JP3616734B2 - Sliding mode controller - Google Patents

Description

状態方程式は、
【００５６】
【数２】

とおけるので、微分方程式（３．１）を（３．２）に代入すると、以下のようになる。
【００５７】
【数３】

４．切換関数の設計
スライディングモード制御は、システムの状態により、フィードバックゲインを切り換えるので、この切換関数Ｓを以下のように置く。
【００５８】
【数４】

切換関数のパラメータによりスライディングモードが発生しない場合があるため、切換関数の設計は非常に重要である。設計方法は、主に以下のような方法がある。
【００５９】
▲１▼極配置法を用いた設計法
▲２▼最適切換超平面の設計法
▲３▼システムの零点を用いた設計法
▲４▼周波数整形による超平面の設計法
α_１，α_２を上記の設計法を適用して求め、α_１：α_２＝γ：１が成立するγを求めると、以下のようになる。
【００６０】
【数５】

しかし、上記のように、通常の教科書とおりに設計された切換関数では、制御対象の実際の位置つまりＶＴＣの実角度ｘの関数としているため、バルブタイミング制御装置に対しては、以下のように不適となる。
【００６１】
まず、γｘの項については、ＶＴＣの目標角度が０°以外の場合、常に正の値がついてしまい、目標角度と実角度とに、無関係な値となるため、ＶＴＣが目標角度に収束しない。
【００６２】
また、ｄｘ／ｄｔの項については、電磁切換弁が不感帯にあるときには、ＶＴＣが動作しないため、実速度ｄｘ／ｄｔは変化せず、微小角度だけ動作させたい場合には、応答性が悪い。
【００６３】
なお、教科書とおりの設計では、エラー量の積分項も付加することが推奨されているが、カムシャフトが目標角度にされたときに、該エラー量の積分項が０でない値で残されることとなり、目標角度への収束を妨げるように機能してしまう。
【００６４】
そこで、切換関数Ｓを以下のようなエラー量の関数として設定する。
【００６５】
【数６】

ここで、該切換関数の設計には、▲３▼のシステムの零点を用いた設計法を利用した。該システムの零点は、（Ｓ，Ａ，Ｂ）の零点を複素平面上左半面に設定する手法である（Ｓ：切換関数、Ａ，Ｂ：（３．２）式の定数）。
【００６６】
５．スライディング条件の算出
スライディングが成立する最も単純な条件は、Ｓ・ｄＳ／ｄｔ＜０である。
Ｓが減少していくときのみ、上記条件が成立する。Ｓは、エラーとエラーの微分値を変数としているので、上記条件成立時はエラーが減少し、目標値に収束していくことを意味する。
【００６７】
最初にＳの展開に必要な式を求める。
制御量ｕを以下のようにおく。
【００６８】
【数７】

これを（３・１）式に代入すると、
【００６９】
【数８】

次にハットｕについて展開する。
【００７０】
ハットｕは、スライディングしているときの入力なので、Ｓ＝ｄＳ／ｄｔ＝０である。
【００７１】
【数９】

ｂｕ＝ハットｕとすると、
【００７２】
【数１０】

スライディングする条件Ｓ・ｄＳ／ｄｔ＜０について考える。
【００７３】
【数１１】

（５・２），（５・３）式より、
【００７４】
【数１２】

したがって、ｋを正の値にとれば、スライディングが成立する。
【００７５】
６．制御量演算式の設計
制御量（フィードバック補正量）ｕは、式（５．１），（５．３）より、以下のようになる。
【００７６】
【数１３】

ＶＴＣの伝達関数を簡素化した（２．１）式を用いると、状態方程式は以下のようにおける。
【００７７】
【数１４】

（６．２）の状態方程式を用いると、（６．１）式は、以下のようになる。
【００７８】
【数１５】

ここで、α＝ｂ^−１（ａ−γ）、ｋ’＝ｂ^−１ｋとおくと、
【００７９】
【数１６】

この式は、切換線Ｓ＝０上をスライディングしながら動くことを保証する式である。
【００８０】
しかし、このように（教科書とおりに）設計された制御量の式では、線形項α・ｄｘ／ｄｔについても、不感帯にあるときに油の給排が行なわれないため、動作速度ｄｘ／ｄｔ＝０→線形項＝０となって有効に機能しない。
【００８１】
そこで、本発明では、不感帯に入ったときでも線形項が有効に機能するように、以下のような処理を行なう。
即ち、上記の制御量ｕの式に、β・Ｓ（βは定数）を加える。ここで、切換線Ｓ＝０上をスライディングしているときは、β・Ｓ≒０であるので、β・Ｓを制御量ｕに加算してもスライディングに何ら影響はない。
【００８２】
【数１７】

