JP4161645B2

JP4161645B2 - Vibration reduction device for internal combustion engine

Info

Publication number: JP4161645B2
Application number: JP2002250313A
Authority: JP
Inventors: 康之浅原
Original assignee: Nissan Motor Co Ltd
Current assignee: Nissan Motor Co Ltd
Priority date: 2002-08-29
Filing date: 2002-08-29
Publication date: 2008-10-08
Anticipated expiration: 2022-08-29
Also published as: JP2004084644A

Description

【０００１】
【発明の属する技術分野】
本発明は、内燃機関の振動低減装置に関する。
【０００２】
【従来の技術】
フライホイールと結合されたクランクシャフトの回転駆動力を伝える駆動力伝達機構と、前記駆動力伝達機構により回転させられて慣性力を生じる副慣性質量体とを備え、且つ、前記駆動力伝達機構に弾性体を持たせて振動系を形成し、この振動系の反共振の周波数を、前記内燃機関の運転状態のうち、略一定回転で運転される運転状態での回転のｎ／２（ｎ＝自然数）倍の周波数と略一致させることにより、その運転状態での内燃機関の振動を低減するようにした内燃機関の振動低減装置が従来より知られている（特開平１１−３２５１８６号公報）。
【０００３】
また、特開平１１−０４４２３１号公報には、モータージェネレーターを内燃機関の回転方向と逆方向に回転させることにより、モータージェネレーターのローター部の慣性力による反力トルクにより内燃機関のトルク変動を打ち消しながら、更にモータージェネレーターから内燃機関のトルク変動を打ち消すトルク変動を発生させて内燃機関のトルク変動に伴う振動を低減させる技術が開示されている。
【０００４】
【発明が解決しようとする課題】
しかしながら、前者の従来例（特開平１１−３２５１８６号公報の記載の従来技術）では、反共振の周波数近傍では大きな振動低減効果が得られるものの、その周波数を外れると効果が小さくなってしまう。つまり、反共振周波数と回転のｎ／２（ｎ＝自然数）倍の周波数が略一致するように設定されている運転条件から、回転数が少しずれただけでも低減効果が小さくなってしまう。
例えば、アイドル運転で反共振の効果が得られるように反共振周波数を設定した場合、エアコン、オルタネーター等の補機負荷の状態や、Ａ／ＴのＮレンジ、Ｄレンジのセレクト状態等によってアイドル運転といってもその運転条件が異なるため、回転数を変えたほうが良い場合があるが、前者の従来例の場合には最大の振動低減効果を得るためには常にアイドル回転数を同じに保たなければならない。また、反共振周波数より少し高い周波数で共振が起きるため、それによる振動悪化を伴うという問題点がある。
また、後者の従来例（特開平１１−０４４２３１号公報）では、モータージェネレーターの駆動にベルトを用いた場合、ベルトが比較的柔らかいためそれがばねとなり、そこにつながるクランクシャフトと一体回転するフライホイールと、モータージェネレーターのローター部がマスとなる振動系が構成される。そしてこの振動系の共振周波数の前後、及びそれより高い周波数ではフライホイールとモータージェネレーターの位相が一致しないため、モータージェネレータから単に逆位相のトルクを発生させるだけでは内燃機関のトルク変動を低減することが出来なくなる。つまり比較的低回転では内燃機関のトルク変動を低減することが出来るが、回転基本次数（回転周波数×気筒数／２）が共振周波数となる回転数の前後及びそれ以上の回転数ではトルク変動を低減することが出来なくなってしまうという問題点がある。
【０００５】
【課題を解決するための手段】
そこで、本発明の振動低減装置は、内燃機関の駆動力を回転駆動力として伝達するクランクシャフトと、このクランクシャフトと一体回転するフライホイールと、前記クランクシャフトの回転に伴って回転する副慣性質量体と、クランクシャフトの回転駆動力を弾性体を介して前記副慣性質量体に伝達して該副慣性質量体を回転駆動する駆動力伝達機構と、前記内燃機関の回転速度を検出する手段と、を備え、前記駆動力伝達機構のばね作用によって生じる振動系と内燃機関のロール振動との反共振の周波数が、前記内燃機関の所定回転速度における回転周波数の（自然数／２）倍した周波数と略一致するようにした内燃機関の振動低減装置において、前記副慣性質量体が前記駆動力伝達機構により伝達されたクランクシャフトの回転駆動力を用いて発電可能なアクチュエータとなっており、かつそのアクチュエータの駆動トルクを前記内燃機関の運転状態に応じて制御することにより前記反共振の周波数及び前記反共振の減衰の大きさのうち少なくとも一方を変更可能なアクチュエータ駆動トルク制御手段を備えている。
【０００６】
【発明の効果】
本発明によれば、振動系の反共振による効果と共に、アクチュエータにより振動を打ち消すトルクを発生させることにより、より大きな振動低減効果を得ることができる。
【０００７】
【発明の実施の形態】
以下、本発明の一実施例を図面に基づいて詳細に説明する。
【０００８】
まず、本発明の第１実施例を図１〜図３に基づいて説明する。
図１及び図２は、エンジン１とモータージェネレーター２,３を組み合わせて車両を駆動するハイブリッド自動車のパワーユニットを示しており、トランスミッションケース４内に設置されてＣＶＴ機構５を介して車両の駆動及び制動時の回生発電を行うモータージェネレータ３（以下、モーターＡと記す）と、エンジン側方に取り付けられてクランクシャフト６前端のクランクプーリー７から補機駆動ベルト８を介して所定の増速比（回転速度比）αで駆動され、エンジンの始動及び発電を行うモータージェネレータ２（以下、モーターＢと記す）の２つのモータージェネレータを持っている。車両停止時にはエンジン１も停止しており、発進するときには、エンジン１とトランスミッション軸間に設置されたクラッチ９が切れた状態で、まずモーターＡの駆動力により発進し、その後モーターＢによりエンジン１を始動してからクラッチ９をつなぎ、エンジン１とモーターＡ両方の駆動力で車両を加速する。バッテリー残量が不足した場合には、停止時であってもエンジン１を運転して、モーターＢにより発電する。
ここで、燃費の良い運転をしようとすると、エンジン１の運転状態としては低回転高負荷運転を多用することが望ましいが、このような運転条件ではエンジン１の燃焼に伴うトルク変動が大きく、それによるエンジン１のロール振動や回転速度変動により車室内の振動や騒音が悪化する。
【０００９】
本実施例のようにモーターＢを補機駆動ベルト８で駆動する場合、ベルト８が比較的柔らかいためそれがばねとなり、そこにつながるクランクシャフト６と一体回転するフライホイール１０と、モーターＢのローター部１１がマスとなる振動系が構成される。この共振周波数においてはエンジン１のロール振動や回転速度変動が悪化するが、それより少し低い周波数にはそれらのレベルが小さくなる反共振周波数が存在し、この反共振周波数とエンジン１の回転基本次数（回転周波数×気筒数／２）が一致する回転数では振動が低減される。
また、モーターＢの回転角変位θ_mから、クランクシャフトの回転角変位θ_cにモーターＢの増速比αをかけた回転角変位を引いた差に定数−Ｋ_mをかけたトルク（トルクは正の値又は負の値をとります）を、モーターＢから発生させることにより、このトルクはモーターＢとクランクシャフト６間に設置したばね定数Ｋのばねのばね力と等価になり、この電気的な振動系とそれ以外の機械的な振動系とが１つの振動系として作用する。それによりベルト３のばねと、モーターＢからの電気的なばねが並列に接続されていることになるため、電気的なばね定数Ｋ_mを変化させることにより、図３のように反共振周波数を変化させることが出来る。
【００１０】
そこでクランクシャフト６及びモーターＢにはそれらの回転角変位θ_c、θ_mを検出するセンサー１２，１３を設置し、検出された回転角変位に基づいて制御ユニット１４でモーターＢから発生するトルクを算出するのだが、そのトルクT_aはモーターＢの増速比をαとすると、次式（１）で表される。
【００１１】
【数４】

