JPH0223741B2

JPH0223741B2 -

Info

Publication number: JPH0223741B2
Application number: JP55078229A
Authority: JP
Inventors: Kenji Sakano; Masahiro Yamashita; Kenji Yamashita; Masahiro Akeda
Original assignee: Kubota Corp
Current assignee: Kubota Corp
Priority date: 1980-06-09
Filing date: 1980-06-09
Publication date: 1990-05-25
Also published as: JPS576142A

Description

[Detailed description of the invention]

本発明は内燃機関のクランク、ピストン機構の
往復動慣性力に起因する起振力の任意のｎ次成分
を相殺する回転偏心錘式バランサ装置に関するも
のである。第１発明は、内燃機関のクランク軸と
ピストン中心軸とを通る起振方向面から一定距離
隔ててバランサ装置を設け、その多軸回転偏心錘
の遠心力の合力によつて起振力の任意のｎ次成分
を打ち消すと同時に、上記遠心力の合力と起振力
のｎ次成分とで発生する偶力による起振モーメン
トを上記遠心力による全合成モーメントで打ち消
すようなバランサ装置を提供し、かつ、回転偏心
錘の回転運動空間は干渉させながら、回転偏心錘
の初期位相差を利用して回転偏心錘の干渉を防ぐ
ことにより、回転偏心錘の軸心間距離を極力小さ
く設定して、バランサ装置の小型化を図ることを
目的とする。第２発明は第１発明のバランサ装置をユニツト
に組込み、そのユニツトを水冷デイーゼル機関の
外側部に突設された燃料噴射ポンプの背部に設
け、バランサ装置を噴射ポンプのカム軸を介して
クランク軸に連動連結することによつてデイーゼ
ル機関の外側部のデツドスペースを利用してバラ
ンサ装置を装備することを目的とするものであ
る。直列型内燃機関は、シリンダ数やシリンダ径を
種々変えることにより、出力のバリエーシヨンが
得られるので、各種産業用機械のエンジンとして
広範に使用されて来たが、その出力性能以外に振
動・騒音が重要な課題となつて来ている。例え
ば、農業機械等においては、エンジンが防振支持
される構造になつておらず、主要な起振源である
エンジン振動そのものの低減が機械各部の振動低
減のきめ手となつている。こうした問題に対して、多気筒化によつてエン
ジン振動を低減する方法も考えられるが、１気筒
当りの排気量が小さくなつて、燃費が増加するの
で多気筒化にも限界がある。従つて、エンジン振動低減の為には、結局バラ
ンサ装置の装着が必要となるが、既存のバランサ
装置には、次のような問題がある。先ず既存の内燃機関において、ピストン、クラ
ンク機構の往復動慣性力の振動による上下方向起
振力の１次成分を緩和する為に、一般にクランク
アーム外側にクランクピン部分に対するカウンタ
ーウエイトを付加した構造である。しかし、この
カウンターウエイトによつては、起振力の２次以
上の高次成分を打ち消すことが出来ない。そこ
で、従来起振力の例えば２次成分を打ち消すた
め、第１２図に仮想線で示すように、クランク軸
Ｓ直下に１対の回転偏心錘Ｗを設け、この回転偏
心錘Ｗの遠心力の合力を、クランク軸Ｓとピスト
ン中心軸とで規定される起振方向面Ｖ内に発生さ
せる構造のバランサ装置B′が用いられていた。
しかし、この構造の場合、バランサ装置B′をク
ランク軸直下に装備するので、機関の高さがかな
り高くなつてしまうという欠点が避けられない。そこで、仮にバランサ装置を内燃機関の側部に
設けるものとすると、上下方向の起振力は回転偏
心錘の遠心力の合力で打ち消されるものの、起振
力と遠心力とが偶力をなし、起振モーメントが発
生してしまい、機関の振動源となる。もつとも、
クランクアーム部に100％カウンターウエイトを
付加する場合には、起振力が水平方向に振り向け
られるので、起振力の例えば２次成分を打ち消す
ためのバランサ装置は符号B″で示すように機関
のクランク軸Ｓの真横に装着することが出来る
が、この構造の場合機関の横幅が相当増加してし
まうという欠点が避けられない。本発明は、偏置多軸回転偏心錘式バランサ装置
によつて上記の諸欠点を解決し、かつ、バランサ
装置を小型化するものである。即ち、起振方向面
Ｖから所定距離ｄだけ隔てて少なくとも３軸以上
の多軸回転偏心錘式のバランサ装置Ｂを設け、こ
れら回転偏心錘の複数の遠心力の合力で機関の起
振力の所望のｎ次成分Ｆ→_Eに釣合せると同時に、
遠心力の合力と起振力のｎ次成分Ｆ→_Eとから成る
起振偶力による起振モーメントＭ→_Vに、遠心力の
全合成モーメントＭ→_Bを釣合わせ、かつ、バラン
サ装置の各回転偏心錘のうち、隣り合う少なくと
も２個の回転偏心錘の各回転運動空間が相互に干
渉するように、その軸心間距離を短かく設定し、
両回転偏心錘はその初期位相αの差により干渉す
ることなく回転できるように互いに等回転速度で
逆方向に回転するように連動連結することを特徴
とするものである。次に、本発明の実施例を図に基き説明する。第１図、第２図は、水冷立型直列４気筒デイー
ゼル機関Ｅの概略図を示し、クランクケースＣの
上部にシリンダブロツクＤが一体に形成され、こ
のシリンダブロツクＤの上部にシリンダヘツドブ
ロツクＨ・動弁カム室Ｍ・燃料噴射弁Ｎが装備さ
れ、機関の前端部にはギヤケースＧが配設されて
いる。シリンダブロツクＤ内部にはウオークジヤ
ケツトが配設され、その中心部にシリンダが形成
