-B(s) is expressed by the following equations 13 to 20.
[0039]
## EQU11 ##

[0040]
And the acceleration / deceleration α_wAnd transmission input shaft rotation speed difference (−Δω_p) Element G having no integral characteristic_11-B(s) ~ G_22-Bx with element (s)₁₁~ X_{twenty two}And an element G having these integral characteristics in these elements_11-A(s) ~ G_22-A(s) to y₁₁~ Y_{twenty two}Then, the feedback control amount T of the transmission input torque command value^* _p-fb, Feedback ratio of gear ratio command value I^* _p-fbAre expressed by the following formulas 21 and 22, respectively.
[0041]
[Expression 12]

[0042]
Next, calculation processing performed by the feedforward compensator 31 and the feedback compensator 33 will be described with reference to the flowchart of FIG. This calculation process is performed at a predetermined sampling period ΔT of about 10 msec., For example. In this calculation process, all the steps for communication are not described, but necessary information is exchanged with other controllers or storage devices as needed, and information obtained in the calculation process is always changed with other controllers. Alternatively, it is exchanged with a storage device.
[0043]
In this calculation process, first, the accelerator opening A detected by the accelerator sensor 24 in step S1._p0Is read.
Next, the process proceeds to step S2 where the wheel speed V detected by the wheel speed sensor 25 is detected._wIs read.
Next, the process proceeds to step S3, where the primary rotational speed ω is transmitted from the transmission controller 5._pSecondary rotational speed ω_s, Gear ratio I which is the ratio of both_pAnd the engine speed ω from the engine torque controller 7_eIs read.
[0044]
Next, the process proceeds to step S4, and the accelerator opening A read in step S1 is read according to the control map shown in FIG._p0, Wheel speed V read in step S2_wBased on target acceleration / deceleration α^* _wIs calculated and set.
Next, the process proceeds to step S5, and the transfer function G shown by the following equation 23 is obtained._bpUsing the bandpass filter (s), the wheel speed V_wRemoves noise from the wheel acceleration / deceleration α only in the specified frequency range_wIs calculated. In the formula, ω_nIs the natural angular frequency, ζ_nIs the attenuation factor, ω_n, Ζ_nIs determined by the noise level of the detected wheel speed.
[0045]
[Formula 13]

[0046]
However, in order to perform arithmetic processing by the microcomputer described above, the wheel acceleration / deceleration α is obtained by discretizing by, for example, Tustin approximation and obtaining a differential equation that can be executed by software as shown by the following equation 23a._wIs calculated. Note that BPN0, BPN1, and BPN2 in the formula are the natural angular frequencies ω_n, Damping rate ζ_n, A constant determined by the sampling period ΔT.
[0047]
[Expression 14]

[0048]
Next, the process proceeds to step S6, for example, the transmission input torque command value T at the time of the previous calculation.^* _pTransmission torque T from engine torque response delay model_pIs calculated. This engine torque response delay model is a first-order delay system represented by the following 25 transfer functions.
[0049]
[Expression 15]

[0050]
However, in order to perform the arithmetic processing by the microcomputer described above, the transmission input torque T is obtained by discretizing by, for example, Tustin approximation or the like and obtaining a differential equation that can be executed by software as shown by the following equation 25a._pIs calculated. In the formula, TEN0, TEN1, and TEN2 are the time constant τ_eng, A constant determined by the sampling period ΔT.
[0051]
[Expression 16]

[0052]
Next, the process proceeds to step S7, where road gradient resistance force F_graIs calculated. Specifically, as described above, the transmission input torque T calculated in the step S6._p, Gear ratio I_p, Final reduction ratio I_fUsing the tire effective radius R, the driving force F according to the following equation 26_dIs calculated. Next, the acceleration / deceleration α calculated in step S5 according to the following equation (27)_wActual driving force F from the product of vehicle mass M_realIs calculated. Then, according to the following equation 28, the driving force F_dTo actual driving force F_realTo reduce the total driving resistance F_allIs calculated. Next, according to the following equation 29, the total running resistance F_allFrom the wheel speed V as described above._wFlat road running resistance F obtained according to_rReduce road slope resistance F before noise removal_gralIs calculated.
[0053]
[Expression 17]