【００１２】
となる。ここでK_mはエンジンの回転数により次式（２）のように変更される。
【００１３】
【数５】

【００１４】
ここで、Ｋ：ベルトのばねによる回転ばね定数、Ｎ_e：エンジンの回転数、Ｎ₀：機械的な振動系のみで回転速度変動が反共振となるエンジンの回転数、である。こうすることにより、常に回転基本次数と回転速度変動の反共振周波数が一致するようになり、大きな回転速度変動低減効果を得ることが出来る。
また、Ｎ₀として機械的な振動系のみでロール振動が反共振となるエンジンの回転数を置いた場合には、大きなロール振動低減効果を得ることが出来る。
次に本発明の第２実施例について説明する。この第２実施例は上述した第１実施例に対して、モーターＢから発生するトルクＴ_aを、次式（３）により決定している。
【００１５】
【数６】

【００１６】
こうすることにより、モーターＢからばね力に加えて機械的な減衰力を打ち消す負の減衰力を発生させていることになり、電気的な減衰力と機械的な減衰力をあわせた振動系全体の減衰力をほぼ０にすることが出来、図４に示すように、第１実施例に比べて反共振の効果を更に大きくすることが出来る。ここで、K_mは第１実施例と同様に求めることができるが、この第２実施例におけるK_mを求める次式（４）においては、N₀がモーターＢの駆動トルクによる電気的なばね力は作用しないが、電気的な減衰力は作用している状態で反共振となる回転数となっている。
【００１７】
【数７】

【００１８】
次に本発明の第３実施例について説明する。
本実施例の自動車用エンジンに２１は、図５及び図６に示すように補機駆動ベルト２２で駆動され、クランクシャフト２３と逆方向に回転し、発電を行うオルタネーター２４が取り付けられている。クランクプーリ２５は、図７及び図８に示すように、クランクシャフト２３に直接結合される内周部２６と、ベルト２２が掛かり、クランクシャフト２３にベアリング２８を介して回転自由に支持される外周部２７に二分割されており、これらは弾性体であるコイルスプリング２９を組み合わせた捩りばね機構３０を介して結合されている。ここで、クランクシャフト２３及びそこに結合されるフライホイール３１のエンジン回転部の慣性質量と、オルタネータ２４ロータ部の慣性質量を、プーリ部のばねで結合した振動系によるエンジンロール振動との反共振の周波数が、アイドル運転時の回転基本次数と一致するようにこの捩りばね機構３０のばね定数が設定されており、そうすることによりアイドル時のエンジンロール振動を低減している。
ここで、オルタネーター２４の発電量ｗはコントローラー３２からの信号により制御することが可能であり、図９ａのように所定の時間間隔をおいて異なる発電量ｗ₁，ｗ₂を繰り返すことにより、オルタネータ２４の駆動トルクＴａを図９ｂに示すように、所定の時間間隔をおいてT₁、T₂を振動的に繰り返す矩形波として制御することができる。この矩形波で与えられる駆動トルクＴａは、次に示すトルクT_a’とその回転基本次数成分が略同一となるように、その振幅、時間間隔、及び位相が決定される。T_a’は機械的な振動系の反共振周波数前後の回転数（後述する図１０の領域▲１▼を参照）では次式（５）で表される。
【００１９】
【数８】

【００２０】
ここでI_mはエンジンの回転数により次式（６）のように変更される。
【００２１】
【数９】

【００２２】
また、T_a’は機械的な振動系の共振周波数前後の回転数(後述する図１０の領域▲２▼を参照)では次式（７）で表される。
【００２３】
【数１０】

【００２４】
正弦曲線で表されるT_a′は共振及び反共振前後の回転数では、回転基本次数成分が大部分を占めるため、T_a′の極大値、極小値間の差Tsを回転基本次数成分の両振幅とみなすことが出来、オルタネーター２４の駆動トルクT_aはその回転基本次数成分がT_sとなるように、つまりT_aのような矩形波の場合には、
T₂-T₁＝0.785T_s
（0.785は矩形波を正弦波に直す場合の一般的な定数です）
となる。
こうすることにより図１０に示すように、反共振周波数前後では、回転数によって慣性力を変更することにより回転基本次数と反共振周波数とが一致するように反共振周波数を変更し、効果領域（図１０における領域▲１▼）を広くすることが出来る。共振周波数前後の回転数（図１０における領域▲２▼）では、電気的に減衰力を付加することにより減衰を大きくし、共振の悪化を防止することが出来る。
【００２５】
また、オルタネーター２４は常に負のトルクしか発生させることが出来ないため、その振幅T₂-T₁を大きく取ろうとすると発電量が過大となってしまい、あまり大きな振幅を発生させることが出来ない。しかし本実施例の場合は、反共振周波数前後で機械的な振動系に対して付加的に発生させるか、振幅の小さな減衰力相当のトルクを発生させるだけなので、オルタネーター２４の駆動トルクは小さくて済み、アクチュエータとしてモータジェネレータを用い、このモータジェネレータの発電量が図１１ａに示すような正弦波となるように、モータジェネレータの駆動トルクを正弦波で与える場合（図１１ｂを参照）に比べ、安価なオルタネーターで制御することが可能である。
上記各実施例から把握し得る本発明の技術的思想について、その効果とともに列記する。
（ａ）フライホイールと結合されたクランクシャフトの回転駆動力を伝える駆動力伝達機構と、前記駆動力伝達機構により回転させられて慣性力を生じる副慣性質量体とを備え、かつ前記駆動力伝達機構に弾性体を持たせて振動系を形成し、この振動系の反共振を用いた内燃機関の振動低減装置において、前記副慣性質量体もしくは前記副慣性質量体の一部が駆動力を発生もしくは吸収するアクチュエータとなっており、かつそのアクチュエータの駆動トルクを前記内燃機関の運転状態に応じて制御することにより前記反共振の周波数を変更または前記反共振の減衰の大きさを変更するアクチュエータ駆動トルク制御手段を備えていることにより、振動系の反共振による効果と共に、前記アクチュエータにより振動を打ち消すトルクを発生することにより、より大きな振動低減効果を得ることが出来る。
（ｂ）前記（ａ）に記載の構成において、前記クランクシャフトと前記副慣性質量体の各回転角変位、前記クランクシャフトと前記副慣性質量体の各回転角速度、前記クランクシャフトと前記副慣性質量体の各回転角変位及び各回転角速度、または前記副慣性質量体の回転角加速度、のいずれかを検出もしくは推定する手段を有し、前記アクチュエータの駆動トルクは、前記クランクシャフトと前記副慣性質量体の各回転角変位、前記クランクシャフトと前記副慣性質量体の各回転角速度、前記クランクシャフトと前記副慣性質量体の各回転角変位及び各回転角速度、または前記副慣性質量体の回転角加速度、のいずれかを用いて制御されていることにより、前記アクチュエータからばね力、減衰力、慣性力に相当するトルクを発生することが出来、その振動系を制御することにより大きな振動低減効果を得ることが出来る。
（ｃ）前記（ｂ）に記載の構成において、前記副慣性質量体は、前記クランクシャフトに対して所定の回転速度比αを持って回転しており、前記アクチュエータの駆動トルクは、前記クランクシャフトの回転角変位に前記回転速度比αを乗じて前記副慣性質量体の回転角変位から引いたものに所定の係数Ｋ_mを乗じた値、前記クランクシャフトの回転角速度に前記回転速度比αを乗じて前記副慣性質量体の回転角速度から引いたものに所定の係数Ｃ_mを乗じた値、前記副慣性質量体の回転角加速度に所定の係数Ｉ_mを乗じた値、もしくは前記クランクシャフトの回転角変位に前記回転速度比αを乗じて前記副慣性質量体の回転角変位から引いたものに所定の係数Ｋ_mを乗じた値と前記クランクシャフトの回転角速度に前記回転速度比αを乗じて前記副慣性質量体の回転角速度から引いたものに所定の係数Ｃ_mを乗じた値との和、のいずれかを用いて制御されていることにより、前記アクチュエータから発生するトルクによって生じる電気的な振動系により、主フライホイールを結合したクランクシャフトと、副慣性質量体との振動系を変更することができ、それを制御することにより大きな振動低減効果が得られる。
（ｄ）前記（ｃ）に記載の構成において、前記内燃機関の回転速度を検出する手段を持ち、前記アクチュータの駆動トルクによる電気的な振動系を含めた前記振動系と、前記内燃機関のロール振動もしくは前記クランクシャフトの回転速度変動との反共振の周波数が、前記内燃機関の所定回転速度における回転周波数の（自然数／２）倍した周波数のうちのいずれかの周波数と略一致するように、前記所定の係数Ｋ_m及びＩ_mを前記内燃機関の回転速度に応じて変更する。すなわち、電気的な振動系を付加し、内燃機関の回転次数と振動系の反共振周波数が一致するようにその振動系を制御することにより、大きな振動低減効果を得ることが出来る。
（ｅ）前記（ｄ）に記載の構成において、前記アクチュエータの駆動トルクは、前記副慣性質量体の回転角変位から前記クランクシャフトの回転角変位に前記回転速度比αをかけたものを引いた差に前記所定の係数Ｋ_mをかけたトルクとなるように制御されており、かつ前記所定の係数Ｋ_mは次式によって算出される。
【００２６】
【数１１】