ここで、β’＝βγ、α’＝α＋βとおくと、
【００８３】
【数１８】

この式は、
【００８４】
【数１９】

の形となる。
【００８５】
このように、上記の加算処理を行なった結果、制御量の線形項にＶＴＣのエラー量（ＰＥＲＲ）が含まれることとなり、これにより、動作不感帯に入ったときでも０でない線形項によって切換線への移動速度が適度に与えられ、切換線上での良好なスライディングも確保されるので、目標角度へ応答性よく収束させることができる。
【００８６】
なお、ｃ、ｄの係数は、通常の線形制御系の設計（応答性、安定性より決定）を用いて決める。例えば、ｃは、実際のバルブタイミング制御装置の９０％応答時間及び行き過ぎ量から決定できる。係数ｄも、大きすぎると目標角度に収束せず、ハンチングが発生するので、発散しないように適度の値に設定する。
【００８７】
Ｋは、正の値を設定する。但し、大きすぎるとハンチングの原因になるので、ハンチングが発生しない最大の値を設定する。
７．チャタリング防止の設計
非線形項ＵｎＬ＝−ｋ・Ｓ／｜Ｓ｜＝−ｋｓｇｎ（Ｓ）をデジタル制御器で用いると、サンプリング周期を無限小にできないため、切換面を滑らず、その近傍でチャタリングを起こす。
【００８８】
そこで、飽和関数、平滑関数等を用いてチャタリングの低減を行なう。これらの関数を図示すると、図７に示すようになる。
いずれを使用してもよいが、平滑関数は、飽和関数に比較して演算式が簡単であるので（条件分岐がない）、使用しやすい。
【００８９】
図８は、上記のように設計されたスライディングモード制御を適用した前記コントローラ４８による電磁アクチュエータ５４のデューティ制御の様子を示すブロック図である。
【００９０】
ＶＴＣ目標角度ＶＴＣＴＲＧとＶＴＣ実角度ＶＴＣＮＯＷとの偏差であるエラー量ＰＥＲＲを算出し、該エラー量ＰＥＲＲにＰ分ゲインｃを乗じた比例分制御量Ｕ_Ｐと、ＶＴＣ実角度ＶＴＣＮＯＷの微分値であるＶＴＣ実速度Ｕ_Ｎに速度ゲインｄを乗じた速度制御量Ｕ_Ｎ’を加算して線形項制御量Ｕ_Ｌを算出する。
【００９１】
また、前記エラー量ＰＥＲＲに傾きγを乗じた値と、エラー量ＰＥＲＲの微分値ｄ（ＰＥＲＲ）／ｄｔとを加算して、切換関数Ｓを算出し、該切換関数Ｓを用いた平滑関数−ｋＳ（｜Ｓ｜＋δ）として非線形項制御量Ｕ_ＮＬを算出する。
【００９２】
前記線形項制御量Ｕ_Ｌは、制御系（ＶＴＣ）の状態を切換線（Ｓ＝０）に近づける速さを調整する役割を有し、非線形項制御量Ｕ_ＮＬは、切換線上に沿ったスライディングモードを生じさせる役割を有する。
【００９３】
そして、前記線形項制御量Ｕ_Ｌと、非線形項制御量Ｕ_ＮＬとを加算して、制御量（フィードバック補正分）ＵＤＴＹを算出し、該フィードバック補正分ＵＤＴＹを、前記不感帯中立位置相当のベースデューティ比ＢＡＳＥＤＴＹと加算して該加算結果を最終的なデューティ比ＶＴＣＤＴＹとして出力する。
【００９４】
このように、スライディング制御によってフィードバック補正量を算出して、予め設定された切換線上に制御系の状態を導くようにフィードバックゲインの切換が行なわれるので、油温や油圧などの外乱による影響を受けにくく、ロバスト性の高い制御を行うことができる（図９参照）。
【００９５】
また、特に、制御量の線形項をエラー量の関数として設定することにより、前記切換弁の不感帯を乗り越えるための複雑なディザー制御が不要若しくは依存率を減少させてマッチングを簡略化でき、ＲＯＭやＲＡＭの容量も節約できる。
【００９６】
また、前記不感帯の影響が低減するため、切換弁の不感帯幅の寸法公差が緩められ、加工コストを軽減できる。
また、前記実施の形態では、線形項を前記エラー量に比例した項に加えて、ＶＴＣの動作速度に比例した項を設けて設定したため、動作不感帯から外れたときには該動作速度に比例する項による切換線への移動速度調整機能も加わって、より適切な移動速度に調整され、応答性がより向上する。
【００９７】
しかし、前記制御量の線形項を、エラー量に比例する項のみで設定した場合でも、動作不感帯に入ったときのみならず、動作不感帯から外れたときでも切換線への移動速度が適度に与えられる機能を有するので、スライディングモード制御を実現することができ、この場合は動作速度に比例する項を省略したことにより、制御（演算）が簡易となってプログラム容量を節約できる（図１０参照）。
【００９８】
なお、本発明は、前記ベーン式の油圧アクチュエータを用いたＶＴＣに限らず、例えば、リニア式の油圧アクチュエータを用いて直線運動を回転運動に変換してカムシャフトの回転位相を可変するようなＶＴＣにも適用できることは勿論であり、動作不感帯を有する制御対象であれば油圧制御式に限るものでもない。
【図面の簡単な説明】
【図１】実施の形態におけるバルブタイミング制御機構を示す断面図。
【図２】図１のＢ−Ｂ断面図。
【図３】上記バルブタイミング制御機構の分解斜視図。
【図４】上記バルブタイミング制御機構における電磁切換弁を示す縦断面図。
【図５】上記バルブタイミング制御機構における電磁切換弁を示す縦断面図。
【図６】上記バルブタイミング制御機構における電磁切換弁を示す縦断面図。
【図７】スライディングモード制御の非線形項制御量に使用される関数の形態を示す図。
【図８】上記バルブタイミング制御機構の制御ブロック図。
【図９】上記バルブタイミング制御機構のスライディングモード制御時の目標角度への収束の様子を示すタイムチャート。
【図１０】別の実施の形態におけるバルブタイミング制御機構の制御ブロック図。
【符号の説明】
２…カムシャフト
４…油圧回路
３２…進角側油圧室
３３…遅角側油圧室
４５…電磁切換弁
４７…オイルポンプ
５３…スプール弁体
１０１…回転センサ
１０２…エアフローメータ
１０３…クランク角センサ
１０４…カムセンサ[0001]
BACKGROUND OF THE INVENTION
The present invention relates to an apparatus for performing sliding mode control on a controlled object having an operation dead band with respect to a controlled variable such as a hydraulic control system.
[0002]
[Prior art]
Conventionally, as a valve timing control device configured to continuously and variably control the rotational phase of a camshaft with respect to a crankshaft of an internal combustion engine by hydraulic control using a switching valve, a vane type valve as disclosed in JP-A-10-141022 There is a timing control device.
[0003]
In this structure, a concave portion is formed on the inner peripheral surface of a cylindrical housing fixed to the cam sprocket, while a vane portion of the impeller fixed to the camshaft is accommodated in the concave portion. The camshaft can be rotated relative to the cam sprocket within a range in which the wing portion can move.
[0004]
And the said wing | blade part hold | maintains the said wing | blade part in the intermediate position of the said recessed part by supplying and discharging oil relatively with respect to a pair of hydraulic chamber formed by dividing the said recessed part into the front and back of a rotation direction. When the hydraulic pressure in the pair of hydraulic chambers is adjusted to a hydraulic pressure at which a target rotational phase is obtained, the hydraulic passage is closed by a switching valve and the oil pressure of the pair of hydraulic chambers is adjusted. It is configured to stop supply and discharge.
[0005]
[Problems to be solved by the invention]
By the way, in the hydraulic control system as described above, the hydraulic passage for supplying and discharging oil in the hydraulic chamber is closed by a switching valve so as to prevent oil leakage from the hydraulic chamber due to the reaction force from the intake / exhaust valves in a steady state. The closing margin of the valve body of the switching valve (spool valve) is set large. For this reason, the amount of oil in the hydraulic chamber and thus the rotational phase of the camshaft to be controlled has an operation dead zone with respect to the control amount of the switching valve (duty ratio or the like in the electromagnetic drive type).
[0006]
As a control method of the camshaft rotational phase, PID control or the like is generally employed, but it is difficult to perform feedback control with good response to the dead zone only by the PID control. For this reason, some dithers are added to perform dither control separately from PID. However, it is necessary to perform addition determination for dithers based on the amount of error, which is complicated control, and ROM or RAM In order to secure capacity and control accuracy by reducing the variation of the dead zone width for each component, it is necessary to increase the processing accuracy of the component and increase the processing cost.
[0007]
In order to execute the PID control with high responsiveness, it is desirable to variably set the feedback gain because the viscosity of the oil changes according to the oil temperature and oil pressure. However, it is not easy to match the settings.
[0008]
On the other hand, in recent years, sliding mode control has attracted attention as feedback control with high robustness and little influence on disturbance.