されている。このシリンダのウオータジヤケツト
横外壁１の前寄部から燃料噴射ポンプＦのハウジ
ング２を横外側に向つて一体に突設し、このハウ
ジング２の背部にバランサ装置Ｂが配置され、ハ
ウジング２の取付座へボルトで固定される。ま
た、燃料噴射ポンプＦの下部にはバランサ装置Ｂ
の前面に作業用油圧ポンプＰ２が装置されると共
に、バランサ装置Ｂの後面にも作業用油圧ポンプ
Ｐ１が装着される。そして、バランサ装置Ｂはクランク軸端のクラ
ンクギヤ３から調時ギヤ４，５を介して回転駆動
される噴射ポンプカム軸６に連動連結される。バランサ装置Ｂは、第３図〜第６図に示すよう
に、３軸の回転偏心錘７，８，９及びこれらを駆
動するためのギヤ類をケーシング３０内に組込ん
だ構造の独立したユニツトに形成され、クランク
軸Ｓに平行且つ上中下３段にバランサ軸１９，２
０，２１が配置され、各バランサ軸１９，２０，
２１の中央部には各偏心錘７，８，９が装着され
る。但し、各バランサ軸１９，２０，２１のクラン
ク軸Ｓに対する位置及び各偏心錘７，８，９の偏
心質量は後述の計算方法で設定される。各バラン
サ軸１９，２０，２１の後端部には同径のバラン
サギヤ１６，１７，１８が装着され、これらが噛
み合つて同じ回転数で回転するように連動連結さ
れている。他方、クランク軸Ｓの回転角速度をωとする
と、調時ギヤ４，５を介して噴射ポンプカム軸６
が1/2ωでクランク軸と同方向に回転駆動され、
この噴射ポンプカム軸６後端からバランサ装置Ｂ
に回転駆動力が入力される。即ち、噴射ポンプカム軸６後端には第１ギヤ１
０、油圧ポンプ駆動軸２２の前端には同軸上に第
２ギヤ１１・第３ギヤ１２、中段のバランサ軸２
０前端には第４ギヤ１３が装着されており、第１
ギヤ１０から第２ギヤ１１へは２倍増速で回転駆
動力が入力され、この回転駆動力が第３ギヤ１２
から第４ギヤ１３へ２倍増速で入力される。そし
て、バランサ軸１９，２０，２１後端のバランサ
ギア１６，１７，１８を介して各バランサ軸１
９，２０，２１は等速且つ隣り合うバランサ軸１
９，２０，２１が逆回転するよう連動連結されて
いるので、結局、各バランサ軸１９，２０，２１
は噴射ポンプカム軸６の４倍の回転角速度（即
ち、2ω）で回転駆動し、中段のバランサ軸２０
はクランク軸Ｓと同方向に、上下段のバランサ軸
１９，２１はクランク軸Ｓと反対方向に回転駆動
されるように形成される。また、各バランサ軸１９，２０，２１及び油圧
ポンプ駆動軸２２は、その前端部及び後端部をボ
ールベアリング２３，２４を介してケーシング３
０に支持される。バランサ装置Ｂの後部の作業用油圧ポンプＰ１
の入力軸２５は油圧ポンプ駆動軸２２の後端に連
動連結され、バランサ装置Ｂのケーシング３０の
後端側に形成された補機取付座２６に油圧ポンプ
Ｐ１がボルトで固定される。また、前側の作業用
油圧ポンプＰ２は第４ギヤ１３・第５ギヤ１４、
第６ギヤ１５を介して中段のバランサ軸２０に連
動連結され、この油圧ポンプＰ２もバランサ装置
Ｂのケーシング３０に固定される。更に、バランサ軸１９，２０，２１間の距離を
極力小さく設定して、バランサ装置Ｂの小型化を
図る為に、次のように形成される。第７図に示す
ように、隣り合う回転偏心錘７，８，９の回転運
動空間２７，２８，２９が干渉するようにバラン
サ軸１９，２０，２１間距離を小さく設定する
が、隣り合う回転偏心錘７，８，９の初期位相α
（後記の計算方法により決定される）の差により、
回転偏心錘７，８，９相互の干渉は起らないよう
に形成される。ここで、回転偏心錘７，８，９相
互の干渉を起りにくくする為、各回転偏心錘７，
８，９の両側の始端部７ａ，８ａ，９ａ及び終端
部７ｂ，８ｂ，９ｂを扁平にし、回転運動空間２
７，２８，２９の外周面より内側に退かせて形成
する。次に、上記の偏置多軸回転偏心錘式のバランサ
装置Ｂによつて、内燃機関の起振力の任意のｎ次
成分Ｆ→_Eを打ち消す力学的構成について説明する。
第８図に示し且つ後記のように、ｘ軸・ｙ軸・ｚ
軸を設定し、第１２図に示すように、クランク軸
軸心Ｏとピストン中心軸とで規定される起振方向
面Ｖを設定する。シリンダブロツクＤの横外側に
装着されているバランサ装置Ｂは、起振方向面Ｖ
から一定距離ｄだけ隔てて配設されている。そし
て、上中下の各回転偏心錘７，８，９の遠心力の
全合力のうち、そのｘ軸分力が内燃機関Ｅの起振
力の任意のｎ次成分Ｆ→_E（但し、実施例の場合ｎ
＝２）と比べて、その大きさがほぼ等しく方向が
逆となるように設定されており、上記ｘ軸分力で
起振力のｎ次成分Ｆ→_Eが打ち消される。また、全
合力のｙ軸成分は殆んど発生せず、ほぼ０に等し
くなるように設定される。そして、上記全合力の
ｘ軸分力と起振力のｎ次成分Ｆ→_Eとが所定距離ｄ
隔てられているために偶力をなし、この起振偶力
によつて生じる起振モーメントＭ→_Vを、複数の回
転偏心錘７，８，９の複数の遠心力による回転モ
ーメントの全合成モーメントＭ→_Bで打ち消す、つ
まり両者の大きさがほぼ等しく回転方向が逆とな
るように設定される。このように、内燃機関Ｅの起振力のｎ次成分Ｆ→
_Ｅは、偏置多軸回転偏心錘式バランサ装置Ｂで打
ち消され、且つ起振モーメントも発生しないの
で、内燃機関Ｅの機械振動が大幅に軽減され、そ
の振動、騒音が格段と改善される。尚、ｎは正整
数である。次に、上記偏置多軸回転偏心錘式バランサ装置
Ｂの主要諸元、即ちバランサ軸１９，２０，２１
のクランク軸Ｓに対する位置、偏心錘７，８，９
の偏心質量mi、偏心錘の偏心距離ri、偏心錘の初
期位相αi等を設定する為の計算方法について詳細
に説明する。 (1) 直列機関の起振力第８図に示すように、機関の重心を通りシリ