【００２７】
ここで、K:前記副慣性質量体の回転に対する前記弾性体の回転ばね定数、N_e:前記内燃機関の回転速度、N₀:前記副慣性質量体の回転角変位から前記クランクシャフトの回転角変位に前記回転速度比αをかけたものを引いた差に所定の係数Ｋ_mをかけた前記トルクを前記アクチュエータから付加しないときに反共振となる前記内燃機関の回転速度、とする。
【００２８】
これによって、アクチュエーターから発生するばね力に相当するトルクを制御することにより、前記（ｄ）に記載の効果を実現できる。
（ｆ）前記（ｄ）に記載の構成において、前記アクチュエータの駆動トルクは、前記副慣性質量体の回転角変位から前記クランクシャフトの回転角変位に前記回転速度比αをかけたものを引いた差に前記所定の係数Ｋ_mをかけた値に、前記副慣性質量体の回転角速度から前記クランクシャフトの回転角速度に前記回転速度比αをかけたものを引いた差に正の係数Ｃ_mをかけた値を加えたトルクとなるように制御されており、かつ前記所定の係数Ｋ_mは次式によって算出される。
【００２９】
【数１２】

【００３０】
ここで、K:前記副慣性質量体の回転に対する前記弾性体の回転ばね定数、N_e:前記内燃機関の回転速度、N₀:前記副慣性質量体の回転角変位から前記クランクシャフトの回転角変位に前記回転速度比αをかけたものを引いた差に所定の係数Ｋ_mをかけた前記トルクを前記アクチュエータから付加しないが、電気的な減衰力は作用している状態で反共振となる前記内燃機関の回転速度、とする。これによって、アクチュエータから負の減衰力に相当するトルクを発生させることにより、振動系の減衰を小さくし、反共振の効果を大きくすることが出来る。
（ｇ）前記（ｆ）に記載の構成において、前記副慣性質量体の回転角速度から前記クランクシャフトの回転角速度に前記回転速度比αをかけたものを引いた差にかける正の係数Ｃ_mは、前記アクチュエータから駆動トルクを付加しないときの、前記副慣性質量体の回転に対する回転減衰定数と略一致している。すなわち、アクチュエータから機械的な減衰力と略一致する負の減衰力に相当するトルクを発生させることにより、振動系の減衰をほぼ０にし、反共振の効果をさらに大きくすることが出来る。
（ｈ）前記（ｃ）〜（ｇ）のいずれかに記載の構成において、前記アクチュエータの駆動トルクを制御しないときの、前記内燃機関のロール振動もしくは前記クランクシャフトの回転速度変動と、前記振動系との反共振周波数近傍の回転速度にて、前記アクチュエータの駆動トルクを制御する。すなわち、機械的な振動系の反共振周波数近傍でのみ制御を行うことにより、アクチュエーターの制御力が小さくても必要な効果を得ることが出来る。
（ｉ）前記（ｈ）に記載の構成において、前記アクチュエータの駆動トルクは、副慣性質量体の回転角加速度に前記所定の係数Ｉ_mをかけた値を用いて制御されており、かつ前記所定の係数Ｉ_mは次式によって算出される。
【００３１】
【数１３】