Therefore, it is conceivable to apply sliding mode control to the hydraulic control type valve timing control device, and when sliding mode control is designed according to the textbook, the influence of disturbances such as changes in the oil temperature and hydraulic pressure is suppressed. However, it has been found that it does not function effectively for the dead zone and does not replace or assist the dither control.
[0009]
The present invention has been made by paying attention to such a conventional problem, and is a robust device that suppresses a decrease in responsiveness due to the dead zone with respect to a control target having an operation dead zone with respect to a control amount of a hydraulic control system or the like. The purpose is to enable high sliding mode control.
[0010]
[Means for Solving the Problems]
For this reason, the invention according to claim 1
A device for setting a feedback correction amount of a control amount of a variable valve mechanism whose valve characteristics are variably controlled by adding a linear term and a nonlinear term, and calculating the nonlinear term by sliding mode control,
The linear term for the feedback correction is the control target that determines the valve characteristics.It is set as a function of the deviation between the target position and the actual position.
[0011]
According to the invention of claim 1, in the sliding mode controlThe linear term for the feedback correction is the control target that determines the valve characteristics.It is set as a function of the deviation between the target position and the actual position.
[0012]
As a result, even when the operation dead zone is entered and the actual position of the controlled object does not change, there is a deviation (≠ 0) between the target position and the actual position. Therefore, the linear term set as a function of the deviation The moving speed to the switching line for switching the feedback gain is appropriately given, so that the target position can be converged with good responsiveness.
[0013]
The invention according to claim 2
SaidThe variable valve mechanismIt is a hydraulic control system.
According to the invention of claim 2,
The sliding mode control according to the present invention is applied to the hydraulic control system.
[0014]
As a result, in the hydraulic control system having a large operation dead zone due to the switching valve or the like, control with good responsiveness that suppresses the influence of the dead zone is executed.
The invention according to claim 3
The variable valve mechanism isThe configuration is such that the rotational phase of the camshaft with respect to the crankshaft is continuously variably controlled by hydraulic control, and is controlled by selectively controlling the supply and discharge of oil to and from the hydraulic actuator that is hydraulically controlled. It is a valve timing control device for an internal combustion engine.
[0015]
According to the invention of claim 3,
The sliding mode control according to the present invention is applied to the valve timing control device for an internal combustion engine having the above configuration.
[0016]
As a result, in the valve timing control device for a hydraulically controlled internal combustion engine having a large operation dead zone due to the switching valve, control with good responsiveness, in which the influence of the dead zone is suppressed, is executed.
[0017]
The invention according to claim 4
SaidFeedback correctionThe linear term is set by adding a term proportional to the deviation between the target position of the controlled object and the actual position and a term proportional to the operating speed of the controlled object.
[0018]
According to the invention of claim 4,
The linear term of the control amount in the sliding mode control is set by adding a term proportional to the deviation between the target position of the control target and the actual position and a term proportional to the operation speed of the control target.
[0019]
As a result, the term proportional to the deviation between the target position to be controlled and the actual position has a function that allows the moving speed to the switching line to be given even when the operation dead zone is entered, and deviates from the operation dead zone. In this case, a function for adjusting the moving speed to the switching line by a term proportional to the operating speed of the controlled object is also added, so that the moving speed is adjusted to a more appropriate speed and the responsiveness is further improved.
[0020]
The invention according to claim 5
SaidFeedback correctionThe linear term is set only by a term proportional to the deviation between the target position to be controlled and the actual position.
[0021]
According to the invention of claim 5,
The linear term of the control amount in the sliding mode control is set only by a term proportional to the deviation between the target position to be controlled and the actual position.
[0022]
Even when the linear term is set only by a term proportional to the deviation between the target position to be controlled and the actual position, the moving speed to the switching line is not only when entering the motion dead zone but also when moving outside the motion dead zone. Since it has an appropriate function, sliding mode control can be realized, and the control (calculation) is simplified by omitting the term proportional to the operation speed of the controlled object (in claim 4). The program capacity can be saved.
[0023]
DETAILED DESCRIPTION OF THE INVENTION
Embodiments of the present invention will be described below.
FIGS. 1-6 shows the mechanism part of the valve timing control apparatus of the internal combustion engine in embodiment, and shows what was applied to the intake valve side.
[0024]
