TW201231842A

TW201231842A - Flexible engagement gear device and method for determining shape of gear tooth of flexible engagement gear device

Info

Publication number: TW201231842A
Application number: TW100102854A
Authority: TW
Inventors: Shinji Yoshida; Masaaki Shiba; Manabu Andoh
Original assignee: Sumitomo Heavy Industries
Priority date: 2011-01-26
Filing date: 2011-01-26
Publication date: 2012-08-01
Also published as: TWI425155B

Abstract

The purpose of this invention is to increase shock resistance and further enhance transmission torque and transmission efficiency. The flexible engagement gear device (100) of this invention comprises: cylinder-shaped external gears (120A, 120B) which are flexible; and an internal gear for deceleration (130A) and an internal gear for output (130B). The external gears (120A, 120B) respectively have rigidity for internal engagement, wherein shapes of gear teeth of the external gears (120A, 120B) that are engaged respectively with the internal gear for deceleration (130A) and the internal gear for output (130B) are identical to each other. The external gears (120A, 120B), the internal gear for deceleration (130A), and the internal gear for output (130B) have gear tooth shapes such that both the number of simultaneous engagement teeth (Nph) between the external gear (120A) and the internal gear for deceleration (130A); and the number of simultaneous engagement teeth (Npl) between the external gear (120B) and the internal gear for output (130B) can be greater than two.

Description

201231842 六、發明說明【發明所屬之技術領域】本發明係有關一種撓性咬合式齒輪裝置及撓性咬合式齒輪裝置之齒形決定方法。【先前技術】專利文獻1所示之撓性咬合式齒輪裝置，具備有：震盪體；筒形外齒輪，配置於該震盪體之外周並具有依該震盪體之旋轉而撓性變形的可撓性；第1內齒輪，具有該外齒輪內咬合的剛性；及第2內齒輪，軸向上與該第1內齒輪並設，且具有與前述外齒輪內咬合的剛性》因此，當第1內齒輪固定於外殼時，依震盪體之旋轉撓性變形的外齒輪內咬合於第1內齒輪，外齒輪依據與第 1內齒輪之齒數差減速。而且，可從第2內齒輪取出其被減速後的外齒輪之輸出。專利文獻1 :日本特開2 006-295 08號公報【發明內容】 (本發明所欲解決之課題）然而，在如專利文獻1所示之撓性咬合式齒輪裝置中，因必須藉由使外齒輪撓性來實現與內齒輪之咬合以及當爲筒形外齒輪時，必須同時檢討與2個內齒輪之咬合等理由，很難使2個內齒輪與外齒輪在理論上相咬合，且作爲剛體齒輪之理論咬合數非常少。因此，使用以往的筒形 -5- 201231842 外齒輪的撓性咬合式齒輪裝置其耐衝擊性低並且傳遞扭矩小，其傳遞效率亦低。因此，本發明係解決前述問題點而完成者，其課題在於提供一種能夠提高耐衝擊性’並使傳遞扭矩或傳遞效率進一步增大的撓性咬合式齒輪裝置及撓性咬合式齒輪裝置之齒形決定方法。 (用以解決課題之手段）本發明係藉由如下解決前述課題者：一種撓性咬合式齒輪裝置，具備：震盪體；筒形外齒輪，配置於該震盪體之外周並具有依該厲盪體之旋轉而撓性變形的可撓性；第 1內齒輪，具有與該外齒輪內咬合的剛性；及第2內齒輪，軸向上與該第1內齒輪並設，且具有與前述外齒輪內咬合的剛性，其中，前述外齒輪分別與前述第1內齒輪及前述第2內齒輪咬合的部份之齒形相同，前述外齒輪、第 1內齒輪及第2內齒輪分別具備如該外齒輪與第1內齒輪之同時咬合數及該外齒輪與第2內齒輪之同時咬合數均成爲2以上的齒形。本發明係使外齒輪、第1內齒輪及第2內齒輪具備將外齒輪與2個內齒輪（第1內齒輪和第2內齒輪）之同時咬合數均設爲2以上之齒形者。因此，能夠提高耐衝擊性，分散施加於咬合齒面之面壓，且傳遞大扭矩。而且，本發明中，作爲其基本結構，與具備有使筒形外齒輪咬合於2個具有剛性的內齒輪之結構的情況相結合，能夠提高 -6- 201231842 耐棘輪性，並且能夠使無負荷時產生於外齒輪之應力少於杯形外齒輪，且能夠增加負荷容量。因此，本發明能夠增大傳遞扭矩，並且能夠增大傳遞效率。又’外齒輪之齒形於分別與第1內齒輪及第2內齒輪咬合的部份相同，故容易加工外齒輪，能夠較低地抑制加工成本，並且能夠實現高精度的形狀加工。又，本發明係藉由如下解決前述課題：一種撓性咬合式齒輪裝置’具備：震盪體；筒形外齒輪，配置於該震盪體之外周並具有依該震盪體之旋轉而撓性變形的可撓性；第1內齒輪，具有該外齒輪內咬合的剛性；及第2內齒輪’軸向上與該第1內齒輪並設，且具有與前述外齒輪內咬合的剛性’其中，當將該外齒輪之外齒設爲圓筒形銷時或假想爲圓筒形銷時，或者，將該第1內齒輪或第2內齒輪之內齒設爲筒柱形銷時或假想爲圓筒形銷時，於穿過前述震盪體之旋轉軸和與前述第1內齒輪或第2內齒輪咬合時前述外齒輪的咬合半徑之中心亦即偏心軸的直線與由該外齒輪與該第1內齒輪及第2內齒輪之咬合產生的接觸點之各個共同法線之交點亦即節距點之間配置該銷中心。將桌1內齒輪或第2內齒輪之內齒假想爲圓筒形銷時，具體而言’依據該假想的銷求出外齒，並依據該外齒形成第i 內齒輪及第2內齒輪之內齒作爲包絡線。本發明中，當將外齒輪之外齒設爲圓筒形銷時，或者將該第1內齒輪或第2內齒輪之內齒設爲圓筒形銷時，於上述2個節距點之間配置該銷中心。因此，與第〗內齒輪 201231842 咬合時施加於筒形外齒輪之外齒的負載和與第2內齒合時施加於筒形外齒輪之外齒的負載具備相互反方向份’並且能夠在外齒輪之外周方向上使施加於外齒輪 2個負載的區域靠近。亦即，從軸向觀察時，進行咬作時2個內齒輪可設爲僅夾入少數外齒之狀態。因此其可以防止外齒輪與內齒輪之咬合因過度扭矩而偏移象（棘輪現象）。亦即，本發明尤其著眼於進行棘輪高之情況，能夠增大所容許之傳遞扭矩，並且能夠增遞效率。根據本發明，能夠提高耐衝擊性，並使傳遞扭矩遞效率增大。【實施方式】以下，參照附圖詳細說明本發明之實施形態之一《第1實施形態》 <結構> 最開始主要利用第1圖和第2圖對本實施形態之結構進行槪略說明。撓性咬合式齒輪裝置100，具有：震盪體104: 輪120A、120B (僅作爲外齒輪120)，配置於震盪儷之外周並具有依震盪體104之旋轉而撓性變形的可撓及作爲第1內齒輪之減速用內齒輪13 0A和作爲第2 輪之輸出用內齒輪130B，具有外齒輪120分別內咬輪咬之成之該合動，尤之現性提大傳及傳例。整體外齒 1 04 性：內齒合的 201231842 剛性。再者，以後將減速用內齒輪1 3 Ο A和輸出用內齒輪 1 3 Ο B僅統稱爲內齒輪1 3 〇。以下’對各構成要件進行詳細說明。如第3圖（A)、第3圖（B)所示，前述震盪體1〇4 爲柱形’於中央形成有***未圖示的輸入軸的輸入軸孔 106。於輸入軸孔1〇6設置有鍵槽108，以使當***輸入軸並旋轉時，震盪體104與輸入軸一體旋轉。如第3圖、第4圖所示，震盪體104以將2個圓弧部 (第1圓弧部FA、第2圓弧部SA)連接在一起的形狀 (雙圓弧形狀）構成。第1圓弧部FA是以點B(稱爲偏心軸）爲中心的曲率半徑r 1之圓弧，構成用以使外齒輪 120和內齒輪130咬合的圓弧部份（亦稱爲咬合範圍）。第2圓弧部S A是以點C爲中心的曲率半徑r 2之圓弧，構成外齒輪120和內齒輪130互不咬合的範圍之圓弧部份 (亦稱爲非咬合範圍）。第1圓弧部FA之長度由作爲長軸X與點A處之法線N所成的角度之咬合角度Θ來決定。此時，如第4圖所示，若將震盪體1 04之長軸X之半徑設爲r ’則將偏心量設爲L，用式（1 )表示第1圓弧部 F A之曲率半徑r 1 ° r l=r -L ... ( 1) 又’如第4圖所示，在第1圓弧部fa與第2圓弧部 S A之連接部份A通用切線τ (法線N )。