ンダ軸方向にｘ軸、クランク軸方向にｚ軸、こ
れら２軸に直角な方向にｙ軸（ただし以後の計
算の便宜上、座標系が右手系となるようにそれ
ぞれの正負の方向を決める）をとると、直列機
関には一般に第８図に示すような各軸方向の併
進振動および各軸まわりの回転振動を合成した
振動が生起する。これは、Ｆ→＝F_xｅ→_x＋F_yｅ→_y＋F_zｅ→_z ………(1) Ｍ→＝M_xｅ→_x＋M_yｅ→_y＋M_zｅ→_z ………(2) M_z＝M_Iz＋M_Gz ………(3) なる起振力Ｆ→と起振偶力Ｍ→によるものであり
（ただしM_zについては、後の便宜のためM_Izと
M_Gzの２成分に分けた）各成分はそれぞれ次の
ようになる。 F_x＝_L 〓^l=1 （F_x）_l ………(4) F_y＝_L 〓^l=1 （F_y）_l ………(5) F_z＝_L 〓^l=1 （F_z）_l ………(6) M_x＝_L 〓^l=1 Z_l・（F_y）_l ………(7) M_y＝_L 〓^l=1 Z_l・（F_x）_l ………(8) M_Iz＝_L 〓^l=1 （M_Iz）_l ………(9) M_Gz＝_L 〓^l=1 （M_Gz）_l ………(10) ここに（F_x）_l＝（ｍ＋m₀）ω²rcos（θ＋α_l）＋mω²r_∞ 〓ⁿ⁼¹ C_2ocos2n（θ＋α_l） ………(11) （F_y）_l＝m₀ω²rsin（θ＋α_l） ………(12) （F_z）_l＝０ ………(13) （M_Iz）_l＝−mω²r² _∞ 〓ⁿ⁼¹ D_oｎ／２sin ｎ（θ＋α_l）−I₁ω² _∞ 〓ⁿ⁼¹ ｎ・H_2osin2n（θ＋α_l） −I₁ω² _∞ 〓ⁿ⁼¹ G_2o+1sin（2n＋１）（θ＋α_l） ………(14) （M_Gz）_l＝−πD²／４・ｒ｛_∞ 〓ⁿ⁼¹ b_o/2cosｎ／２（θ＋β_l）＋_∞ 〓ⁿ⁼¹ a_o/2sinｎ／２（θ＋β_l）｝ ………(15) ｅ→_x，ｅ→_y，ｅ→_z：ｘ、ｙ、ｚ方向基底ベクトル（F_x）_l，（F_y）_l，（F_z）_l：第１番目気筒のｘ、ｙ
、
ｚ軸方向の起振力（M_x）_l、（M_y）_l：第１番目気筒のｘ、ｙ、軸ま
わりの起振偶力（M_Iz）_l：第１番目気筒の往復運動部質量およ
び修正慣性モーメントによるｚ軸まわりの起
振偶力（M_Gz）_l：第１番目気筒のガス圧力によるｚ軸
まわりの起振偶力Ｌ：気筒数Ｄ：シリンダ径ｒ：クランク半径ｍ：１気筒あたりの往復運動部質量 m₀：１気筒あたりの回転運動部質量 θ：第１気筒上死点よりのクランク軸回転角 ω：クランク軸回転角速度（dθ／dt） Z_l：１番目気筒のｚ座標 α_l：１番目気筒と１番目気筒のクランク位相差
（クランク角） β_l：１番目気筒と１番目気筒の着火間隔位相差
（クランク角） I₁：１気筒あたりのコンロツド修正慣性モーメ
ント ρ：ｒ／lo（lo：コンロツドの長さ） The present invention relates to a rotating eccentric weight type balancer device that cancels out any n-order component of an excitation force caused by reciprocating inertia of a crank and piston mechanism of an internal combustion engine. The first invention provides a balancer device that is spaced a certain distance from a plane in the vibration direction passing through the crankshaft and the piston center axis of an internal combustion engine, and uses the resultant force of the centrifugal force of the multi-axis rotating eccentric weight to generate an arbitrary vibration force. Provided is a balancer device which cancels out the n-order component of the centrifugal force and at the same time cancels out the excitation moment due to the couple generated by the resultant force of the centrifugal force and the n-order component of the excitation force with the total resultant moment due to the centrifugal force, In addition, the distance between the axes of the rotating eccentric weights is set as small as possible by preventing the interference of the rotating eccentric weights by using the initial phase difference of the rotating eccentric weights while allowing the rotational motion spaces