The valve timing control device shown in the figure is provided so as to be rotatable relative to the cam sprocket 1 (timing sprocket) that is rotationally driven via a timing chain by an engine crankshaft (not shown). Camshaft 2, rotating member 3 fixed to the end of camshaft 2 and rotatably accommodated in cam sprocket 1, and hydraulic circuit for rotating the rotating member 3 relative to cam sprocket 1 4 and a lock mechanism 10 that selectively locks the relative rotational position of the cam sprocket 1 and the rotating member 3 at a predetermined position.
[0025]
The cam sprocket 1 includes a rotating portion 5 having a tooth portion 5a meshing with a timing chain (or timing belt) on the outer periphery, and a housing 6 disposed in front of the rotating portion 5 and rotatably accommodating the rotating member 3. A disc-shaped front cover 7 serving as a lid for closing the front end opening of the housing 6, and a substantially disc-shaped rear cover disposed between the housing 6 and the rotating portion 5 to close the rear end of the housing 6. 8, and the rotating portion 5, the housing 6, the front cover 7, and the rear cover 8 are integrally coupled from the axial direction by four small-diameter bolts 9.
[0026]
The rotating portion 5 has a substantially annular shape, and has four female screw holes 5b through which the small-diameter bolts 9 are screwed at equal circumferential positions of about 90 ° in the circumferential direction. A step-diameter fitting hole 11 into which a passage-forming sleeve 25 described later is fitted is formed through. Further, a disc-shaped fitting groove 12 into which the rear cover 8 is fitted is formed on the front end surface.
[0027]
The housing 6 has a cylindrical shape in which both front and rear ends are formed with openings, and four partition walls 13 project from the circumferential position of the inner peripheral surface at 90 °. The partition wall 13 has a trapezoidal shape in cross section, is provided along the axial direction of the housing 6, and both end edges are flush with the both end edges of the housing 6. Four bolt insertion holes 14 through which the small-diameter bolts 9 are inserted are formed penetrating in the axial direction. Further, a U-shaped seal member 15 and a leaf spring 16 that presses the seal member 15 inward are fitted into a holding groove 13a that is cut out along the axial direction at the center position of the inner end face of each partition wall 13. Is retained.
[0028]
Further, the front cover 7 has a relatively large-diameter bolt insertion hole 17 at the center, and four bolt holes 18 at positions corresponding to the bolt insertion holes 14 of the housing 6. ing.
[0029]
The rear cover 8 has a disc portion 8a fitted and held in the fitting groove 12 of the rotating member 5 on the rear end surface, and a small-diameter annular portion 25a of the sleeve 25 is fitted in the center. A fitting hole 8c is formed, and four bolt holes 19 are also formed at positions corresponding to the bolt insertion holes 14.
[0030]
The camshaft 2 is rotatably supported at the upper end portion of the cylinder head 22 via a cam bearing 23, and a cam (not shown) for opening the intake valve via a valve lifter is integrated at a predetermined position on the outer peripheral surface. In addition, a flange portion 24 is integrally provided at the front end portion.
[0031]
The rotating member 3 is fixed to the front end portion of the camshaft 2 by a fixing bolt 26 inserted from the axial direction through the sleeve 25 whose front and rear portions are fitted in the flange portion 24 and the fitting hole 11, respectively. An annular base portion 27 having a bolt insertion hole 27a through which the fixing bolt 26 is inserted, and four vanes 28a, 28b, 28c, 28d integrally provided at 90 ° positions in the circumferential direction of the outer peripheral surface of the base portion 27; It has.
[0032]
The first to fourth vanes 28a to 28d each have a substantially inverted trapezoidal cross section, and are disposed in the recesses between the partition walls 13, and the recesses are separated in the front and rear in the rotation direction. An advance side hydraulic chamber 32 and a retard side hydraulic chamber 33 are formed between both sides and both side surfaces of each partition wall portion 13. In addition, a U-shaped seal member 30 slidably contacting the inner peripheral surface 6a of the housing 6 and the seal member 30 are pressed outwardly into a holding groove 29 cut in the axial direction at the center of the outer peripheral surface of each of the vanes 28a to 28d. The leaf springs 31 are fitted and held.
[0033]
The lock mechanism 10 is formed through an engagement groove 20 formed at a predetermined position on the outer peripheral side of the fitting groove 12 of the rotating portion 5 and a predetermined position of the rear cover 8 corresponding to the engagement groove 20. An engagement hole 21 whose inner peripheral surface is tapered, a sliding hole 35 formed through the substantially central position of the one vane 28 corresponding to the engagement hole 21 along the internal axis direction, and the one A lock pin 34 slidably provided in the sliding hole 35 of the vane 28, a coil spring 39 that is a spring member elastically mounted on the rear end side of the lock pin 34, and the lock pin 34 and the sliding hole 35 is formed with a pressure receiving chamber 40 formed between them.
[0034]
The lock pin 34 includes a middle-side main body 34a on the center side, an engaging portion 34b formed in a substantially tapered shape on the front end side of the main body 34a, and a large step formed on the rear end side of the main body 34a. The engaging hole 21 is formed by the spring force of the coil spring 39 elastically mounted between the bottom surface of the internal concave groove 34d of the stopper portion 34c and the inner end surface of the front cover 7. And the hydraulic pressure in the pressure receiving chamber 40 formed between the outer peripheral surface between the main body 34a and the stopper portion 34c and the inner peripheral surface of the sliding hole 35. It slides in the direction of coming out of the engagement hole 21. The pressure receiving chamber 40 communicates with the retard angle side hydraulic chamber 33 through a through hole 36 formed in a side portion of the vane 28. Further, the engaging portion 34 b of the lock pin 34 is configured such that the engaging portion 34 b is engaged with the engaging hole 21 at the rotation position on the maximum retard angle side of the rotating member 3.