【００３２】
ここで、I:前記副慣性質量体の慣性モーメント、N_e:前記内燃機関の回転速度、N₀:前記副慣性質量体の回転角加速度に前記所定の係数Ｉ_mをかけた前記トルクを前記アクチュエータから付加しないときに反共振となる前記内燃機関の回転速度、とする。これによって、アクチュエータから発生する慣性力に相当するトルクを制御することにより、前記（ｄ）に記載の効果を実現できる。
（ｊ）前記（ｃ）または（ｈ）または（ｉ）のいずれかに記載の構成において、前記アクチュエータの駆動トルクを制御しないときの、前記内燃機関のロール振動もしくは前記クランクシャフトの回転速度変動と、前記振動系との共振周波数近傍の回転速度において、前記アクチュエータの駆動トルクは、前記副慣性質量体の回転角速度から前記クランクシャフトの回転角速度に前記回転速度比αをかけたものを引いた差に負の係数をかけたトルクとなるように制御されていることによって、共振周波数近傍で、アクチュエータから正の減衰力に相当するトルクを発生させることにより、振動系の減衰を大きくし、共振の悪化を小さくすることが出来る。
（ｋ）前記（ａ）〜（ｊ）のいすれかに記載の構成において、前記アクチュエータの駆動トルクは、その主たる周波数成分において、振幅、位相及び周期を制御する。すなわち、前記アクチュエータから、例えば矩形波のような少なくともその主たる周波数成分に関して制御できる制御トルクを発生することにより、制御を容易にすることが出来る。
（ｌ）前記（ａ）〜（ｋ）のいずれかに記載の構成において、前記アクチュエータの駆動トルクを制御しないときの、前記内燃機関のロール振動もしくは前記クランクシャフトの回転速度変動と、前記振動系との反共振周波数が、前記内燃機関の運転状態のうち、略一定回転速度で運転される運転状態での前記内燃機関の回転速度おける回転周波数の（自然数／２）倍した周波数にうちのいずれかの周波数と略一致している。すなわち、機械的な振動系の反共振周波数が、内燃機関の略一定回転で運転される運転条件での回転次数と一致していることにより、この運転条件では機械的な振動系の反共振による効果が得られるため、アクチュエータによる制御力を発生させる必要が無くなる。
（ｍ）上記（ｌ）に記載の構成において、前記アクチュエータの駆動トルクを制御しないときの、内燃機関のロール振動もしくは前記クランクシャフトの回転速度変動と、前記振動系との反共振周波数が、前記内燃機関の運転状態のうち、略一定回転速度で運転される運転状態での前記内燃機関の回転速度における回転周波数の（内燃機関の気筒数／２）倍した周波数と略一致している。すわわち、機械的な振動系の反共振周波数が、内燃機関の略一定回転で運転される運転条件での回転次数のうち、通常最もレベルが大きい回転基本次数（気筒数／２次）と一致していることにより、この運転条件では機械的な振動系の反共振により大きな効果が得られる。
（ｎ）前記（ｌ）または（ｍ）に記載の構成において、前記略一定回転速度で運転される運転状態が、アイドル運転であることによって、内燃機関のロール振動や回転速度変動による問題が発生しやすいアイドル運転に機械的な振動系の反共振周波数を設定することにより、大きな振動低減効果を得ることが出来る。
（ｏ）前記（ａ）〜（ｎ）のいずれかに記載の構成において、前記アクチュエータが内燃機関に取り付けられたオルタネーターもしくはモータージェネレーターである。これによって、発電を行うオルタネーターもしくは発電と共に内燃機関の始動や車両の駆動等も行うモータージェネレーターとアクチュエーターを兼ねることにより、少ない部品点数で本発明を実現できる。
【図面の簡単な説明】
【図１】本発明に係る内燃機関の振動低減装置を備えたパワーユニットの正面図。
【図２】本発明に係る内燃機関の振動低減装置を備えたパワーユニットの側面図。
【図３】本発明の第１実施例における効果を示す説明図。
【図４】本発明の第２実施例における効果を示す説明図。
【図５】本発明に係る内燃機関の振動低減装置を備えた本発明の第３実施例における自動用エンジンの正面図。
【図６】本発明に係る内燃機関の振動低減装置を備えた本発明の第３実施例における自動用エンジンの側面図。
【図７】図５に示すクランクプーリの正面図。
【図８】図５に示すクランクプーリの断面図。
【図９】本発明の第３実施例におけるオルタネータ発電量と駆動トルクとの関係を示す説明図。
【図１０】本発明の第３実施例における効果を示す説明図。
【図１１】モータジェネレータの発電量とモータジェネレータの駆動トルクとの関係を示す説明図。
【符号の説明】
１…エンジン
２…モータジェネレータ（モーターＢ）
３…モータジェネレータ（モーターＡ）
６…クランクシャフト
７…クランクプーリ
８…補機駆動ベルト
１０…フライホイール[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a vibration reduction device for an internal combustion engine.
[0002]
[Prior art]
A driving force transmission mechanism that transmits a rotational driving force of a crankshaft coupled with a flywheel; and a sub-inertial mass body that is rotated by the driving force transmission mechanism to generate an inertial force, and the driving force transmission mechanism includes An elastic body is provided to form a vibration system, and the anti-resonance frequency of the vibration system is set to n / 2 of rotation in an operation state in which the internal combustion engine is operated at a substantially constant rotation (n = A vibration reduction device for an internal combustion engine that reduces vibrations of the internal combustion engine in its operating state by making it substantially coincide with a frequency that is a natural number) times is known (Japanese Patent Laid-Open No. 11-325186).
[0003]
Japanese Patent Application Laid-Open No. 11-044331 discloses a method in which the motor generator is rotated in the direction opposite to the rotation direction of the internal combustion engine, thereby canceling the torque fluctuation of the internal combustion engine by the reaction force torque caused by the inertial force of the rotor portion of the motor generator. Furthermore, a technique is disclosed in which a torque fluctuation that cancels the torque fluctuation of the internal combustion engine is generated from the motor generator to reduce the vibration accompanying the torque fluctuation of the internal combustion engine.
[0004]
[Problems to be solved by the invention]
However, in the former conventional example (conventional technology described in Japanese Patent Application Laid-Open No. 11-325186), a large vibration reduction effect is obtained in the vicinity of the anti-resonance frequency, but if the frequency is deviated, the effect is reduced. In other words, even if the rotational speed is slightly deviated from the operating conditions set so that the anti-resonance frequency and the frequency n / 2 (n = natural number) times the rotational speed substantially coincide with each other, the reduction effect is reduced.
For example, when the anti-resonance frequency is set so that the anti-resonance effect can be obtained in the idle operation, the idle operation depends on the load condition of auxiliary equipment such as an air conditioner and an alternator, the A / T N range, the D range selected state, etc. However, because the operating conditions are different, it may be better to change the rotation speed, but in the former conventional example, the idle rotation speed is always kept the same to obtain the maximum vibration reduction effect. There must be. Further, since resonance occurs at a frequency slightly higher than the anti-resonance frequency, there is a problem in that vibration is accompanied by deterioration.
In the latter conventional example (Japanese Patent Application Laid-Open No. 11-044331), when a belt is used to drive the motor generator, the belt is relatively soft so that it becomes a spring and rotates integrally with a crankshaft connected therewith. Thus, a vibration system in which the rotor portion of the motor generator becomes a mass is configured. And since the phases of the flywheel and the motor generator do not match before and after the resonance frequency of this vibration system and higher, the torque fluctuation of the internal combustion engine can be reduced simply by generating a reverse phase torque from the motor generator. Cannot be done. In other words, the torque fluctuation of the internal combustion engine can be reduced at a relatively low rotation speed, but the torque fluctuation is reduced at a rotation speed before and after the rotation speed at which the basic rotation order (rotation frequency x number of cylinders / 2) becomes the resonance frequency. There is a problem that it cannot be reduced.
[0005]
[Means for Solving the Problems]
Therefore, the vibration reducing device of the present invention includes a crankshaft that transmits the driving force of the internal combustion engine as a rotational driving force, a flywheel that rotates integrally with the crankshaft, and a sub inertia mass that rotates as the crankshaft rotates. A body, a driving force transmission mechanism for transmitting the rotational driving force of the crankshaft to the sub inertial mass body via an elastic body to rotationally drive the subinertial mass body, and means for detecting the rotational speed of the internal combustion engine; The frequency of the antiresonance between the vibration system generated by the spring action of the driving force transmission mechanism and the roll vibration of the internal combustion engine is a frequency obtained by multiplying the rotation frequency at the predetermined rotation speed of the internal combustion engine by (natural number / 2). in the vibration reducing device for an internal combustion engine so as to substantially coincide, the rotational driving force of the crankshaft to the secondary inertial mass is transmitted by the driving force transmission mechanism Used has a power generation capability actuator, and at least one of the attenuation of the frequency and the anti-resonance of the anti-resonance size by controlling in accordance with the driving torque of the actuator to the operating state of the internal combustion engine A changeable actuator drive torque control means is provided.
[0006]
【The invention's effect】
According to the present invention, it is possible to obtain a greater vibration reduction effect by generating torque that cancels vibration by the actuator, in addition to the effect of anti-resonance of the vibration system.
[0007]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, an embodiment of the present invention will be described in detail with reference to the drawings.
[0008]
First, that describes a first embodiment of the present invention with reference to FIGS. 1 to 3.
1 and 2 show a hybrid vehicle power unit that drives a vehicle by combining an engine 1 and