of the rotating eccentric weights to interfere with each other. The purpose is to downsize the balancer device. The second invention incorporates the balancer device of the first invention into a unit, and the unit is installed on the back of a fuel injection pump protruding from the outside of a water-cooled diesel engine, and the balancer device is connected to the crankshaft via the camshaft of the injection pump. The purpose of this device is to utilize the dead space outside the diesel engine to equip the balancer device by interlocking and connecting the balancer device to the diesel engine. In-line internal combustion engines have been widely used as engines for various industrial machines because they can achieve variations in output by varying the number of cylinders and cylinder diameters, but in addition to their output performance, they also suffer from vibration and noise. is becoming an important issue. For example, agricultural machinery and the like do not have a structure in which the engine is supported in a vibration-proof manner, and the key to reducing the vibration of each part of the machine is to reduce the engine vibration itself, which is the main source of vibration. One possible solution to this problem is to reduce engine vibration by increasing the number of cylinders, but since the displacement per cylinder becomes smaller and fuel consumption increases, there is a limit to increasing the number of cylinders. Therefore, in order to reduce engine vibration, it becomes necessary to install a balancer device, but the existing balancer devices have the following problems. First, in existing internal combustion engines, in order to alleviate the primary component of the vertical excitation force caused by the vibration of the reciprocating inertia force of the piston and crank mechanism, a counterweight for the crank pin part is generally added to the outside of the crank arm. be. However, this counterweight cannot cancel the second-order or higher-order components of the excitation force. Therefore, in order to cancel, for example, the second-order component of the excitation force, conventionally, a pair of rotating eccentric weights W is provided directly below the crankshaft S, as shown by the imaginary line in FIG. A balancer device B' has been used which is structured to generate a resultant force within a plane V in the vibration direction defined by the crankshaft S and the piston center axis.