[0035]
The hydraulic circuit 4 includes two systems, a first hydraulic passage 41 that supplies and discharges hydraulic pressure to the advance side hydraulic chamber 32 and a second hydraulic passage 42 that supplies and discharges hydraulic pressure to the retard side hydraulic chamber 33. The hydraulic passages 41 and 42 are connected to a supply passage 43 and a drain passage 44 through passage-switching electromagnetic switching valves 45, respectively. The supply passage 43 is provided with an oil pump 47 that pumps the oil in the oil pan 46, while the downstream end of the drain passage 44 communicates with the oil pan 46.
[0036]
The first hydraulic passage 41 is branched and formed in the head portion 26a from the cylinder head 22 through the first passage portion 41a formed in the axial center of the camshaft 2 and the axial direction in the fixing bolt 26. A first oil passage 41b that communicates with the first passage portion 41a, a small-diameter outer peripheral surface of the head portion 26a, and an inner peripheral surface of the bolt insertion hole 27a that is provided in the base portion 27 of the rotating member 3 are the first. An oil chamber 41c that communicates with the oil passage 41b, and four branch passages 41d that are formed substantially radially in the base portion 27 of the rotating member 3 and communicate with the oil chamber 41c and each advance-side hydraulic chamber 32. Yes.
[0037]
On the other hand, the second hydraulic passage 42 is formed into a second passage portion 42 a formed in the cylinder head 22 and on one side of the camshaft 2, and is bent into a substantially L shape inside the sleeve 25. A second oil passage 42b communicating with the passage portion 42a, four oil passage grooves 42c formed at the outer peripheral side edge of the fitting hole 11 of the rotating member 5 and communicating with the second oil passage 42b, and the periphery of the rear cover 8. The four oil holes 42 d are formed at positions of about 90 ° in the direction and communicate with each oil passage groove 42 c and the retard side hydraulic chamber 33.
[0038]
The electromagnetic switching valve 45 is configured such that an internal spool valve body switches and controls each of the hydraulic passages 41 and 42, the supply passage 43, and the drain passages 44a and 44b, and a control signal from the controller 48. It is designed to be switched by.
[0039]
Specifically, as shown in FIGS. 4 to 6, a cylindrical valve body 51 inserted and fixed in the holding hole 50 of the cylinder block 49 and a valve hole 52 in the valve body 51 are slidable. The spool valve body 53 is provided to switch the flow path, and a proportional solenoid type electromagnetic actuator 54 that operates the spool valve body 53.
[0040]
The valve body 51 is formed with a supply port 55 penetrating the downstream end of the supply passage 43 and the valve hole 52 at a substantially central position of the peripheral wall, and the first and the second on both sides of the supply port 55. A first port 56 and a second port 57 that communicate with the other end of the second hydraulic passages 41 and 42 and the valve hole 52 are formed penetratingly. Further, third and fourth ports 58 and 59 are formed through both ends of the peripheral wall so as to communicate the drain passages 44a and 44b with the valve hole 52.
[0041]
The spool valve body 53 has a substantially cylindrical first valve portion 60 that opens and closes the supply port 55 at the center of the small diameter shaft portion, and opens and closes the third and fourth ports 58 and 59 at both ends. It has substantially cylindrical second and third valve portions 61 and 62. The spool valve body 53 is a conical valve spring 63 elastically mounted between an umbrella portion 53b provided at one end edge of the support shaft 53a on the front end side and a spring seat 51a provided on the inner peripheral wall of the front end side of the valve hole 52. Therefore, the supply valve 55 and the second hydraulic passage 42 are urged in the right direction in FIG.
[0042]
The electromagnetic actuator 54 includes a core 64, a moving plunger 65, a coil 66, a connector 67, and the like, and a driving rod 65 a that presses the umbrella portion 53 b of the spool valve body 53 is fixed to the tip of the moving plunger 65.
[0043]
The controller 48 detects the current operating state (load, rotation) based on signals from the rotation sensor 101 that detects the engine rotation speed and the air flow meter 102 that detects the intake air amount, and from the crank angle sensor 103 and the cam sensor 104. The relative rotation position of the cam sprocket 1 and the camshaft 2, that is, the rotational phase of the camshaft 2 with respect to the crankshaft is detected by the above signal.
[0044]
The controller 48 controls the energization amount for the electromagnetic actuator 54 based on a duty control signal.