[0054]
Next, the following equation 30 transfer function F₂Road slope resistance force F by removing unnecessary high-frequency noise components with the low-pass filter shown in (s)_graIs calculated. In the formula, τ_graIs the cut-off frequency of the low-pass filter.
[0055]
[Expression 18]

[0056]
However, in order to perform the arithmetic processing by the microcomputer described above, the road gradient resistance force F is obtained by discretizing by, for example, Tustin approximation or the like and obtaining a differential equation that can be executed by software as shown in the following equation 31._graIs calculated. Note that SLPNO, SLPN1 and SLPD1 in the equation are the cut-off frequency τ._gra, A constant determined by the sampling period ΔT.
[0057]
[Equation 19]

[0058]
Next, the process proceeds to step S8, and the road gradient resistance force F calculated in step S7 is calculated according to the control map shown in FIG._graTo target acceleration / deceleration correction value α_graIs calculated. This control map shows the road gradient resistance force F acting on the uphill._graIs a positive value, road slope resistance F acting on the downhill_graThe absolute value of road gradient resistance force | F_graFirst predetermined value F that is relatively small_a1In the following areas, the target acceleration / deceleration correction value α_graIs "0" and road gradient resistance force F_graIs a relatively large second predetermined value F_a2In the above region, the target acceleration / deceleration correction value α_graIs equal to the acceleration / deceleration (g · sin θ) due to the road gradient, and a predetermined value α_VmaxIt is set so that it becomes small only, and road gradient resistance force F_graIs a second predetermined value (−F_a2) Target acceleration / deceleration correction value α_graIs equal to the acceleration / deceleration (g · sin θ) due to the road gradient, and a predetermined value α_VmaxIs set to be larger by a predetermined value (−F_a2) To a predetermined value (-F_a1) Or a predetermined value F_a1To predetermined value F_a2Between the target acceleration / deceleration correction value α_graIs a relatively small slope and the road gradient resistance F_graIt is set to increase or decrease linearly with the increase or decrease of. The predetermined value α_VmaxIs a value obtained by dividing the maximum driving force of the vehicle by the vehicle mass.
[0059]
Next, the process proceeds to step S9, where the target acceleration / deceleration α^* _wTo the target acceleration / deceleration correction value α calculated in step S8._graIs added to the new target acceleration / deceleration α^* _wCorrect as
Next, the process proceeds to step S10, where the reference transmission input shaft rotational speed ω according to the engine operating constraint condition described above is obtained._p-refAnd transmission input shaft rotation speed ω_pDifference of input shaft speed Δω_pIs calculated. Specifically, first, the transmission input torque T calculated in step S6._pTo the transmission input shaft rotation speed on the optimum fuel consumption driving line or engine brake characteristic line from the target transmission input shaft rotation speed ω^* _pThis is the transfer function G of the following 32 equations_ref-wpNormalized using the reference model response characteristic shown in (s), the reference transmission input rotational speed ω_p-refIs calculated.
[0060]
[Expression 20]

[0061]
However, in order to perform arithmetic processing by the microcomputer described above, for example, it is discretized by Tustin approximation or the like, and a differential equation that can be executed by software is obtained as shown by the following equation 32a to obtain a reference transmission input rotational speed ω._p-refIs calculated. Note that PRN0, PRN1, and PRD2 in the formula are the time constants τ_ref-wp, A constant determined by the sampling period ΔT.
[0062]
[Expression 21]

[0063]
Then, as shown by the following equation 33, the obtained standard transmission input rotational speed ω_p-refTo the transmission input shaft rotational speed ω_pReduce the transmission input shaft rotation speed difference Δω_pIs calculated.
[0064]
[Expression 22]

[0065]
Next, the process proceeds to step S11, where the feedforward compensator 31 feeds the target transmission input torque command value to the feedforward control amount T.^* _p-ffAnd target forward gear ratio command value feedforward control I^* _p-ffIs calculated.
Next, the process proceeds to step S12, and the feedback control standard acceleration / deceleration α_w-refAnd acceleration / deceleration α_wAcceleration / deceleration difference Δα_wIs calculated. Specifically, the transfer function G shown by the following equation 34_ref-aA delay compensation (second-order lag model) corresponding to the reference model response of acceleration / deceleration consisting of_bpa delay compensation (second-order lag model) corresponding to the acceleration / deceleration bandpass filter comprising (s) and a target acceleration / deceleration α^* _wApplied to the norm acceleration / deceleration α_w-refIs calculated.
[0066]
[Expression 23]