因此，第2圓弧部SA之曲率半徑Γ2爲（曲率半徑ri+長度BC)，故 201231842 用式（2 )表示。 r 2= r 1 +長度 BC =r 1 +L/cos9 ... ( 2 ) 震盪體軸承110A爲配置於震盪體丨04之外側與外齒輪12 0 A內側之間的軸承，如第2圖、第5圖所示，包括內圈112、保持器114A'作爲轉動體之滾子116A及外圈 118A。內圈112之內側與震盪體104抵接，內圏112與震盪體104—體變形的同時旋轉。滾子ι16Α爲圓筒形 (包括滾針）。因此，與轉動體爲球時相比，增加了滾子 116A與內圏112及外圈118A接觸之部份，故可加大負荷容量。亦即，藉由使用滾子1 16A，可使震盪體軸承1 10A 之傳遞扭矩增大，且可使之長壽命化。外圈118A配置於滾子M6A之外側。外圈11 8A依震盪體104之旋轉而撓性變形，並使配置於其外側的外齒輪1 2 0 A變形。再者，如第2圖所示，震盪體軸承U0B與震盪體軸承110A相同，由內圈112、保持器114B、滾子116B及外圈11 8B所構成。震盪體104及內圈112在震盪體軸承 110A、110B中通用。而且，保持器114B、滾子116B及外圈118B作爲單體構件（組件），與保持器114A '滾子 116A及外圈118A相同》如第2圖所示，外齒輪120A與減速用內齒輪130A 內咬合。外齒輪120A由基礎構件122和外齒124A所構成。基礎構件122爲支承外齒124A且具有可撓性的筒形構件，配置於震盪體軸承Π 〇A之外側。外齒1 24A成爲 -10- 201231842 半徑Ρ1之圓筒形銷（因此’本實施形態之外齒 (124Β )或外齒輪12〇a ( 120Β )或撓性咬合式齒輪 1〇〇亦僅稱爲銷型）》外齒124Α由環形構件126Α保基礎構件122上。如第2圖所示，外齒輪120Β與輸出用內齒輪內咬合。而且，外齒輪120Β與外齒輪120Α相同，礎構件122和外齒124Β所構成。外齒124Β的數量齒124 Α相同並且由相同的圓筒形銷構成，且由環形 126B保持於基礎構件122上。亦即，基礎構件122 支承外齒124A和外齒124B。亦即，外齒輪120A、爲相同形狀之齒形。震盪體1 04之偏心量L以同相位外齒124A和外齒124B。以後，將外齒124A、124B 爲外齒124。如第2圖所示，減速用內齒輪1 3 0 A由具有剛性件形成。減速用內齒輪130A具備比外齒輪120A的 124A之齒數僅多2倍數之齒數（關於齒數在後面進細敘述）。於減速用內齒輪1 3 Ο A上’透過螺栓孔固定未圖示之外殻。而且，減速用內齒輪13 0A藉由齒輪120A咬合，而有助於震盪體1〇4的旋轉之減速 6圖（A )中表示外齒輪120A和減速用內齒輪130A 的樣子，第7圖（A)中表示x軸上的外齒124A和 1 2 8 A之樣子。另一方面，輸出用內齒輪13 0B亦與減速用內 1 3 0 A相同，由具有剛性的構件形成。輸出用內齒輪201231842 VI. Description of the Invention [Technical Field] The present invention relates to a method for determining the tooth profile of a flexible snap-in gear device and a flexible snap-in gear device. [Prior Art] The flexible snap-in gear device disclosed in Patent Document 1 includes an oscillating body and a cylindrical external gear that is disposed on the outer periphery of the oscillating body and has flexibility to be flexibly deformed in accordance with the rotation of the oscillating body. The first internal gear has a rigidity that is engaged in the external gear; and the second internal gear is axially disposed with the first internal gear and has a rigidity that engages with the external gear. Therefore, when the first inner When the gear is fixed to the outer casing, the outer gear that is flexibly deformed by the rotation of the oscillating body is engaged with the first internal gear, and the external gear is decelerated according to the difference in the number of teeth from the first internal gear. Further, the output of the decelerated external gear can be taken out from the second internal gear. [Problem to be Solved by the Invention] However, in the flexible snap gear device shown in Patent Document 1, it is necessary to When the external gear is flexible to achieve engagement with the internal gear and when it is a cylindrical external gear, it is necessary to simultaneously review the engagement with the two internal gears, etc., and it is difficult to theoretically engage the two internal gears with the external gear, and The theoretical number of occlusions for rigid body gears is very small. Therefore, the flexible snap-in gear device using the conventional cylindrical -5-201231842 external gear has low impact resistance and small transmission torque, and its transmission efficiency is also low. Therefore, the present invention has been made to solve the above problems, and an object thereof is to provide a flexible snap-in gear device and a tooth of a flexible snap-in gear device capable of improving impact resistance and further increasing transmission torque or transmission efficiency. Shape determination method. (Means for Solving the Problems) The present invention solves the above-mentioned problems by providing a flexible snap-in gear device including an oscillating body and a cylindrical external gear disposed on the outer periphery of the oscillating body and having the same a flexible body that is flexibly deformed by rotation; a first internal gear having rigidity engaged with the external gear; and a second internal gear axially coupled to the first internal gear and having the external gear The rigidity of the inner engagement is the same as the tooth shape of the portion of the external gear that meshes with the first internal gear and the second