However, in this structure, since the balancer device B' is installed directly below the crankshaft, the disadvantage is that the height of the engine becomes considerably high. Therefore, if the balancer device were to be installed on the side of the internal combustion engine, the vertical excitation force would be canceled by the resultant force of the centrifugal force of the rotating eccentric weight, but the excitation force and the centrifugal force would form a couple. An excitation moment is generated, which becomes a source of vibration for the engine. However,
When a 100% counterweight is added to the crank arm, the excitation force is distributed horizontally, so a balancer device for canceling out the secondary component of the excitation force is used as shown by the symbol B'' on the engine. Although it can be installed right next to the crankshaft S, this structure inevitably has the drawback that the width of the engine increases considerably.The present invention uses an eccentric multi-axis rotary eccentric weight balancer device. The above-mentioned drawbacks are solved and the balancer device is miniaturized. That is, a multi-axis rotating eccentric weight type balancer device B having at least three axes is installed at a predetermined distance d from the plane V in which the vibration is generated. At the same time, the desired n-dimensional component F→ _E of the vibrational force of the engine is balanced by the resultant force of a plurality of centrifugal forces of these rotating eccentric weights.
The total resultant moment of centrifugal force M→ _B is balanced with the excitation moment M→ _V due to the excitation couple consisting of the resultant force of centrifugal force and the n-dimensional component F→ _E of excitation force, and each of the balancer devices Among the rotating eccentric weights, the distance between their axes is set short so that the respective rotational motion spaces of at least two adjacent rotating eccentric weights interfere with each other,
The two rotating eccentric weights are characterized in that they are interlocked and connected so that they rotate in opposite directions at equal rotational speeds so that they can rotate without interference due to the difference in their initial phases α. Next, embodiments of the present invention will be described based on the drawings. 1 and 2 show schematic diagrams of a water-cooled vertical inline four-cylinder diesel engine E, in which a cylinder block D is integrally formed in the upper part of a crankcase C, and a cylinder head block H is formed in the upper part of this cylinder block D. -Equipped with a valve train cam chamber M and a fuel injection valve N, and a gear case G is installed at the front end of the engine. A walk jacket is disposed inside the cylinder block D, and a cylinder is formed in the center of the walk jacket. A housing 2 of a fuel injection pump F is integrally protruded from the front part of the horizontal outer wall 1 of the water jacket of this cylinder toward the outside laterally, and a balancer device B is disposed at the back of this housing 2. It is bolted to the seat. In addition, a balancer device B is installed at the bottom of the fuel injection pump F.
A working hydraulic pump P2 is installed on the front side of the balancer device B, and a working hydraulic pump P1 is also installed on the rear side of the balancer device B. The balancer device B is interlocked and connected to an injection pump camshaft 6 which is rotationally driven from a crank gear 3 at the end of the crankshaft via timing gears 4 and 5. As shown in FIGS. 3 to 6, the balancer device B is an independent unit having a structure in which three-axis rotating eccentric weights 7, 8, and 9 and gears for driving these are incorporated in a casing 30. Balancer shafts 19, 2 are formed parallel to the crankshaft S and are arranged in upper, middle, and lower three stages.
0, 21 are arranged, and each balancer shaft 19, 20,
Eccentric weights 7, 8, and 9 are attached to the central portion of 21. However, the position of each balancer shaft 19, 20, 21 with respect to the crankshaft S and the eccentric mass of each eccentric weight 7, 8, 9 are set by the calculation method described later. Balancer gears 16, 17, and 18 having the same diameter are attached to the rear ends of each of the balancer shafts 19, 20, and 21, and are interlocked and connected so that they mesh and rotate at the same number of rotations. On the other hand, if the rotational angular velocity of the crankshaft S is ω, the injection pump camshaft 6
is rotated in the same direction as the crankshaft at 1/2ω,
From the rear end of this injection pump camshaft 6 to the balancer device B
Rotational driving force is input to. That is, the first gear 1 is located at the rear end of the injection pump camshaft 6.
0. At the front end of the hydraulic pump drive shaft 22, there are a second gear 11 and a third gear 12 coaxially, and a balancer shaft 2 in the middle stage.