[0067]
However, in order to perform the arithmetic processing by the microcomputer described above, it is discretized by, for example, Tustin approximation or the like, and a differential equation that can be executed by software is obtained as shown by the following equation 34a to obtain the reference acceleration / deceleration α_w-refIs calculated. In the formula, REN0, REN1, REN2, REN3, REN4, RED1, RED2, RED3, RED4 are the cut-off frequency ω_r, Damping constant ζ_r, Natural angular frequency ω_n, Damping rate ζ_n, A constant determined from the sampling period ΔT. In addition, (k-4) indicates the value before the last time.
[0068]
[Expression 24]

[0069]
Next, the process proceeds to step S13, and the acceleration / deceleration speed difference Δα as described above._w, Transmission input shaft rotational speed difference (−Δω_p) For the feedback compensator G excluding the integral characteristic_11-B(s) ~ G_22-BFeedback control amount T of transmission input torque command value by (s)^* _p-fbElement x₁₁, X₁₂, And feedback control amount I of the gear ratio command value^* _p-fbElement x_{twenty one}, X_{twenty two}Is calculated.
[0070]
Next, the process proceeds to step S14, and a feedback control amount T of the transmission input torque command value calculated in step S13 is calculated according to the calculation process of FIG.^* _p-fbElement x₁₁, X₁₂, And feedback control amount I of the gear ratio command value^* _p-fbElement x_{twenty one}, X_{twenty two}In contrast, the feedback compensator G having the integral characteristic_11-A(s) ~ G_22-AFeedback control amount T of transmission input torque command value by (s)^* _p-fbElement y₁₁, Y₁₂, And feedback control amount I of the gear ratio command value^* _p-fbElement y_{twenty one}, Y_{twenty two}Is calculated.
[0071]
Next, the process proceeds to step S15, and the element y calculated in step S14 is calculated according to the above equations 21 and 22.₁₁, Y₁₂The feedback control amount T of the transmission input torque command value from the added value of^* _p-fbAnd the element y_{twenty one}, Y_{twenty two}The feedback control amount I of the gear ratio command value from the added value of I^* _p-fbIs calculated.
Next, the process proceeds to step S16, and according to the following formulas 35 and 36, the feed-forward control amount T of the transmission input torque command value is set.^* _p-ffAnd feedback control amount T^* _p-fbTransmission input torque command value T^* _pAs well as the feedforward control amount I of the gear ratio command value.^* _p-ffAnd feedback control I^* _p-fbTo the gear ratio command value I^* _pIs calculated.
[0072]
[Expression 25]