motor generators

2 and 3. The vehicle is driven and braked via a CVT mechanism 5 installed in a transmission case 4. A motor generator 3 that performs regenerative power generation (hereinafter referred to as motor A), and a predetermined speed increasing ratio (rotation) from the crank pulley 7 at the front end of the crankshaft 6 via the auxiliary drive belt 8 attached to the engine side. The motor generator 2 (hereinafter referred to as “motor B”), which is driven at a speed ratio (α) and starts the engine and generates power, has two motor generators. When the vehicle is stopped, the engine 1 is also stopped. When starting the vehicle, the clutch 9 installed between the engine 1 and the transmission shaft is disengaged, the vehicle 1 is first started by the driving force of the motor A, and then the engine 1 is driven by the motor B. After starting, the clutch 9 is engaged, and the vehicle is accelerated by the driving force of both the engine 1 and the motor A. When the remaining battery level is insufficient, the engine 1 is operated even when stopped, and the motor B generates power.
Here, in order to perform an operation with good fuel consumption, it is desirable to frequently use a low-rotation and high-load operation as the operation state of the engine 1, but under such an operation condition, the torque fluctuation accompanying the combustion of the engine 1 is large, Due to the roll vibration and rotational speed fluctuation of the engine 1 caused by the above, the vibration and noise in the passenger compartment deteriorate.
[0009]
When the motor B is driven by the accessory drive belt 8 as in the present embodiment, the belt 8 is relatively soft and thus becomes a spring, and the flywheel 10 that rotates integrally with the crankshaft 6 connected thereto, and the rotor of the motor B A vibration system in which the portion 11 becomes a mass is configured. At this resonance frequency, the roll vibration and rotational speed fluctuation of the engine 1 are deteriorated. However, there is an anti-resonance frequency at which the level is slightly lower than this, and this anti-resonance frequency and the basic rotation order of the engine 1 are present. Vibrations are reduced at a rotational speed where (rotational frequency x number of cylinders / 2) matches.
Further, the torque (torque is the torque is the difference between the rotational angular displacement θ _m of the motor B and the rotational angular displacement obtained by multiplying the rotational angular displacement θ _c of the crankshaft by the acceleration ratio α of the motor B and the constant −K _m. (Takes a positive value or a negative value) from the motor B, this torque becomes equivalent to the spring force of the spring of the spring constant K installed between the motor B and the crankshaft 6, and this electrical A simple vibration system and other mechanical vibration systems act as one vibration system. Whereby the spring of the belt 3, for electrical spring from the motor B is ing that are connected in parallel, by changing the electrical spring constant K _m, the anti-resonance frequency as shown in FIG. 3 Can be changed.
[0010]
Therefore, the crankshaft 6 and the motor B are provided with

sensors

12 and 13 for detecting their rotational angular displacements θ _c and θ _m, and the torque generated from the motor B by the control unit 14 based on the detected rotational angular displacements. I calculated, but the torque T _a is the speed increasing ratio of the motor B and alpha, is expressed by the following equation (1).
[0011]
[Expression 4]

[0012]
It becomes. Here, K _m is changed as shown in the following equation (2) _depending on the engine speed.
[0013]
[Equation 5]

[0014]
Here, K is a rotational spring constant by the spring of the belt, N _{e is the} engine speed, and N ₀ is the engine speed at which the rotational speed fluctuation is anti-resonant only by the mechanical vibration system. By doing this, the rotation basic order always matches the anti-resonance frequency of the rotational speed fluctuation, and a large rotational speed fluctuation reducing effect can be obtained.
In addition, when N ₀ is set to an engine speed at which roll vibration becomes anti-resonance only with a mechanical vibration system, a large roll vibration reduction effect can be obtained.
Next, a second embodiment of the present invention will be described. For this second embodiment the first embodiment described above, the torque T _a which is generated from the motor B, and determined by the following equation (3).
[0015]
[Formula 6]

[0016]
In this way, the motor B generates a negative damping force that cancels the mechanical damping force in addition to the spring force, and the entire vibration system that combines the electrical damping force and the mechanical damping force. As shown in FIG. 4, the anti-resonance effect can be further increased as compared with the first embodiment. Here, K _m can be obtained in the same manner as in the first embodiment, but in the following equation (4) for obtaining K _m in the second embodiment, N ₀ is an electric spring based on the driving torque of the motor B. The force does not act, but the electric damping force is the rotational speed that causes anti-resonance in the acted state.
[0017]
[Expression 7]

[0018]
Next, a third embodiment of the present invention will be described.
As shown in FIGS. 5 and 6, the automobile engine 21 of this embodiment is driven by an auxiliary machine drive belt 22, and is attached with an alternator 24 that rotates in the opposite direction to the crankshaft 23 and generates electric power. As shown in FIGS. 7 and 8, the crank pulley 25 has an inner peripheral portion 26 that is directly coupled to the crankshaft 23 and a belt 22, and an outer periphery that is rotatably supported by the crankshaft 23 via a bearing 28. The portion 27 is divided into two parts, and these are coupled via a torsion spring mechanism 30 that combines a coil spring 29 that is an elastic body. Here, the anti-resonance between the engine roll vibration by the vibration system in which the inertia mass of the engine rotating portion of the crankshaft 23 and the flywheel 31 coupled thereto and the inertia mass of the alternator 24 rotor portion are coupled by the spring of the pulley portion. The spring constant of the torsion spring mechanism 30 is set so that the frequency of the torque coincides with the rotation basic order during idling, thereby reducing engine roll vibration during idling.
Here, the power generation amount w of the alternator 24 can be controlled by a signal from the controller 32. By repeating different power generation amounts w ₁ and w ₂ at predetermined time intervals as shown in FIG. As shown in FIG. 9B, the drive torque Ta of 24 can be controlled as a rectangular wave that repeats T ₁ and T ₂ oscillating at a predetermined time interval. The amplitude, time interval, and phase of the driving torque Ta given by the rectangular wave are determined so that the following torque T _a ′ and the rotation basic order component thereof are substantially the same. T _a ′ is represented by the following equation (5) at the rotational speed around the anti-resonance frequency of the mechanical vibration system (see region (1) in FIG. 10 described later).
[0019]
[Equation 8]

[0020]
Here, _Im is changed as shown in the following equation (6) _depending on the engine speed.
[0021]
[Equation 9]