internal gear, and the external gear, the first internal gear, and the second internal gear are respectively provided The number of simultaneous engagement of the gear and the first internal gear and the number of simultaneous engagement of the external gear and the second internal gear are both two or more. In the present invention, the external gear, the first internal gear, and the second internal gear are provided with a tooth shape in which the number of engagements between the external gear and the two internal gears (the first internal gear and the second internal gear) is two or more. Therefore, it is possible to improve the impact resistance, disperse the surface pressure applied to the nip surface, and transmit a large torque. Further, in the present invention, as a basic configuration, in combination with a configuration in which the cylindrical external gear is engaged with two rigid internal gears, the ratchet resistance of -6-201231842 can be improved, and no load can be applied. The stress generated in the external gear is smaller than that of the cup-shaped external gear, and the load capacity can be increased. Therefore, the present invention can increase the transmission torque and can increase the transmission efficiency. Further, since the tooth profile of the external gear is the same as the portion that is engaged with the first internal gear and the second internal gear, the external gear can be easily processed, the processing cost can be suppressed low, and high-precision shape processing can be realized. Further, the present invention solves the above-described problems by providing a flexible snap-in gear device that includes an oscillating body and a cylindrical external gear that is disposed on the outer circumference of the oscillating body and that is flexibly deformed in accordance with the rotation of the oscillating body. Flexibility; a first internal gear having a rigidity engaged in the external gear; and a second internal gear 'axially disposed with the first internal gear and having a rigidity engaged with the external gear, wherein When the external gear is a cylindrical pin or a cylindrical pin, or when the internal gear of the first internal gear or the second internal gear is a cylindrical pin or a cylinder a straight line passing through the rotating shaft of the oscillating body and the center of the occlusion radius of the external gear, that is, the eccentric axis, and the first gear and the first one when the pin is inserted through the rotating body and the first internal gear or the second internal gear The intersection of the respective common normals of the contact points generated by the engagement of the internal gear and the second internal gear, that is, the center of the pin is disposed between the pitch points. When the internal gear of the internal gear of the table 1 or the second internal gear is assumed to be a cylindrical pin, specifically, the external tooth is obtained based on the imaginary pin, and the ith internal gear and the second internal gear are formed according to the external tooth. The internal tooth acts as an envelope. In the present invention, when the external gear is a cylindrical pin, or when the internal gear of the first internal gear or the second internal gear is a cylindrical pin, the two pitch points are Configure the pin center. Therefore, the load applied to the external teeth of the cylindrical external gear when engaged with the internal gear 201231842 and the load applied to the external external teeth of the cylindrical external gear when the second internal tooth is engaged have the opposite directions and can be in the external gear The area applied to the two loads of the external gear is brought closer in the outer circumferential direction. That is, when viewed from the axial direction, the two internal gears can be set to a state in which only a small number of external teeth are sandwiched. Therefore, it is possible to prevent the engagement of the external gear and the internal gear from being shifted due to excessive