[0073]
Next, the process proceeds to step S17, where the transmission input torque command value T is determined from the control maps shown in FIGS.^* _p, Gear ratio command value I^* _pA restriction process is applied to. That is, the target value is limited by the upper and lower limit values of the control amount that can actually be generated.
Next, the process proceeds to step S18, where the transmission input torque command value T^* _p, Gear ratio command value I^* _pAre output to the engine torque controller 7 and the transmission controller 5, respectively, and then returned to the main program.
[0074]
Next, the minor program executed in step S14 of the calculation process of FIG. 6 will be described with reference to the flowchart of FIG. In this calculation process, first, in step S21, the previous value T of the transmission input torque command value is set.^* _p(k-1) is an upper limit value and the acceleration / deceleration difference Δα_wIs greater than or equal to “0” or the previous value T of the transmission input torque command value^* _p(k-1) is the lower limit and the acceleration / deceleration difference Δα_wIs determined to be “0” or less. If any of the conditions is satisfied, the process proceeds to step S22. If not, the process proceeds to step S23.
[0075]
In step S23, a feedback control amount T of the transmission input torque command value is obtained.^* _p-fbElement x₁₁In contrast, the feedback compensator G having the integral characteristic_11-AFeedback control amount T of transmission input torque command value by (s)^* _p-fbElement y₁₁After calculating, the process proceeds to step S24. Specifically, the current value y of the element based on the difference equation obtained by discretization as described above.₁₁Update (k).
[0076]
In step S22, a feedback control amount T of the transmission input torque command value is obtained.^* _p-fbLast value y of the element of₁₁(k-1) is the current value y₁₁After outputting as (k), the process proceeds to step S24. In essence, the current value y of the elements of the difference equation described in step S23.₁₁(k) is not updated and the previous value y₁₁Save as (k-1).
In step S24, the previous value T of the transmission input torque command value.^* _p(k-1) is an upper limit value and the transmission input shaft rotational speed difference Δω_pIs greater than or equal to “0” or the previous value T of the transmission input torque command value^* _p(k-1) is a lower limit value and the transmission input shaft rotational speed difference Δω_pIs determined to be less than or equal to “0”. If any of the conditions is satisfied, the process proceeds to step S25, and if not, the process proceeds to step S26.
[0077]
In step S26, a feedback control amount T of the transmission input torque command value is obtained.^* _p-fbElement x₁₂In contrast, the feedback compensator G having the integral characteristic_12-AFeedback control amount T of transmission input torque command value by (s)^* _p-fbElement y₁₂After calculating, the process proceeds to step S27. Specifically, the current value y of the element based on the difference equation obtained by discretization as described above.₁₂Update (k).
[0078]
In step S25, a feedback control amount T of the transmission input torque command value is obtained.^* _p-fbLast value y of the element of₁₂(k-1) is the current value y₁₂After outputting as (k), the process proceeds to step S27. In essence, the current value y of the element of the difference equation described in step S26.₁₂(k) is not updated and the previous value y₁₂Save as (k-1).
In step S27, the previous value I of the gear ratio command value is obtained.^* _p(k-1) is an upper limit value and the acceleration / deceleration difference Δα_wIs the predetermined value Δα_w1Or the previous value I of the gear ratio command value^* _p(k-1) is the lower limit and the acceleration / deceleration difference Δα_wIs the predetermined value Δα_w2It is determined whether or not the following conditions are satisfied. If any of the conditions is satisfied, the process proceeds to step S28, and if not, the process proceeds to step S29. The predetermined value Δα_w1Is a positive value and means an acceleration command, and the predetermined value Δα_w2Is a negative value and means a deceleration command.
[0079]
In the step S29, the feedback ratio I of the gear ratio command value^* _p-fbElement x_{twenty one}In contrast, the feedback compensator G having the integral characteristic_21-AFeedback ratio of gear ratio command value by (s) I^* _p-fbElement y_{twenty one}Is calculated, and then the process proceeds to step S30. Specifically, the current value y of the element based on the difference equation obtained by discretization as described above._{twenty one}Update (k).
[0080]
In step S28, a feedback control amount I of the gear ratio command value is obtained.^* _p-fbLast value y of the element of_{twenty one}(k-1) is the current value y_{twenty one}After outputting as (k), the process proceeds to step S30. In effect, the current value y of the elements of the difference equation described in step S29._{twenty one}(k) is not updated and the previous value y_{twenty one}Save as (k-1).
In step S30, the previous value I of the gear ratio command value^* _p(k-1) is an upper limit value and the transmission input shaft rotational speed difference Δω_pIs greater than or equal to “0” or the previous value I of the gear ratio command value^* _p(k-1) is the lower limit and transmission input shaft rotational speed difference Δω_pIs determined to be less than or equal to “0”, the process proceeds to step S31 if any of the conditions is satisfied, and the process proceeds to step S32 otherwise.
[0081]