For example, when a control signal (OFF signal) with a duty ratio of 0% is output from the controller 48 to the electromagnetic actuator 54, the spool valve body 53 moves to the position shown in FIG. . As a result, the first valve portion 60 opens the open end 55a of the supply port 55 to communicate with the second port 57, and at the same time, the second valve portion 61 opens the open end of the third port 58, and the fourth The valve part 62 closes the fourth port 59. Therefore, the hydraulic oil pressure-fed from the oil pump 47 is supplied to the retarded hydraulic chamber 33 through the supply port 55, the valve hole 52, the second port 57, and the second hydraulic passage 42, and at the advanced side. The hydraulic oil in the hydraulic chamber 32 is discharged from the first drain passage 44a into the oil pan 46 through the first hydraulic passage 41, the first port 56, the valve hole 52, and the third port 58.
[0045]
Therefore, the internal pressure of the retard side hydraulic chamber 33 is high and the internal pressure of the advance side hydraulic chamber 32 is low, and the rotating member 3 rotates in one direction at the maximum via the vanes 28a to 28b. As a result, the cam sprocket 1 and the camshaft 2 rotate relative to one side to change the phase. As a result, the opening timing of the intake valve is delayed and the overlap with the exhaust valve is reduced.
[0046]
On the other hand, when a control signal (ON signal) with a duty ratio of 100% is output from the controller 48 to the electromagnetic actuator 54, the spool valve element 53 is maximally leftward against the spring force of the valve spring 63 as shown in FIG. At the same time as the third valve portion 61 closes the third port 58 by sliding, the fourth valve portion 62 opens the fourth port 59, and the first valve portion 60 includes the supply port 55 and the first port 56. To communicate with. Therefore, the hydraulic oil is supplied into the advance side hydraulic chamber 32 through the supply port 55, the first port 56, and the first hydraulic passage 41, and the hydraulic oil in the retard side hydraulic chamber 33 is second. The oil is discharged to the oil pan 46 through the hydraulic passage 42, the second port 57, the fourth port 59, and the second drain passage 44b, and the retard side hydraulic chamber 33 becomes low pressure.
[0047]
For this reason, the rotating member 3 rotates to the maximum in the other direction via the vanes 28a to 28d, whereby the cam sprocket 1 and the camshaft 2 are relatively rotated to the other side and the phase is changed. The opening timing of the intake valve is advanced (advanced), and the overlap with the exhaust valve is increased.
[0048]
The controller 48 has a position in which the first valve portion 60 closes the supply port 55, the third valve portion 61 closes the third port 58, and the fourth valve portion 62 closes the fourth port 59. The relative duty position (rotation phase) between the cam sprocket 1 and the camshaft 2 detected based on signals from the crank angle sensor 103 and the cam sensor 104, and the operation state A feedback correction amount UDTY for matching the target value (target advance value) of the relative rotation position (rotation phase) set in accordance with is set by sliding mode control as will be described later, and the base duty ratio BASEDTY is set. The final duty ratio VTCDTY is obtained by adding the feedback correction amount UDTY and the duty ratio. A control signal TCDTY are to be output to the electromagnetic actuator 54. The base duty ratio BASEDTY is a substantially median value of the duty ratio range in which the supply port 55, the third port 58, and the fourth port 59 are all closed and no oil is supplied or discharged in any of the hydraulic chambers 32 and 33. (For example, 50%).
[0049]
That is, when it is necessary to change the relative rotation position (rotation phase) in the retarding direction, the duty ratio is reduced by the feedback correction UDTY, and the hydraulic oil pumped from the oil pump 47 is retarded. While being supplied to the hydraulic chamber 33, the hydraulic oil in the advance side hydraulic chamber 32 is discharged into the oil pan 46. Conversely, the relative rotation position (rotation phase) is changed in the advance direction. When it is necessary to increase the duty ratio, the duty ratio is increased by the feedback correction UDTY, the hydraulic oil is supplied into the advance hydraulic chamber 32, and the hydraulic oil in the retard hydraulic chamber 33 is supplied to the oil pan 46. Will be discharged. When the relative rotation position (rotation phase) is maintained in the current state, the absolute value of the feedback correction UDTY is decreased to return to a duty ratio near the base duty ratio. The internal pressures of the hydraulic chambers 32 and 33 are controlled by closing the 55, the third port 58, and the fourth port 59 (stopping the supply and discharge of hydraulic pressure).
[0050]
Here, the feedback correction amount UDTY is calculated as follows by sliding mode control. In the following description, the detected relative rotation position (rotation phase) between the cam sprocket 1 and the camshaft 2 will be described as the actual angle of the valve timing control device (VTC), and its target value will be described as the VTC target angle.
[0051]
1. Mathematical model calculation
In the sliding mode control, the controller parameters are determined by the mathematical model to be controlled. Therefore, the VTC mathematical model is first calculated.
[0052]
The mathematical model can be obtained by, for example, establishing an equation of motion or obtaining the system by identifying the system. Here, system identification is used.
As a result of system identification when input u (k): duty and output y (k): VTC actual angle, the following transfer function was obtained.
[0053]
G (s) = b / (s²+ A₂・ S + a₁)
2. Simplification of transfer function
The model obtained by system identification may be a higher-order model, and the transfer function is simplified in order to simplify the controller configuration.
[0054]
G (s) = b / [s (s + a₂]] …… (2.1)
3. Calculation of equation of state
From the obtained transfer function, the differential equation of VTC is given as follows. Where x: actual angle of VTC, u: input (duty)
[0055]
[Expression 1]