[0022]
Further, T _a ′ is represented by the following expression (7) in the rotational speeds before and after the resonance frequency of the mechanical vibration system (refer to region (2) in FIG. 10 described later).
[0023]
[Expression 10]

[0024]
Since T _a ′ represented by a sine curve is predominantly the rotation basic order component at the rotation speeds before and after resonance and anti-resonance, the difference Ts between the maximum and minimum values of T _a ′ is expressed as the rotation basic order component. It can be regarded as both the amplitude, so that the driving torque T _a of the alternator 24 is rotation basic order component and T _s, in the case of a rectangular wave, such as that is T _a is
T ₂ -T ₁ = 0.785T _s
(0.785 is a general constant for converting a square wave into a sine wave.)
It becomes.
As a result, as shown in FIG. 10, before and after the anti-resonance frequency, the anti-resonance frequency is changed so that the rotation basic order and the anti-resonance frequency coincide with each other by changing the inertial force according to the rotation speed. The area (1) in FIG. 10 can be widened. At the rotational speeds around the resonance frequency (region (2) in FIG. 10), the damping can be increased by electrically adding a damping force to prevent the deterioration of the resonance.
[0025]
In addition, since the alternator 24 can always generate only a negative torque, if the amplitude T ₂ -T ₁ is increased, the amount of power generation becomes excessive, and a very large amplitude cannot be generated. However, in the case of the present embodiment, the drive torque of the alternator 24 is small because it is generated in addition to the mechanical vibration system around the anti-resonance frequency or only a torque corresponding to a damping force with a small amplitude is generated. Compared with the case where a motor generator is used as an actuator and the driving torque of the motor generator is given as a sine wave so that the power generation amount of the motor generator becomes a sine wave as shown in FIG. 11a (see FIG. 11b). It is possible to control with a simple alternator.
The technical ideas of the present invention that can be grasped from the above embodiments will be listed together with their effects.
(A) a driving force transmission mechanism that transmits a rotational driving force of a crankshaft coupled to a flywheel; and a sub-inertial mass body that is rotated by the driving force transmission mechanism to generate an inertial force; In a vibration reduction device for an internal combustion engine that uses an anti-resonance of the vibration system by forming an elastic body in the mechanism, the sub inertia mass body or a part of the sub inertia mass body generates a driving force. Alternatively, an actuator that is an absorbing actuator and that changes the anti-resonance frequency or changes the anti-resonance attenuation level by controlling the driving torque of the actuator according to the operating state of the internal combustion engine. By providing the torque control means, the actuator generates a torque that cancels the vibration, together with the effect of anti-resonance of the vibration system. Therefore, a greater vibration reduction effect can be obtained.
(B) In the configuration described in (a), each rotational angular displacement of the crankshaft and the sub inertial mass body, each rotational angular velocity of the crankshaft and the subinertial mass body, the crankshaft and the subinertial mass Means for detecting or estimating any of the rotational angular displacements and rotational angular velocities of the body, or the rotational angular acceleration of the sub inertial mass body, and the drive torque of the actuator includes the crankshaft and the sub inertial mass. Each rotational angular displacement of the body, each rotational angular velocity of the crankshaft and the sub-inertial mass body, each rotational angular displacement and each rotational angular velocity of the crankshaft and the sub-inertial mass body, or rotational angular acceleration of the sub-inertial mass body , By generating a torque corresponding to a spring force, a damping force, and an inertial force from the actuator. A large vibration reduction effect can be obtained by controlling the vibration system.
(C) In the configuration described in (b), the sub-inertial mass body rotates with a predetermined rotational speed ratio α with respect to the crankshaft, and the drive torque of the actuator is the crankshaft. Is obtained by multiplying the rotational angular displacement of the sub-inertial mass body by the rotational speed ratio α and multiplying the rotational angular displacement of the sub-inertial mass body by a predetermined coefficient K _m, and the rotational speed ratio α of the rotational angular velocity of the crankshaft. multiplied by the value obtained by multiplying a predetermined coefficient C _m minus the angular velocity of the secondary inertial mass, the value obtained by multiplying a predetermined coefficient I _m in the rotational angular acceleration of the auxiliary inertial mass or the crankshaft, A value obtained by multiplying the rotational angular displacement by the rotational speed ratio α and subtracting the rotational angular displacement of the sub-inertial mass body by a predetermined coefficient K _m and the rotational angular velocity of the crankshaft are multiplied by the rotational speed ratio α. Said It is controlled by using the sum of the value obtained by multiplying a predetermined coefficient C _m minus the angular velocity of the secondary inertial mass, one of the electrical vibration caused by the torque generated from the actuator The vibration system of the crankshaft coupled with the main flywheel and the sub inertia mass body can be changed by the system, and a large vibration reduction effect can be obtained by controlling the vibration system.
(D) In the configuration described in (c), the vibration system including means for detecting the rotational speed of the internal combustion engine, including an electric vibration system based on the driving torque of the actuator, and the roll of the internal combustion engine The frequency of anti-resonance with vibration or the rotational speed fluctuation of the crankshaft substantially coincides with any one of the frequencies (natural number / 2) times the rotational frequency at the predetermined rotational speed of the internal combustion engine. The predetermined coefficients K _m and I _m are changed according to the rotational speed of the internal combustion engine. That is, by adding an electrical vibration system and controlling the vibration system so that the rotational order of the internal combustion engine and the anti-resonance frequency of the vibration system coincide with each other, a large vibration reduction effect can be obtained.
(E) In the configuration described in (d), the driving torque of the actuator is obtained by subtracting the rotational angular displacement of the crankshaft multiplied by the rotational speed ratio α from the rotational angular displacement of the sub inertial mass body. said predetermined is controlled such that the torque multiplied by the coefficient K _m in, and the predetermined coefficient K _m is calculated from the following equation.
[0026]
## EQU11 ##

[0027]
Where, K: rotation spring constant of the elastic body with respect to rotation of the sub inertial mass body, N _e : rotation speed of the internal combustion engine, N ₀ : rotation angle displacement of the sub inertial mass body from the rotation angle displacement of the crankshaft The rotational speed of the internal combustion engine that is anti-resonant when the torque obtained by multiplying the displacement multiplied by the rotational speed ratio α and the predetermined coefficient _Km is not applied from the actuator.
[0028]
Thus, the effect described in (d) can be realized by controlling the torque corresponding to the spring force generated from the actuator.
(F) In the configuration described in (d), the driving torque of the actuator is obtained by subtracting the rotational angular displacement of the crankshaft multiplied by the rotational speed ratio α from the rotational angular displacement of the sub inertial mass body. the value obtained by multiplying the predetermined coefficient K _m on the difference, the secondary inertial mass positive coefficient C _m the difference obtained by subtracting the rotational angular velocities multiplied by the said rotational speed ratio α of the crankshaft from the rotational angular velocity of the The torque is controlled by adding the multiplied value, and the predetermined coefficient _Km is calculated by the following equation.
[0029]
[Expression 12]