In step S32, a feedback control amount I of the gear ratio command value is obtained.^* _p-fbElement x_{twenty two}In contrast, the feedback compensator G having the integral characteristic_22-AFeedback ratio of gear ratio command value by (s) I^* _p-fbElement y_{twenty two}After calculating, the process proceeds to step S15 of the calculation process of FIG. Specifically, the current value y of the element based on the difference equation obtained by discretization as described above._{twenty two}Update (k).
[0082]
In step S31, a feedback control amount I of the gear ratio command value is obtained.^* _p-fbLast value y of the element of_{twenty two}(k-1) is the current value y_{twenty two}After the output as (k), the process proceeds to step S15 of the calculation process of FIG. In essence, the current value y of the elements of the difference equation described in step S32_{twenty two}(k) is not updated and the previous value y_{twenty two}Save as (k-1).
[0083]
According to these calculation processes, the target acceleration / deceleration α^* _wFeedforward control amount T of target transmission input torque command value according to^* _p-ffAnd target forward gear ratio command value feedforward control I^* _p-ffIs calculated and set, and the acceleration / deceleration difference Δα_wAnd transmission input shaft rotation speed difference (−Δω_p) According to the feedback control amount T of the transmission input torque command value according to^* _p-fbAnd the feedback control component I of the gear ratio command value I^* _p-fbIs calculated and set, and the transmission input torque command value T is calculated from the sum of the two.^* _pAnd gear ratio command value I^* _pIs calculated and set. However, the transmission input torque command value T which is the operation amount^* _pAnd gear ratio command value I^* _pIs saturated, the feedback control amount T of the transmission input torque command value^* _p-fbAnd the feedback control component I of the gear ratio command value I^* _p-fbOnly the feedback compensator having the integral characteristic when calculating and setting is stopped. As described above, the feedback compensator having an integral characteristic has a characteristic of accumulating the difference between the target value and the actual value. Therefore, when the manipulated variable is saturated, the feedback compensator having the integral characteristic is If the operation stops, the difference between the target value and the actual value is not accumulated, and the overshoot of the control amount is suppressed and prevented when the manipulated variable is no longer saturated.
[0084]
Further, regarding the stop condition of the feedback compensator having the integral characteristic, the transmission input torque command value T^* _pAcceleration / deceleration difference Δα_wIs a positive value, that is, when further acceleration is required, or transmission input torque command value T^* _pIs the lower limit, but the acceleration / deceleration difference Δα_wIs a negative value, that is, when further deceleration is required, the acceleration / deceleration difference Δα_wTo transmission input torque command value T^* _pElement y to calculate₁₁Only the operation is stopped. Further, the transmission input torque command value T^* _pIs the upper limit, but the transmission input shaft rotational speed difference Δω_pIs a positive value, that is, when further speedup is required, or transmission input torque command value T^* _pIs the lower limit, but the transmission input shaft rotational speed difference Δω_pIs a negative value, that is, when further deceleration is required, the transmission input shaft rotational speed difference Δω_pTo transmission input torque command value T^* _pElement y to calculate₁₂Only the operation is stopped. Further, the gear ratio command value I^* _pAcceleration / deceleration difference Δα_wIs the predetermined value Δα_w1That is, that is, when further acceleration is required, or the gear ratio command value I^* _pIs the lower limit, but the acceleration / deceleration difference Δα_wIs the predetermined value Δα_w2When it is below, that is, when further deceleration is required, the acceleration / deceleration difference Δα_wTo gear ratio command value I^* _pElement y to calculate_{twenty one}Only the operation is stopped. Further, the gear ratio command value I^* _pIs the upper limit, but the transmission input shaft rotational speed difference Δω_pIs a positive value, that is, when further speedup is required, or the gear ratio command value I^* _pIs the lower limit, but the transmission input shaft rotational speed difference Δω_pIs a negative value, that is, when further deceleration is required, the transmission input shaft rotational speed difference Δω_pTo gear ratio command value I^* _pElement y to calculate_{twenty two}Only the operation is stopped. Therefore, the calculation of the element in which the operation amount is not saturated is continued, and more robust feedback control can be performed correspondingly.
[0085]
Further, as described above, when the manipulated variable is saturated, the calculation of only the feedback compensator having the integral characteristic is stopped, and the calculation of the other feedback compensators is continued. As described above, the feedback compensator that separates the feedback compensator based on modern control theory and excludes the integral characteristic, conversely, has a differential characteristic, that is, a phase advance characteristic. Therefore, stopping the entire feedback compensator may cause the feedback compensator having this differential characteristic to react sensitively when restarting the feedback compensator operation. In the present embodiment, by continuing the operation of the feedback compensator excluding the integral characteristic in this way, it is possible to suppress and prevent the sensitive response of the control amount.
[0086]