A fourth gear 13 is attached to the front end of the
A rotational driving force is input from the gear 10 to the second gear 11 at twice the speed, and this rotational driving force is input to the third gear 12.
is inputted to the fourth gear 13 at twice the speed. Then, each balancer shaft 1
9, 20, 21 are constant velocity and adjacent balancer shafts 1
Since the balancer shafts 9, 20, and 21 are interlocked to rotate in the opposite direction, each balancer shaft 19, 20, and 21
is rotated at a rotational angular velocity four times that of the injection pump camshaft 6 (i.e., 2ω), and the middle balancer shaft 20
is formed to be rotated in the same direction as the crankshaft S, and the upper and lower balancer shafts 19 and 21 are rotationally driven in the opposite direction to the crankshaft S. Further, each of the balancer shafts 19, 20, 21 and the hydraulic pump drive shaft 22 has its front end and rear end connected to the casing 3 via ball bearings 23, 24.
Supported by 0. Hydraulic pump P1 for work at the rear of balancer device B
The input shaft 25 is interlocked and connected to the rear end of the hydraulic pump drive shaft 22, and the hydraulic pump P1 is fixed with bolts to an auxiliary equipment mounting seat 26 formed on the rear end side of the casing 30 of the balancer device B. In addition, the front working hydraulic pump P2 has a fourth gear 13, a fifth gear 14,
The hydraulic pump P2 is interlocked and connected to the middle stage balancer shaft 20 via the sixth gear 15, and this hydraulic pump P2 is also fixed to the casing 30 of the balancer device B. Further, in order to reduce the size of the balancer device B by setting the distance between the balancer shafts 19, 20, and 21 as small as possible, the balancer device B is formed as follows. As shown in FIG. 7, the distance between the balancer shafts 19, 20, 21 is set small so that the rotational motion spaces 27, 28, 29 of the adjacent rotating eccentric weights 7, 8, 9 interfere with each other. Initial phase α of eccentric weights 7, 8, 9
(determined by the calculation method described below),
The rotating eccentric weights 7, 8, and 9 are formed so as not to interfere with each other. Here, in order to prevent mutual interference between the rotating eccentric weights 7, 8, and 9, each rotating eccentric weight 7,
Starting ends 7a, 8a, 9a and terminal ends 7b, 8b, 9b on both sides of 8, 9 are flattened to create a rotational movement space 2.
7, 28, and 29 to the inside. Next, a description will be given of a mechanical configuration for canceling any n-order component F→ _E of the excitation force of the internal combustion engine using the eccentric multi-axis rotating eccentric weight type balancer device B described above.
As shown in Figure 8 and described later, the x-axis, y-axis, z
The axis is set, and as shown in FIG. 12, a vibration generating direction plane V defined by the crankshaft axis O and the piston center axis is set. The balancer device B installed on the lateral outer side of the cylinder block D is
It is arranged at a fixed distance d from the center. Of the total resultant force of the centrifugal forces of the upper, middle, and lower rotating eccentric weights 7, 8, and 9, the x-axis component is an arbitrary n-dimensional component of the excitation force of the internal combustion engine _E In the example case n
= 2), the magnitudes are set to be approximately equal and the directions are opposite, and the n-th order component F→ _E of the excitation force is canceled out by the x-axis component force. Further, the y-axis component of the total resultant force is hardly generated and is set to be approximately equal to 0. Then, the x-axis component of the total resultant force and the n-order component F→ _E of the excitation force are separated by a predetermined distance d.
Since they are separated, they form a couple, and the excitation moment M→ _V generated by this excitation couple is the total resultant moment of the rotational moment due to the multiple centrifugal forces of the multiple rotating eccentric weights 7, 8, and 9. It is set so that M→ _B cancels out, that is, the sizes of both are approximately equal and the rotation directions are opposite. In this way, the n-dimensional component F of the excitation force of the internal combustion engine E →
_E is canceled by the eccentric multi-axis rotating eccentric weight balancer device B, and no vibration moment is generated, so the mechanical vibration of the internal combustion engine E is significantly reduced, and its vibration and noise are significantly improved. Note that n is a positive integer. Next, the main specifications of the eccentric multi-axis rotating eccentric weight type balancer device B, that is, the balancer shafts 19, 20, 21
position relative to the crankshaft S, eccentric weights 7, 8, 9
A calculation method for setting the eccentric mass mi, the eccentric distance ri of the eccentric weight, the initial phase αi of the eccentric weight, etc. will be explained in detail. (1) Excitation force of a series engine As shown in Figure 8, the x-axis passes through the center of gravity of the engine in the direction of the cylinder axis, the z-axis runs in the direction of the crankshaft, and the y-axis runs in a direction perpendicular to these two axes (however, in the following For convenience of calculation, if the positive and negative directions are determined so that the coordinate system is right-handed, then a series engine will generally have translational vibrations in each axis direction and rotational vibrations around each axis as shown in Figure 8. A vibration is generated that is a combination of the two. _This _is _, _{_} _{_} _{_} _{_} _{_} _{_} _{_} _{_} _{_} ) M _z = M _Iz + M _Gz ......(3) This is due to the excitation force F→ and the excitation couple M→ (However, for later convenience, M _z will be referred to as M _Iz and