[0030]
Where, K: rotation spring constant of the elastic body with respect to rotation of the sub inertial mass body, N _e : rotation speed of the internal combustion engine, N ₀ : rotation angle displacement of the sub inertial mass body from the rotation angle displacement of the crankshaft The torque obtained by multiplying the displacement multiplied by the rotational speed ratio α and the predetermined coefficient K _m is not applied from the actuator, but the resonance is anti-resonant in the state where the electric damping force is applied. The rotational speed of the internal combustion engine. Thus, by generating a torque corresponding to a negative damping force from the actuator, it is possible to reduce the damping of the vibration system and increase the anti-resonance effect.
(G) In the configuration described in (f), the positive coefficient C _m applied to the difference obtained by subtracting the rotational angular velocity of the crankshaft multiplied by the rotational angular velocity α from the rotational angular velocity of the sub inertial mass body is When the driving torque is not applied from the actuator, the rotational damping constant substantially coincides with the rotation of the sub inertial mass body. That is, by generating a torque corresponding to a negative damping force that substantially matches the mechanical damping force from the actuator, the damping of the vibration system can be made substantially zero, and the anti-resonance effect can be further increased.
(H) In the configuration according to any one of (c) to (g), when the driving torque of the actuator is not controlled, roll vibration of the internal combustion engine or rotational speed fluctuation of the crankshaft, and the vibration system The drive torque of the actuator is controlled at a rotational speed near the antiresonance frequency. That is, by performing control only in the vicinity of the anti-resonance frequency of the mechanical vibration system, the necessary effect can be obtained even if the control force of the actuator is small.
(I) In the configuration described in the (h), the driving torque of the actuator is controlled using a value which the multiplied by a predetermined coefficient I _m in the rotational angular acceleration of the auxiliary inertial mass, and the predetermined the coefficients I _m of is calculated by the following equation.
[0031]
[Formula 13]

[0032]
Here, I: moment of inertia of the secondary inertial mass, N _e: rotational speed of the internal combustion engine, N _0: the said the torque multiplied by the predetermined coefficient I _m in the rotational angular acceleration of the auxiliary inertial mass The rotational speed of the internal combustion engine is anti-resonant when not added from the actuator. Thus, the effect described in (d) can be realized by controlling the torque corresponding to the inertial force generated from the actuator.
(J) In the configuration according to any one of (c), (h), and (i), when the drive torque of the actuator is not controlled, the roll vibration of the internal combustion engine or the rotational speed fluctuation of the crankshaft The driving torque of the actuator is a difference obtained by subtracting the rotational angular speed of the crankshaft multiplied by the rotational speed ratio α from the rotational angular speed of the sub-inertial mass body at a rotational speed near the resonance frequency with the vibration system. By controlling the torque so that the torque is multiplied by a negative coefficient, by generating a torque equivalent to a positive damping force from the actuator near the resonance frequency, the damping of the vibration system is increased, and the resonance Deterioration can be reduced.
(K) In the configuration described in any one of (a) to (j), the driving torque of the actuator controls the amplitude, phase, and period in its main frequency component. That is, control can be facilitated by generating a control torque that can be controlled with respect to at least its main frequency component, such as a rectangular wave, from the actuator.
(L) In the configuration according to any one of (a) to (k), when the driving torque of the actuator is not controlled, roll vibration of the internal combustion engine or rotational speed fluctuation of the crankshaft, and the vibration system The anti-resonance frequency of the internal combustion engine is a frequency obtained by multiplying the rotational frequency of the internal combustion engine at a substantially constant rotational speed (natural number / 2). It almost matches the frequency. That is, the anti-resonance frequency of the mechanical vibration system matches the rotational order under the operating condition in which the internal combustion engine is operated at a substantially constant rotation. Since the effect is obtained, it is not necessary to generate a control force by the actuator.
(M) In the configuration described in (l) above, when the driving torque of the actuator is not controlled, roll vibration of the internal combustion engine or rotational speed fluctuation of the crankshaft and antiresonance frequency with the vibration system are Of the operating states of the internal combustion engine, the frequency substantially coincides with the frequency obtained by multiplying the rotational frequency at the rotational speed of the internal combustion engine in the operating state operated at a substantially constant rotational speed (number of cylinders of the internal combustion engine / 2). In other words, the anti-resonance frequency of the mechanical vibration system is usually the highest rotation basic order (number of cylinders / secondary) among the rotation orders under the operating conditions where the internal combustion engine is operated at a substantially constant rotation. Due to the coincidence, a great effect is obtained by anti-resonance of the mechanical vibration system under this operating condition.
(N) In the configuration described in the above (l) or (m), when the operation state operated at the substantially constant rotation speed is an idle operation, a problem due to roll vibration of the internal combustion engine or a rotation speed fluctuation occurs. By setting the anti-resonance frequency of the mechanical vibration system for easy idling operation, a large vibration reduction effect can be obtained.
(O) In the configuration according to any one of (a) to (n), the actuator is an alternator or a motor generator attached to an internal combustion engine. Thus, the present invention can be realized with a small number of parts by combining an alternator for generating power or a motor generator and an actuator for starting an internal combustion engine and driving a vehicle together with power generation.
[Brief description of the drawings]
FIG. 1 is a front view of a power unit including a vibration reduction device for an internal combustion engine according to the present invention.
FIG. 2 is a side view of a power unit including a vibration reduction device for an internal combustion engine according to the present invention.
FIG. 3 is an explanatory diagram showing effects in the first embodiment of the present invention.
FIG. 4 is an explanatory diagram showing effects in the second embodiment of the present invention.
FIG. 5 is a front view of an automatic engine according to a third embodiment of the present invention equipped with a vibration reducing device for an internal combustion engine according to the present invention.
FIG. 6 is a side view of an automatic engine according to a third embodiment of the present invention equipped with a vibration reducing device for an internal combustion engine according to the present invention.
7 is a front view of the crank pulley shown in FIG. 5. FIG.
8 is a cross-sectional view of the crank pulley shown in FIG.
FIG. 9 is an explanatory diagram showing a relationship between an alternator power generation amount and drive torque in a third embodiment of the present invention.
FIG. 10 is an explanatory diagram showing effects in the third embodiment of the present invention.
FIG. 11 is an explanatory diagram showing a relationship between a power generation amount of a motor generator and a driving torque of the motor generator.
[Explanation of symbols]
1 ... Engine 2 ... Motor generator (Motor B)
3. Motor generator (Motor A)
6 ... Crankshaft 7 ... Crank pulley 8 ... Auxiliary drive belt 10 ... Flywheel

Claims

A crankshaft that transmits the driving force of the internal combustion engine as a rotational driving force, a flywheel that rotates integrally with the crankshaft, a sub-inertial mass that rotates as the crankshaft rotates, and a rotational driving force of the crankshaft. A driving force transmission mechanism that transmits the sub inertial mass body through an elastic body to rotationally drive the subinertial mass body; and a means for detecting the rotational speed of the internal combustion engine, the driving force transmission mechanism comprising: The vibration of the internal combustion engine in which the frequency of anti-resonance between the vibration system caused by the spring action and the roll vibration of the internal combustion engine substantially coincides with a frequency (natural number / 2) times the rotational frequency at the predetermined rotational speed of the internal combustion engine. In the reduction device,
The sub-inertial mass body is an actuator capable of generating electric power using the rotational driving force of the crankshaft transmitted by the driving force transmission mechanism , and the driving torque of the actuator is controlled according to the operating state of the internal combustion engine. An internal combustion engine vibration reduction device comprising: an actuator drive torque control means capable of changing at least one of the anti-resonance frequency and the anti-resonance attenuation.