The equation of state is
[0056]
[Expression 2]

Therefore, substituting the differential equation (3.1) into (3.2) gives the following.
[0057]
[Equation 3]

4). Switching function design
In the sliding mode control, the feedback gain is switched depending on the system state, so this switching function S is set as follows.
[0058]
[Expression 4]

Since the sliding mode may not occur depending on the parameters of the switching function, the design of the switching function is very important. There are mainly the following design methods.
[0059]
(1) Design method using pole placement method
(2) Optimum switching hyperplane design method
(3) Design method using zero of system
(4) Hyperplane design method by frequency shaping
α₁, Α₂Is obtained by applying the above design method, and α₁: Α₂When γ satisfying γ = 1 is obtained, it is as follows.
[0060]
[Equation 5]

However, as described above, the switching function designed according to a normal textbook is a function of the actual position of the control object, that is, the actual angle x of the VTC. It becomes inappropriate.
[0061]
First, with regard to the term of γx, when the target angle of VTC is other than 0 °, a positive value is always given and becomes an irrelevant value between the target angle and the actual angle, so that VTC does not converge to the target angle.
[0062]
As for the term dx / dt, when the electromagnetic switching valve is in the dead zone, the VTC does not operate, so the actual speed dx / dt does not change, and the response is poor when it is desired to operate only at a minute angle.
[0063]
In addition, in the design according to the textbook, it is recommended to add an integral term for the error amount, but when the camshaft is set to the target angle, the integral term for the error amount is left as a non-zero value. It will function to prevent convergence to the target angle.
[0064]
Therefore, the switching function S is set as a function of the following error amount.
[0065]
[Formula 6]

Here, the design method using the zero of the system of (3) was used for designing the switching function. The zero of the system is a method of setting the zero of (S, A, B) to the left half of the complex plane (S: switching function, A, B: constant of equation (3.2)).
[0066]
5). Calculation of sliding conditions
The simplest condition for achieving sliding is S · dS / dt <0.
The above condition is satisfied only when S decreases. Since S is a variable of the error and the differential value of the error, it means that when the above condition is satisfied, the error decreases and converges to the target value.
[0067]
First, an expression necessary for the expansion of S is obtained.
The control amount u is set as follows.
[0068]
[Expression 7]

Substituting this into equation (3.1),
[0069]
[Equation 8]

Next, the hat u is developed.
[0070]
Since hat u is an input when sliding, S = dS / dt = 0.
[0071]
[Equation 9]

If bu = hat u,
[0072]
[Expression 10]

Consider the sliding condition S · dS / dt <0.
[0073]
## EQU11 ##

From formulas (5.2) and (5.3)
[0074]
[Expression 12]

Therefore, if k is a positive value, sliding is established.
[0075]
6). Control amount calculation formula design
The control amount (feedback correction amount) u is as follows from equations (5.1) and (5.3).
[0076]
[Formula 13]

Using equation (2.1), which simplified the VTC transfer function, the equation of state is as follows.
[0077]
[Expression 14]

Using the state equation of (6.2), the equation (6.1) becomes as follows.
[0078]
[Expression 15]

Where α = b^-1(A−γ), k ′ = b^-1If k is set,
[0079]
[Expression 16]

This equation is an equation that guarantees movement while sliding on the switching line S = 0.
[0080]
However, in the control amount equation designed in this way (as in the textbook), the linear term α · dx / dt is not supplied or discharged when it is in the dead zone, so that the operation speed dx / dt = Since 0 → linear term = 0, it does not function effectively.
[0081]
Therefore, in the present invention, the following processing is performed so that the linear term functions effectively even when the dead zone is entered.
That is, β · S (β is a constant) is added to the equation of the control amount u. Here, when sliding on the switching line S = 0, since β · S≈0, adding β · S to the control amount u has no effect on the sliding.
[0082]
[Expression 17]

Here, if β ′ = βγ and α ′ = α + β,
[0083]
[Expression 18]

This formula is
[0084]
[Equation 19]