torque (ratchet phenomenon). That is, the present invention pays particular attention to the case where the ratchet height is high, the allowable transmission torque can be increased, and the efficiency can be increased. According to the present invention, the impact resistance can be improved and the transmission torque transfer efficiency can be increased. [Embodiment] Hereinafter, a first embodiment of the present invention will be described in detail with reference to the accompanying drawings. <First Embodiment> <Configuration> The structure of the present embodiment will be briefly described mainly with reference to Figs. 1 and 2 . The flexible snap-in gear device 100 includes an oscillating body 104: wheels 120A and 120B (only as the external gear 120), and is disposed outside the oscillating cymbal and has flexibility to be flexibly deformed according to the rotation of the oscillating body 104. The internal gear 13 0A for deceleration of the internal gear and the internal gear 130B for output of the second gear have the combination of the biting of the external gear 120, and the present invention is particularly advantageous. Overall external teeth 1 04 Sex: Internal geared 201231842 Rigidity. Further, the internal gear 1 3 Ο A for deceleration and the internal gear 1 3 Ο B for output will be collectively referred to simply as the internal gear 13 3 以后. The following is a detailed description of each component. As shown in Figs. 3(A) and 3(B), the oscillating body 1〇4 has a columnar shape. An input shaft hole 106 into which an input shaft (not shown) is inserted is formed at the center. A keyway 108 is provided in the input shaft hole 1〇6 such that the oscillator body 104 rotates integrally with the input shaft when the input shaft is inserted and rotated. As shown in Fig. 3 and Fig. 4, the oscillating body 104 is formed in a shape (double arc shape) in which two circular arc portions (the first circular arc portion FA and the second circular arc portion SA) are connected together. The first circular arc portion FA is an arc having a curvature radius r 1 centered on a point B (referred to as an eccentric shaft), and constitutes a circular arc portion (also referred to as a nip range) for engaging the external gear 120 and the internal gear 130. ). The second circular arc portion S A is an arc of a curvature radius r 2 centering on the point C, and constitutes a circular arc portion (also referred to as a non-engagement range) of a range in which the external gear 120 and the internal gear 130 do not engage each other. The length of the first circular arc portion FA is determined by the occlusion angle Θ which is the angle formed by the long axis X and the normal N at the point A. At this time, as shown in Fig. 4, if the radius of the major axis X of the vibrating body 104 is r ', the eccentric amount is L, and the radius of curvature r of the first circular arc portion FA is expressed by the formula (1). 1 ° rl = r - L (1) Further, as shown in Fig. 4, the tangent line τ (normal line N) is common to the connection portion A between the first circular arc portion fa and the second circular arc portion SA. Therefore, the radius of curvature Γ2 of the second arc portion SA is (curvature radius ri + length BC), so 201231842 is expressed by the formula (2). r 2 = r 1 + length BC = r 1 + L / cos9 ( 2 ) The oscillating body bearing 110A is a bearing disposed between the outer side of the oscillating body 丨 04 and the inner side of the outer gear 12 0 A, as shown in Fig. 2 As shown in Fig. 5, the inner ring 112 and the retainer 114A' are used as the rotating body of the roller 116A and the outer ring 118A. The inner side of the inner ring 112 abuts against the oscillating body 104, and the inner cymbal 112 rotates while the body 104 is deformed. The roller ι16Α is cylindrical (including the needle roller). Therefore, compared with the case where the rotating body is a ball, the portion where the roller 116A is in contact with the inner bore 112 and the outer ring 118A is increased, so that the load capacity can be increased. That