FIG. 11 shows acceleration / deceleration feedback control apparatus according to the present embodiment. Acceleration / deceleration, transmission input shaft rotation speed, transmission input torque, transmission ratio, and traveling speed when the accelerator pedal is turned on / off at predetermined intervals after starting acceleration. FIG. 4 shows the change over time in the throttle opening. In this simulation, priority is given to following the target value of acceleration / deceleration. In this simulation, time t₀₁After and time t₀₂Thereafter, the transmission input torque is saturated to the upper limit value. In this embodiment, the calculation of the feedback compensator having the integral characteristic is stopped thereafter. Therefore, the subsequent overshoot of acceleration / deceleration is suppressed. On the other hand, FIG. 12 does not stop the operation of the feedback compensator at all. Therefore, time t₀₁And after time t₀₂After that, the acceleration / deceleration overshoots, respectively.
[0087]
  In this embodiment, the road gradient resistance force F_graDetect target acceleration / deceleration α^* _wCorrect. At this time, absolute value of road gradient resistance force | F_graIs the first predetermined value F_a1Target acceleration / deceleration correction value α when the following area, that is, flat road or slope is small_graIs “0”, that is, target acceleration / deceleration α^* _wIs not substantially corrected, for example, when the road gradient resistance, that is, the road gradient is erroneously detected, the driver does not feel uncomfortable. The absolute value of road gradient resistance | F_graIs the second predetermined value F_a1In the above region, that is, when the gradient is large, the target acceleration / deceleration correction value α_graIs equivalent to the road gradient direction component (g · sin θ) of gravitational acceleration, and the predetermined value α_VmaxIs set to be smaller by an absolute value by the minute, and the predetermined value α_VmaxIs set to the maximum acceleration obtained by dividing the maximum driving force by the vehicle mass, so that the acceleration / deceleration feedback control device does not require the driving force exceeding the maximum driving force of the vehicle, and saturation of the driving force can be prevented. . Furthermore, the absolute value of road gradient resistance | F_gra| Is largeWhen compared to when it is smallTarget acceleration / deceleration correction value α_graSince the absolute value of the change rate of the vehicle is increased, early driving force saturation can be prevented, and the driver can maintain the linearity of acceleration / deceleration with respect to the accelerator pedal operation amount in a wide range.
[0088]
In the above-described embodiment, the case where a microcomputer is applied as each controller has been described. However, instead of this, electronic circuits such as a counter and a comparator may be combined.
[Brief description of the drawings]
FIG. 1 is a schematic configuration diagram of an acceleration / deceleration feedback control device showing an embodiment of a driving force control device of the present invention.
2 is a system configuration diagram of the acceleration / deceleration feedback control apparatus of FIG. 1; FIG.
FIG. 3 is a configuration diagram showing the plant model of FIG. 2;
4 is a configuration diagram of the feedforward compensator of FIG. 2;
FIG. 5 is an engine operation restriction map.
6 is a flowchart of arithmetic processing performed in the feedback compensator and the feedforward compensator of FIG. 2;
FIG. 7 is a control map used for the arithmetic processing of FIG.
FIG. 8 is a control map used in the arithmetic processing of FIG.
FIG. 9 is a control map used for the arithmetic processing of FIG. 6;
10 is a control map used for the arithmetic processing in FIG. 6. FIG.
FIG. 11 is a flowchart of a minor program performed in the arithmetic processing of FIG.
FIG. 12 is an explanatory diagram of the operation of the present embodiment.
FIG. 13 is an explanatory diagram of the operation of a conventional driving force control device.
[Explanation of symbols]
1 is the engine
2 is a continuously variable transmission
3 is a torque converter
4 is the wheel
5 is a transmission controller
6 is the acceleration / deceleration controller
7 is the engine torque controller
11 is a throttle actuator
12 is a throttle valve
13 is the primary pulley
14 is a secondary pulley
16 is a lock-up clutch
31 is a feedforward compensator
33 is a feedback compensator

Claims (3)

In a driving force control apparatus that performs feedback control so that the acceleration / deceleration of a vehicle matches a target acceleration / deceleration, a road gradient detection unit that detects a road gradient, and a target acceleration / deceleration based on the road gradient detected by the road gradient detection unit A target acceleration / deceleration correction unit that sets a speed correction value and corrects the target acceleration / deceleration by adjusting the target acceleration / deceleration correction value; and the target acceleration / deceleration correction unit detects the road detected by the road gradient detection unit. A driving force control device characterized by increasing the rate of change of the target acceleration / deceleration correction value stepwise when the gradient is large compared to when the gradient is small.   The target acceleration / deceleration correction unit sets the target acceleration / deceleration correction value to zero during a period from a flat state to a first predetermined value detected by the road gradient detection unit. Item 2. The driving force control device according to Item 1.   When the road gradient detected by the road gradient detection unit is greater than or equal to a second predetermined value, the target acceleration / deceleration correction unit sets the target acceleration / deceleration correction value equal to or near the road gradient direction component of gravity acceleration. The driving force control device according to claim 1, wherein:

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