It becomes the form.
[0085]
As described above, as a result of performing the above addition processing, the VTC error amount (PERR) is included in the linear term of the control amount, so that even when the operation dead zone is entered, a linear term that is not 0 leads to the switching line. Is moderately given, and good sliding on the switching line is ensured, so that the target angle can be converged with good responsiveness.
[0086]
The coefficients c and d are determined by using a normal linear control system design (determined from responsiveness and stability). For example, c can be determined from the 90% response time and overshoot of the actual valve timing controller. If the coefficient d is too large, it does not converge to the target angle and hunting occurs. Therefore, the coefficient d is set to an appropriate value so as not to diverge.
[0087]
K is set to a positive value. However, if it is too large, it will cause hunting, so set the maximum value that does not cause hunting.
7). Anti-chattering design
If the non-linear term UnL = −k · S / | S | = −ksgn (S) is used in the digital controller, the sampling period cannot be made infinitely small, so that the switching surface does not slip and chattering occurs in the vicinity thereof.
[0088]
Therefore, chattering is reduced using a saturation function, a smooth function, or the like. These functions are illustrated in FIG.
Any of these may be used, but the smooth function is easy to use because it has a simpler arithmetic expression than the saturation function (no conditional branching).
[0089]
FIG. 8 is a block diagram showing a state of duty control of the electromagnetic actuator 54 by the controller 48 to which the sliding mode control designed as described above is applied.
[0090]
An error amount PERR that is a deviation between the VTC target angle VTCTRG and the VTC actual angle VTCNOW is calculated, and a proportional control amount U obtained by multiplying the error amount PERR by a P-component gain c._PVTC actual speed U which is a differential value of VTC actual angle VTCNOW_NSpeed control amount U multiplied by speed gain d_N′ Is added to the linear term control amount U_LIs calculated.
[0091]
Further, a value obtained by multiplying the error amount PERR by a slope γ and a differential value d (PERR) / dt of the error amount PERR is calculated to calculate a switching function S, and a smoothing function using the switching function S− kS (| S | + δ) as a nonlinear term control amount U_NLIs calculated.
[0092]
The linear term control amount U_LHas a role of adjusting the speed at which the state of the control system (VTC) is brought close to the switching line (S = 0), and the nonlinear term control amount U_NLHas a role of generating a sliding mode along the switching line.
[0093]
And the linear term control amount U_LAnd the nonlinear term control amount U_NLIs added to calculate the control amount (feedback correction amount) UDTY, and the feedback correction amount UDTY is added to the base duty ratio BASEDTY corresponding to the dead zone neutral position, and the addition result is used as the final duty ratio VTCDTY. Output.
[0094]
In this way, the feedback correction amount is calculated by the sliding control, and the feedback gain is switched so as to guide the state of the control system on the preset switching line, so that it is affected by disturbances such as oil temperature and hydraulic pressure. It is difficult to perform control with high robustness (see FIG. 9).
[0095]
In particular, by setting the linear term of the control amount as a function of the error amount, complicated dither control for overcoming the dead zone of the switching valve is unnecessary or the dependency rate can be reduced, and matching can be simplified. RAM capacity can also be saved.
[0096]
In addition, since the influence of the dead zone is reduced, the dimensional tolerance of the dead zone width of the switching valve is relaxed, and the processing cost can be reduced.
Further, in the above embodiment, since the linear term is set by setting a term proportional to the VTC operating speed in addition to the term proportional to the error amount, the term proportional to the operating speed is used when it is out of the operation dead zone. The function of adjusting the moving speed to the switching line is also added to adjust the moving speed to a more appropriate moving speed, and the responsiveness is further improved.
[0097]
However, even when the linear term of the control amount is set only with a term proportional to the error amount, not only when entering the operation dead zone, but also when moving out of the operation dead zone, the moving speed to the switching line is given moderately. Therefore, the sliding mode control can be realized. In this case, by omitting the term proportional to the operation speed, the control (calculation) is simplified and the program capacity can be saved (see FIG. 10). .
[0098]
The present invention is not limited to the VTC using the vane type hydraulic actuator. For example, a VTC that changes the rotational phase of the camshaft by converting linear motion into rotational motion using a linear hydraulic actuator. Needless to say, the present invention is not limited to the hydraulic control type as long as it is a control target having an operation dead zone.
[Brief description of the drawings]
FIG. 1 is a cross-sectional view showing a valve timing control mechanism in an embodiment.
2 is a cross-sectional view taken along the line BB in FIG.
FIG. 3 is an exploded perspective view of the valve timing control mechanism.
FIG. 4 is a longitudinal sectional view showing an electromagnetic switching valve in the valve timing control mechanism.
FIG. 5 is a longitudinal sectional view showing an electromagnetic switching valve in the valve timing control mechanism.
FIG. 6 is a longitudinal sectional view showing an electromagnetic switching valve in the valve timing control mechanism.
FIG. 7 is a diagram illustrating a form of a function used for a nonlinear term control amount in sliding mode control.
FIG. 8 is a control block diagram of the valve timing control mechanism.
FIG. 9 is a time chart showing how the valve timing control mechanism converges to a target angle during sliding mode control.
FIG. 10 is a control block diagram of a valve timing control mechanism in another embodiment.
[Explanation of symbols]
2 ... Camshaft
4 ... Hydraulic circuit
32 ... Advance side hydraulic chamber
33 ... retarded-side hydraulic chamber
45 ... Electromagnetic switching valve
47 ... Oil pump
53 ... Spool valve body
101 ... Rotation sensor
102 ... Air flow meter
103 ... Crank angle sensor
104: Cam sensor

Claims (5)

A device for setting a feedback correction amount of a control amount of a variable valve mechanism whose valve characteristics are variably controlled by adding a linear term and a non-linear term calculated using sliding mode control,
A sliding mode control apparatus, wherein the linear term for the feedback correction is set as a function of a deviation between a target position to be controlled and a real position for determining valve characteristics . The sliding mode control device according to claim 1, wherein the variable valve mechanism is a hydraulic control system. The variable valve mechanism is configured to continuously and variably control the rotation phase of the camshaft with respect to the crankshaft by hydraulic control, and selectively controls oil supply / discharge of the hydraulically controlled hydraulic actuator by a switching valve. The sliding mode control device according to claim 2, wherein the control device is a valve timing control device for an internal combustion engine configured to perform control. Claims, characterized in that the linear term of the feedback correction amount, a term which is proportional to the deviation between the actual position and the position of the target of the controlled object, and set by adding the term proportional to the operating speed of the control object The sliding mode control device according to any one of claims 1 to 3 . The sliding mode according to any one of claims 1 to 3, wherein the linear term for the feedback correction is set only by a term proportional to a deviation between a target position to be controlled and an actual position. Control device.

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