is, by using the roller 1 16A, the transmission torque of the oscillating body bearing 1 10A can be increased, and the life can be extended. The outer ring 118A is disposed on the outer side of the roller M6A. The outer ring 11 8A is flexibly deformed in accordance with the rotation of the vibrating body 104, and the outer gear 1 2 0 A disposed on the outer side thereof is deformed. Further, as shown in Fig. 2, the oscillating body bearing U0B is the same as the oscillating body bearing 110A, and is composed of an inner ring 112, a retainer 114B, a roller 116B, and an outer ring 11 8B. The oscillating body 104 and the inner ring 112 are common to the oscillating body bearings 110A, 110B. Further, the retainer 114B, the roller 116B, and the outer ring 118B are used as a single member (assembly), and are identical to the retainer 114A 'roller 116A and outer ring 118A.>> As shown in Fig. 2, the external gear 120A and the internal gear for deceleration Engage in 130A. The outer gear 120A is composed of a base member 122 and outer teeth 124A. The base member 122 is a cylindrical member that supports the external teeth 124A and has flexibility, and is disposed on the outer side of the vibrating body bearing ΠA. The external tooth 1 24A becomes a cylindrical pin of -10-201231842 radius Ρ1 (so 'the outer tooth (124Β) or the outer gear 12〇a (120Β) or the flexible bite gear 1〇〇 is also called only The pin type) outer teeth 124 are secured by the ring member 126 to the base member 122. As shown in Fig. 2, the external gear 120 is engaged with the inner gear for output. Further, the outer gear 120 is the same as the outer gear 120, and the base member 122 and the outer teeth 124 are formed. The number of external teeth 124 is the same as the teeth 124 and is composed of the same cylindrical pin and is held by the ring 126B on the base member 122. That is, the base member 122 supports the outer teeth 124A and the outer teeth 124B. That is, the external gear 120A has a tooth shape of the same shape. The eccentric amount L of the oscillating body 104 is the same phase outer teeth 124A and outer teeth 124B. Thereafter, the external teeth 124A, 124B are external teeth 124. As shown in Fig. 2, the internal gear for deceleration 1 300 A is formed of a rigid member. The reduction internal gear 130A has a number of teeth that is only twice as large as the number of teeth of the external gear 120A 124A (the number of teeth will be described later). The outer casing of the reduction gear 1 3 Ο A is fixed to the casing (not shown) through the bolt holes. Further, the reduction internal gear 130A is engaged by the gear 120A, and contributes to the deceleration of the rotation of the oscillation body 1〇4. FIG. 7(A) shows the appearance of the external gear 120A and the reduction internal gear 130A, FIG. 7 ( A) indicates the appearance of the external teeth 124A and 1 2 8 A on the x-axis. On the other hand, the output internal gear 130B is also formed of a member having rigidity as in the deceleration inner 1130A. Output internal gear

1 24A 裝置持於 1 30B 由基與外構件共同 1 20B 傳至統稱的構外齒行詳 1 32A 與外。第咬合內齒齒輪 1 30B -11 - 201231842 具備與外齒輪120B的外齒124B之齒數相同的內齒128B 之齒數（等速傳遞）。再者’於輸出用內齒輪130B上，透過螺栓孔132B安裝未圖示之輸出軸’並且與外齒輪 120B之自轉相同的旋轉輸出至外部。第6圖（B)中表示外齒輪120B和輸出用內齒輪130B相咬合的樣子，第7 圖（B)中表示X軸上的外齒124B和內齒128B之樣子》以後，將內齒128A、128B統稱爲內齒128。在本實施形態中，將外齒輪1 20A與減速用內齒輪 130A之同時咬合數Nph和外齒輪120B與輸出用內齒輪 13 0B之同時咬合數Npl均設爲2以上並且將其咬合設爲理論咬合》因此，扭矩之傳遞效率不會變低，即可實現順暢的扭矩傳遞並可使傳遞扭矩增大。 <齒形決定方法> 對外齒輪120、減速用內齒輪130A及輸出用內齒輪 13 0B之齒形決定方法進行說明。首先，以下槪略說明齒形之求法。最開始定義外齒輪1 20之齒形。其次，由次擺線曲線公式表示外齒輪120之齒形軌跡，用其次擺線曲線公式定義內齒輪130之齒形。其次，由外齒輪120和內齒輪130 之大小和齒數，對定義外齒輪120和內齒輪130之齒形的多數個參數建立相互關聯。其次，決定內齒輪130的齒形之齒頂和齒根之修正範圍。其次，用已建立關聯之參數求出修正範圍以外的齒形部份，並由其齒形部份求出同時咬 -12- 201231842 合數。而且，以將同時咬合數均設爲2以上之方式決定最佳參數。在決定參數時，以同時滿足扭矩、齒面之容許面壓、各部位之主應力、軸承壽命等目標値之方式不斷摸索。以下，進行詳細說明。最開始定義外齒輪120之齒形。當將外齒1 24設爲半徑p 1之圓筒形銷時，將從偏心軸B至外齒輪120之咬合範圍中作爲外齒124之銷中心之位置（pl=0 )的距離R1稱爲外齒輪120之咬合範圍中的齒形實體半徑。又，當將內齒輪130之內齒128設爲半徑 p2之圓筒形銷時（包括僅在設計上假想之情況）’將從震盪體1〇4之旋轉軸Fc (軸向0上的點）至作爲（包括假想）內齒128之銷中心之位置（p2 = 0)的距離R稱爲內齒輪130之齒形實體半徑。如此’如第8圖所示’由式 (3)表示半徑R與半徑R1之關係。 R1=R-L …（3) 在本實施形態中’外齒輪120透過震盪體軸承110配置於震盪體104之外周。