Each component (divided into two components of M _Gz ) is as follows. F _x = _L 〓 ^l=1 (F _x ) _l ………(4) F _y = _L 〓 ^l=1 (F _y ) _l ………(5) F _z = _L 〓 ^l=1 (F _z ) _l ………(6) M _x = _L 〓 ^l=1 Z _l・(F _y ) _l ………(7) M _y = _L 〓 ^l=1 Z _l・(F _x ) _l ………(8 ) M _Iz = _L 〓 ^l=1 (M _Iz ) _l ………(9) M _Gz = _L 〓 ^l=1 (M _Gz ) _l ………(10) Here (F _x ) _l = (m+m ₀ )ω ² rcos (θ+α _l ) +mω ² r _∞ 〓 ⁿ⁼¹ C _2o cos2n (θ+α _l ) ………(11) (F _y ) _l = m ₀ ω ² rsin (θ+α _l ) ………(12) (F _z ) _l = 0 ………(13) (M _Iz ) _l =−mω ² r ² _∞ 〓 ⁿ⁼¹ D _o n/2sin n(θ+α _l )−I ₁ ω ² _∞ 〓 ⁿ⁼¹ n・H _2o sin2n (θ+α _l ) −I ₁ ω ² _∞ 〓 ⁿ⁼¹ G _2o+1 sin (2n+1) (θ+α _l ) ………(14) (M _Gz ) _l =−πD ² /4・r{ _∞ 〓 ⁿ⁼¹ b _o/2 cosn/2 (θ+β _l )+ _∞ 〓 ⁿ⁼¹ a _o/2 sinn/2 (θ+β _l )} ………(15) e→ _x , e→ _y , e→ _z : x, y, z direction basis vector (F _x ) _l , (F _y ) _l , (F _z ) _l : x, y of the first cylinder
,
Excitation force in the z-axis direction (M _x ) _l , (M _y ) _l : Excitation couple around the x, y, and axes of the first cylinder (M _Iz ) _l : Mass of the reciprocating part of the first cylinder and the excitation couple around the z-axis due to the modified moment of inertia (M _Gz ) _l : Excitation couple around the z-axis due to the gas pressure of the first cylinder L: Number of cylinders D: Cylinder diameter r: Crank radius m: 1 Mass of reciprocating parts per cylinder m ₀ : Mass of rotating parts per cylinder θ : Crankshaft rotation angle from the top dead center of the first cylinder ω : Crankshaft rotational angular velocity (dθ/dt) Z _l : Crankshaft rotation angle from the top dead center of the first cylinder Z coordinate α _l : Crank phase difference between the 1st cylinder and the 1st cylinder (crank angle) β _l : Ignition interval phase difference between the 1st cylinder and the 1st cylinder (crank angle) I ₁ : Conrod corrected moment of inertia per cylinder ρ: r/lo (lo: length of conrod)

【表】 …
…
〓１ 3
【table】 …
…
〓 1 3

Claims

[Claims] 1. Vibration direction plane V on which the excitation force of the internal combustion engine E acts
Assuming an excitation direction plane V that passes through the axis O of the crankshaft S and the piston center axis, a balancer device B is placed on one side at a predetermined distance d from the excitation direction plane V, and the balancer device B is relatively fixed to the engine body E, and the balancer device B is configured such that at least three rotating eccentric weights 7, 8, and 9 can be rotationally driven by a crankshaft S, and the centrifugal rotation of these rotating eccentric weights 7, 8, and 9 is Of the total resultant force, x parallel to the direction of piston movement
The axial component force is set to a value that is approximately equal in magnitude and opposite in direction compared to any n-dimensional component F→ _E of the vibrational force of the engine E, and the y-axis component force perpendicular to this is set to approximately 0. The total resultant moment M→ _B generated by the centrifugal force of the rotating eccentric weights 7, 8, and 9 is set so that any n-dimensional component F→ _E of the excitation force of the engine E, x-axis component force and distance d between these two
Compared to the excitation moment M→ _V due to the excitation couple resulting from the relationship, the magnitudes are set to be approximately equal and the rotation direction is opposite, and each rotating eccentric weight 7, 8, 9 of the balancer device B is Among them, each rotational movement space 2 of at least two adjacent rotating eccentric weights 7, 8, 9
7, 28, 29 interfere with each other, the distance between their axes is set short, and both rotating eccentric weights 7, 8,
Reference numeral 9 refers to a rotary eccentric weight type balancer device for an internal combustion engine, characterized in that they are interlocked and connected so that they rotate in opposite directions at equal rotational speeds so that they can rotate without interference due to the difference in their initial phases α. 2. In the balancer device described in claim 1, the total resultant moment M→ _B caused by the rotating eccentric weights 7, 8, and 9 is applied to the crankshaft S of the engine E by the excitation moment M→ _V due to the excitation couple. Compared to the sum of the arbitrary n-dimensional components M→ _E of the excitation moment around the parallel Z-axis, the magnitude is approximately equal and the direction of rotation is set to be opposite. 3. In the balancer device according to claim 1 or 2, there are three rotating eccentric weights 7, 8, and 9, and one of them is arranged so that the rotation direction is opposite to that of the other two. What I did. 4. In the balancer device according to claim 1, 2 or 3, the rotational movement space 2
7, 28, 29 interfere with each other, each rotating eccentric weight 7,
The starting end portions 7a, 8a, 9a and the terminal end portions 7b, 8b, 9b in the rotational direction of 8, 9 are connected to the rotational eccentric weight 7,
8, 9 is formed by retreating inward from the outer peripheral surfaces of the rotational movement spaces 27, 28, 29. 5. In the balancer device according to any one of claims 1 to 4, an internal combustion engine E
is a water-cooled type, and a balancer device B is installed on the lateral outer side of the lateral outer wall 1 of the water jacket of the cylinder.