A crankshaft that transmits the driving force of the internal combustion engine as a rotational driving force, a flywheel that rotates integrally with the crankshaft, a sub-inertial mass that rotates as the crankshaft rotates, and a rotational driving force of the crankshaft. A driving force transmission mechanism that transmits the sub inertial mass body through an elastic body to rotationally drive the subinertial mass body; and a means for detecting the rotational speed of the internal combustion engine, the driving force transmission mechanism comprising: The frequency of the anti-resonance between the vibration system generated by the spring action and the rotational fluctuation of the crankshaft is approximately the same as the frequency (natural number / 2) times the rotational frequency at the predetermined rotational speed of the internal combustion engine. In the vibration reduction device,
The sub-inertial mass body is an actuator capable of generating electric power using the rotational driving force of the crankshaft transmitted by the driving force transmission mechanism , and the driving torque of the actuator is controlled according to the operating state of the internal combustion engine. An internal combustion engine vibration reduction device comprising: an actuator drive torque control means capable of changing at least one of the anti-resonance frequency and the anti-resonance attenuation.

Means for detecting rotational angular displacements of the crankshaft and the sub inertial mass body;
The sub inertia mass body rotates with a predetermined rotational speed ratio α with respect to the crankshaft,
Driving torque of the actuator, a torque said from the rotation angle displacement of the secondary inertial mass multiplied by the predetermined coefficient K _m the the rotational speed ratio minus multiplied by α difference in rotational angular displacement of the crankshaft And the predetermined coefficient K _m is set so that the anti-resonance frequency substantially coincides with a frequency (natural number / 2) times the rotational frequency at the predetermined rotational speed of the internal combustion engine. The vibration reduction device for an internal combustion engine according to claim 1 , wherein the vibration reduction device is calculated by an equation.

Here, K: rotation spring constant of the elastic body with respect to rotation of the sub inertial mass body, N _e : rotation speed of the internal combustion engine, N ₀ : rotation angle of the crankshaft from the rotation angle displacement of the sub inertial mass body The rotational speed of the internal combustion engine that is anti-resonant when the torque obtained by multiplying the displacement multiplied by the rotational speed ratio α and the predetermined coefficient _Km is not applied from the actuator.

Means for detecting each rotational angular displacement and each rotational angular velocity of the crankshaft and the sub inertial mass body;
The sub inertia mass body rotates with a predetermined rotational speed ratio α with respect to the crankshaft ,
Driving torque of the actuator, said the value of the multiplied the predetermined coefficient K _m in the rotational speed ratio minus multiplied by α difference in rotational angular displacement of the crank shaft from the rotation angle displacement of the secondary inertial mass , wherein is controlled to be said to the rotational speed ratio minus multiplied by α difference multiplied by the positive coefficient C _m value plus the torque to the rotational angular velocity of the crank shaft from the rotational angular velocity of the secondary inertial mass And the predetermined coefficient _Km is calculated by the following equation so that the anti-resonance frequency substantially coincides with a frequency (natural number / 2) times the rotational frequency at the predetermined rotational speed of the internal combustion engine. 3. The vibration reduction device for an internal combustion engine according to claim 1 , wherein the vibration reduction device is an internal combustion engine.

Here, K: rotation spring constant of the elastic body with respect to rotation of the sub inertial mass body, N _e : rotation speed of the internal combustion engine, N ₀ : rotation angle of the crankshaft from the rotation angle displacement of the sub inertial mass body Although the torque obtained by multiplying the displacement multiplied by the rotational speed ratio α and the predetermined coefficient K _m is not applied from the actuator, an electrical damping force , that is , a damping force generated in the actuator is not applied. The rotational speed of the internal combustion engine that is anti-resonant in a running state.

The positive coefficient C _m applied to the difference obtained by subtracting the rotational angular velocity of the crankshaft multiplied by the rotational angular velocity α from the rotational angular velocity of the sub-inertial mass body is the value when the driving torque is not applied from the actuator. 5. The vibration reduction device for an internal combustion engine according to claim 4 , wherein the vibration damping device substantially coincides with a rotational damping constant for the rotation of the sub inertial mass body.

When no control driving torque of the actuator, the rotational speed of the frequency near the front Symbol antiresonance, the internal combustion engine according to claim 1, characterized in that for controlling the driving torque of the actuator Vibration reduction device.

Means for detecting rotational angular acceleration of the sub inertial mass body;
Driving torque of the actuator is controlled using a value which the multiplied by a predetermined coefficient I _m in the rotational angular acceleration of the auxiliary inertial mass, and the predetermined coefficient I _m, the frequency of the antiresonance said 7. The vibration reduction device for an internal combustion engine according to claim 6, wherein the vibration reduction device is calculated by the following equation so as to substantially coincide with a frequency multiplied by (natural number / 2) of a rotation frequency at a predetermined rotation speed of the internal combustion engine.

Here, I: moment of inertia of the secondary inertial mass, N _e: rotational speed of the internal combustion engine, N _0: the said the torque multiplied by the predetermined coefficient I _m in the rotational angular acceleration of the auxiliary inertial mass The rotational speed of the internal combustion engine is anti-resonant when not added from the actuator.

When the drive torque of the actuator is not controlled, the drive torque of the actuator is calculated from the rotational angular velocity of the sub inertia mass body at the rotational speed near the resonance frequency of the roll vibration of the internal combustion engine and the vibration system. 8. The internal combustion engine according to claim 6, wherein a torque obtained by multiplying a difference obtained by multiplying the rotational angular velocity of the shaft by the rotational speed ratio α is a torque obtained by multiplying a negative coefficient. 9. Engine vibration reduction device.

When the driving torque of the actuator is not controlled, the driving torque of the actuator is calculated from the rotational angular velocity of the sub inertial mass body at the rotational speed fluctuation of the crankshaft and the rotational speed in the vicinity of the resonance frequency with the vibration system. 8. The control according to claim 6, wherein a torque obtained by multiplying a difference obtained by multiplying a rotational angular ratio of the crankshaft by the rotational speed ratio α is a torque obtained by multiplying a negative coefficient. 9. A vibration reduction device for an internal combustion engine.

The vibration reduction device for an internal combustion engine according to any one of claims 1 to 9 , wherein the drive torque of the actuator controls an amplitude, a phase, and a period in a main frequency component.

When no control driving torque of the actuator, frequency before Symbol antiresonance among the operating state of the internal combustion engine, the rotational speed definitive rotational frequency of the internal combustion engine under operating conditions which is operated at a substantially constant rotational speed vibration reduction apparatus for an internal combustion engine according to any one of claims 1 to 10, characterized in that it (a natural number / 2) multiplied by the frequency substantially coincides.

When no control driving torque of the actuator, frequency before Symbol antiresonance among the operating state of the internal combustion engine, the rotational frequency in the rotational speed of the internal combustion engine under operating conditions which is operated at a substantially constant rotational speed 12. The vibration reduction device for an internal combustion engine according to claim 11 , wherein the vibration reduction device substantially matches the frequency multiplied by (the number of cylinders of the internal combustion engine / 2).

The vibration reduction device for an internal combustion engine according to claim 11 or 12 , wherein the operation state operated at the substantially constant rotation speed is an idle operation.

14. The vibration reduction device for an internal combustion engine according to claim 1, wherein the actuator is an alternator or a motor generator attached to the internal combustion engine.