震盪體軸承110和外齒輪120之半徑方向之厚度均爲一定。因此’由於震邊體104爲雙圓弧形狀’故外齒輪1 2 0亦成爲雙圓弧形狀。相當於震邊體 104之咬合範圍之曲率半徑rl的外齒輪120之咬合範圍中的齒形實體半徑成爲R1。因此，若將相當於震盪體104 之非咬合範圍之曲率半徑r2的外齒輪120之非咬合範圍中的齒形實體半徑設爲R2，則可利用式（2 )、式 -13- 201231842 (3 )，由式（4)表示半徑R2。 R2 = R1 -L/cos0 ... ( 4 ) 如第9圖所示，外齒124成爲於咬合範 B位於半徑R1(=R-L)之圓周上的半徑（： (因此，偏心軸B成爲外齒輪1 20和內齒輪齒輪120的咬合半徑之中心）。因此，依半徑P1、偏心量L、半徑R 定義外齒輪120之齒形。其次，定義內齒輪130之齒形。求出外形實體位置（半徑pl =〇之位置）之軌跡之僅移動半徑pi者設爲內齒輪130之齒形。具體說明。再者，將外齒輪1 20爲齒形實體形齒輪（稱爲假想齒輪）時的減速比稱爲假如第1 0圖所示，使外齒輪120以震盪軸Fc爲中心公轉角度α。亦即，偏心軸時，外齒輪120之齒形實體位置的座標（X: 想減速比η向反方向僅自轉角度α/η而移動 y2)。因此，由式（5) 、（6)示出表示外形實體位置之軌跡的座標（xPfe，ypf(：)。 [數1] ζχ1 24A device held at 1 30B by the base and external components together 1 20B passed to the collective external gear details 1 32A and outside. The first engaging internal gear 1 30B -11 - 201231842 has the number of teeth (equal speed transmission) of the internal teeth 128B which are the same as the number of teeth of the external teeth 124B of the external gear 120B. Further, in the output internal gear 130B, an output shaft 'not shown (not shown) is attached through the bolt hole 132B, and the same rotation as that of the external gear 120B is outputted to the outside. Fig. 6(B) shows a state in which the external gear 120B and the output internal gear 130B are engaged, and Fig. 7(B) shows the appearance of the external teeth 124B and the internal teeth 128B on the X-axis. 128B is collectively referred to as internal teeth 128. In the present embodiment, the number Neng of the simultaneous engagement of the external gear 1 20A and the internal gear 130A for deceleration and the number Npl of the simultaneous engagement of the external gear 120B and the output internal gear 130B are both 2 or more and the occlusion is theoretical. Therefore, the torque transmission efficiency does not become low, and smooth torque transmission can be achieved and the transmission torque can be increased. <Tooth shape determination method> The tooth profile determination method of the external gear 120, the reduction internal gear 130A, and the output internal gear 130B will be described. First, the following is a brief description of the method of determining the tooth shape. The tooth profile of the external gear 1 20 is initially defined. Next, the tooth profile of the external gear 120 is represented by the trochoid curve formula, and the tooth profile of the internal gear 130 is defined by the second cycloid curve formula. Secondly, the majority of the parameters defining the tooth profile of the external gear 120 and the internal gear 130 are correlated by the size and the number of teeth of the external gear 120 and the internal gear 130. Next, the correction range of the tooth tip and the root of the tooth profile of the internal gear 130 is determined. Secondly, the toothed portion outside the correction range is obtained by the parameters of the established correlation, and the toothed portion is used to find the simultaneous bite -12-201231842. Further, the optimum parameters are determined such that the number of simultaneous occlusions is 2 or more. When determining the parameters, the target is continuously evaluated by satisfying the torque, the allowable surface pressure of the tooth surface, the principal stress of each part, and the bearing life. The details will be described below. The tooth profile of the external gear 120 is initially defined. When the external teeth 1 24 are set as the cylindrical pins of the radius p 1 , the distance R1 which is the position ( pl=0 ) of the pin center of the external teeth 124 from the eccentric shaft B to the external gear 120 is referred to as the distance R1. The radius of the toothed solid in the nip range of the outer gear 120. Further, when the internal teeth 128 of the internal gear 130 are set to the cylindrical pin of the radius p2 (including the case of imaginary design only) 'the rotation axis Fc from the oscillation body 1〇4 (the point on the axial direction 0) The distance R to the position (p2 = 0) of the pin center of the inner tooth 128 (including the imaginary) is referred to