Placed. 6. In the balancer device described in any one of claims 1 to 5, the balancer device B includes rotating eccentric weights 7, 8, 9 in the casing 30.
This balancer device unit is constructed independently as an internal unit, and this balancer device unit is fixed to the engine body E. 7 Excitation direction plane V on which the excitation force of internal combustion engine E acts
Assuming an excitation direction plane V that passes through the axis O of the crankshaft S and the piston center axis, a balancer device B is placed on one side at a predetermined distance d from the excitation direction plane V, and the balancer device B is relatively fixed to the engine body E, and the balancer device B is configured such that at least three or more rotating eccentric weights 7, 8, and 9 can be rotationally driven by a crankshaft S, and the centrifugal rotation of these rotating eccentric weights 7, 8, and 9 is Of the total resultant force, x parallel to the direction of piston movement
The axial component force is set to a value that is approximately equal in magnitude and opposite in direction compared to any n-dimensional component F→ _E of the vibrational force of the engine E, and the y-axis component force perpendicular to this is set to approximately 0. The total resultant moment M→ _B generated by the centrifugal force of the rotating eccentric weights 7, 8, and ₉ is set so that Compared to the excitation moment M→ _V due to the excitation couple resulting from the relationship between the component force and the distance d between the two, the magnitudes are set to be approximately equal and the direction of rotation is opposite, and the balancer device B is Each rotational movement space 27, 2 of at least two adjacent rotational eccentric weights 7, 8, 9 among the rotational eccentric weights 7, 8, 9
The distance between the axes 8 and 29 is set short so that they interfere with each other, and the rotating eccentric weights 7, 8, and 9 rotate equally with each other so that they can rotate without interference due to the difference in their initial phases α. The internal combustion engine E is a water-cooled diesel internal combustion engine, and the housing 2 of the fuel injection pump F is directed laterally outward from the front part of the lateral outer wall 1 of the walk jacket of the cylinder. The balancer device B has each rotary eccentric weight 7,
Constructed independently as a unit with 8 and 9 inside,
This balancer device unit is arranged at the back of the housing 2 of the fuel injection pump F and fixed to the pump housing 2, and each rotating eccentric weight 7, 8, 9 of the balancer device B is cranked via the fuel injection pump camshaft 6. A rotating eccentric weight type balancer device for an internal combustion engine, characterized in that it is interlocked and connected to a shaft S. 8 In the balancer device described in claim 7, the total resultant moment M→ _B caused by the rotating eccentric weights 7, 8, and 9 is applied to the crankshaft S of the engine E by the excitation moment M→ _V due to the excitation couple. Compared to the sum of arbitrary n-order components M→ _E of the excitation moment around the parallel Z-axis, the magnitude is approximately equal and the direction of rotation is set to be opposite. 9 In the balancer device according to claim 7 or 8, there are three rotating eccentric weights 7, 8, and 9, and one of them is arranged so that the rotation direction is opposite to that of the other two. What I did. 10 In the balancer device according to claim 7, 8, or 9, an auxiliary equipment mounting seat 26 is formed on the outer surface of the casing 30 of the balancer device B, and each rotating eccentric weight 7, 8, 9 An accessory drive shaft 22 is attached to the rotational transmission mechanism, and the accessory drive shaft 22 faces an accessory mounting seat 26. 11. The balancer device according to any one of claims 7 to 10, wherein the water-cooled diesel internal combustion engine is of an in-line multi-cylinder type.