as the toothed solid radius of the internal gear 130. Thus, as shown in Fig. 8, the relationship between the radius R and the radius R1 is represented by the formula (3). R1 = R - L (3) In the present embodiment, the external gear 120 is disposed in the outer periphery of the oscillating body 104 through the oscillating body bearing 110. The thickness of the oscillating body bearing 110 and the external gear 120 in the radial direction is constant. Therefore, since the base body 104 has a double circular arc shape, the outer gear 1 200 also has a double arc shape. The radius of the toothed solid in the nip portion of the external gear 120 corresponding to the radius of curvature rl of the occlusion range of the rim body 104 becomes R1. Therefore, if the radius of the toothed solid in the non-engaging range of the external gear 120 corresponding to the radius of curvature r2 of the non-engaging range of the vibrating body 104 is R2, the formula (2), the formula-13-201231842 (3) can be utilized. ), the radius R2 is represented by the formula (4). R2 = R1 - L / cos0 (4) As shown in Fig. 9, the external teeth 124 become a radius on the circumference of the radius R1 (= RL) of the occlusion unit B (: (therefore, the eccentric axis B becomes external) The center of the nip radius of the gear 120 and the internal gear 120. Therefore, the tooth profile of the external gear 120 is defined by the radius P1, the eccentric amount L, and the radius R. Next, the tooth profile of the internal gear 130 is defined. The movement radius pi of the trajectory (the position of the radius pl = 〇) is set to the tooth profile of the internal gear 130. Further, when the external gear 1 20 is a tooth-shaped solid gear (referred to as a virtual gear) The reduction ratio is referred to as the first embodiment shown in FIG. 10, and the external gear 120 is rotated by the angle α around the oscillation axis Fc. That is, when the eccentric shaft is used, the coordinate of the tooth-shaped solid position of the external gear 120 (X: the reduction ratio η) In the opposite direction, only the rotation angle α/η is moved by y2). Therefore, the coordinates (xPfe, ypf(:). [number 1] 表示 indicating the trajectory of the physical position of the shape are shown by the equations (5) and (6).

Xpfc = (Λ - L) * cos — + L * cos α … η ct yPfc = {R-L)* sin--Z- * sin a .·· n 在此，如第1 1圖所示，就內齒輪13 0 圍中距偏心軸 1之圓筒形銷 1 3 0咬合時外及咬合角度 Θ 齒輪1 20之齒後，將向內側以下，進一步 [半徑R1之圓想減速比η » 體104之旋轉 Β旋轉a。此，yi )依據假至座標（X 2，齒輪1 2 0之齒 (5) (6) 之齒形而言， •14- 201231842 由於與外齒輪120進行理論咬合，因此內齒輪130之齒形實體位置的座標由內次擺線曲線公式（圓內次擺線曲線公式）表示。亦即，若利用以旋轉軸F c爲中心固定的基圓 BA之半徑bl、沿基圓BA之圓周不滑動地旋轉的旋轉圓 A A之半徑a 1、描繪點之半徑L1及旋轉角β 1，則由式 (7)、式（8)表示內齒輪130之齒形實體位置的座標 (Xpfc ’ ypfc)。 [數2] …⑺ 工妙=(办1-al)*cos/3 l+Ll*cos —~—β 1Xpfc = (Λ - L) * cos — + L * cos α ... η ct yPfc = {RL)* sin--Z- * sin a .·· n Here, as shown in Fig. 1, the internal gear 13 0 Cylindrical pin 1 in the middle of the eccentric shaft 1 3 0 Outside the occlusal and the occlusal angle 后 After the tooth of the gear 1 20, it will be below the inside, and further [Ring R1 circle wants to reduce the ratio η » The rotation of the body 104 Β Rotate a. Thus, yi) according to the coordinates of the coordinates (X 2, the teeth (5) (6) of the gears 120, • 14-201231842 due to the theoretical engagement with the external gear 120, the tooth shape of the internal gear 130 The coordinates of the solid position are represented by the inner trochoid curve formula (the inner trochoid curve formula), that is, if the radius bl of the base circle BA fixed around the rotation axis F c is used, the circumference along the base circle BA is not The radius a 1 of the rotational circle AA that slidably rotates, the radius L1 of the drawing point, and the rotation angle β1, the coordinates of the tooth-shaped solid position of the internal gear 130 (Xpfc ' ypfc) are expressed by the equations (7) and (8). [Number 2] ...(7) Miao Miao = (do 1-al) *cos/3 l+Ll*cos —~—β 1

V yp/c = (b\-a\)*sm^ 1-Ll*sin ΟΙ …（8) 在此，若利用式（9 )〜（1 1)之關係，則即可求出式 (1 2 )、式（1 3 )之關係。 [數3] al=丄... (9) η bl^-(R-L) …（10) η β 1 二 /3、乙 1 =L - -(11)V yp/c = (b\-a\)*sm^ 1-Ll*sin ΟΙ (8) Here, if the relationship of equations (9) to (1 1) is used, the equation (1) can be found. 2), the relationship of formula (1 3 ). [Equation 3] al=丄... (9) η bl^-(R-L) (10) η β 1 2 /3, B 1 =L - -(11)

Xp/c = {R — L)* cos β +L* cos(n * j3) ...(12) yPfv = (R -L)*s'mβ~ L*sin(«* j3 ) ..· (1 3) 再者’由於式（5)和式（12)(式（6)和式 (13))表不相问座標（Xpfc，ypfc) ’故可求出式 -15- 201231842 (14)。 α = η * β ".(14) 其次，如第1 2圖所示，藉由使內齒輪1 3 0之齒形實體位置的座標（xpf(：，ypfc)向內側（內齒輪130側）僅移動外齒124之半徑pi，可以由式（丨5)〜（I7)表示內齒輪130之齒形的座標（Xf。，yfe)。 [數4] x/c = (R-L)* cos j3 +L * cos(« * j3) + pi * cos 77 …（1 5) y/c = (R-L)* sinj3 ~L * sin(w * β) + ρ\* sin η …（1 6) ▲ (Λ — Z.) * sin β +L* n* sin(« * /3 ) tan 77 =—------ (Λ — Z,) * cos β ~L*n* cos(n *0 ) -· (1 7) 亦即，代入半徑R、P1、偏心量L及假想減速比η(用以製作減速用內齒輪1 3 〇 Α之齒形的假想減速比nh、用以製作輸出用內齒輪130B之齒形的假想減速比ηι)並改變角度β，從而可以求出減速用內齒輪13 0A和輸出用內齒輪1 30Β之齒形各自的座標（Xf。、yf。）。其次，對規定外齒輪12〇和內齒輪13〇的參數建立相互關聯。如上述，外齒輪120之形狀爲由半徑Ri、R2規定的雙圓弧形狀。因此，利用表示外齒輪120A與減速用內齒輪130 A之齒數差的參數k(2以上）及用以導出減速比N 的參數i(減速用內齒輪130A時，i = 1、輸出用內齒輪 130B時，i = 0)，可以將表示第13圖所示的外齒輪120、內齒輪130各自的大小（由齒形®體半徑R、R1求出的 -16- 201231842 周長LC (圓周之長度）和利用假想齒輪之假想減速比η 時的節距P(1個齒之周期的外周方向長度））以及齒數 NT示於表中。其中，由於基於假想齒輪之節距P與基於外齒輪120之節距（=LC/NT)相等，故存在式（18)之關係。 NT = LC/P ... ( 18) 若利用式（18 )，則可從第13圖之表中導出式 (19)、式（20 )。 [數5] (N + iYk = ~^· …（19)Xp/c = {R — L)* cos β +L* cos(n * j3) (12) yPfv = (R -L)*s'mβ~ L*sin(«* j3 ) ..· (1 3) Furthermore, since equations (5) and (12) (formula (6) and (13)) do not match the coordinates (Xpfc, ypfc) ', equation -15-201231842 can be found. ). α = η * β " (14) Next, as shown in Fig. 12, the coordinates (xpf (:, ypfc) of the toothed solid position of the internal gear 1 300 are inward (the internal gear 130 side) Only the radius pi of the external teeth 124 is moved, and the coordinates (Xf., yfe) of the tooth shape of the internal gear 130 can be expressed by the equations (丨5) to (I7). [Number 4] x/c = (RL)* cos J3 +L * cos(« * j3) + pi * cos 77 (1 5) y/c = (RL)* sinj3 ~L * sin(w * β) + ρ\* sin η (1 6) ▲ (Λ — Z.) * sin β +L* n* sin(« * /3 ) tan 77 =—------ (Λ — Z,) * cos β ~L*n* cos(n *0 - (1 7) In other words, the radius R, P1, the eccentric amount L, and the virtual reduction ratio η (the virtual reduction ratio nh for forming the tooth profile of the internal gear 13 3 for deceleration) are used for the output. The virtual reduction ratio ηι) of the tooth profile of the internal gear 130B changes the angle β, and the coordinates (Xf, yf.) of the tooth profiles of the reduction internal gear 130A and the output internal gear 1 30Β can be obtained. The parameters defining the external gear 12 〇 and the internal gear 13 建立 are related to each other. As described above, the shape of the external gear 120 is defined by the radii Ri, R2. Therefore, the parameter k (2 or more) indicating the difference in the number of teeth between the external gear 120A and the internal gear 130A for deceleration and the parameter i for deriving the reduction ratio N (the internal gear 130A for deceleration, i = 1) In the case of the output internal gear 130B, i = 0), the size of each of the external gear 120 and the internal gear 130 shown in Fig. 13 can be obtained (-16 - 201231842 weeks obtained from the tooth profile body radius R, R1) The length P (the length of the circumference) and the pitch P (the length in the outer circumferential direction of the period of one tooth) when the virtual reduction ratio η of the imaginary gear is used, and the number of teeth NT are shown in the table, in which the pitch is based on the imaginary gear. P is equal to the pitch based on the external gear 120 (=LC/NT), so there is a relationship of the formula (18). NT = LC/P ... (18) If the formula (18) is used, it can be seen from the 13th figure. In the table, equations (19) and (20) are derived. [5] (N + iYk = ~^· ... (19)

2n{R-L) + AL^- …（20) N*k = sin(^ - θ) 2n{R-L) η 其次，導入參數Gp (稱爲銷型節距係數）。其中，將穿過偏心軸B和旋轉軸Fc的直線與由外齒輪120(之外齒124)和內齒輪130(之內齒128)咬合產生的接觸點之共同法線之交點稱爲基於外齒輪120和內齒輪130之節點。銷型節距係數Gp係爲了能夠容易掌握外齒輪1 2〇和內齒輪1 3 0各自的齒形實體位置與節距點的相對位置關係且能夠容易調整這些參數彼此而導入者。具體而言，如式（21)所示，銷型節距係數Gp由半徑R1( = R-L)與從偏心軸B至基於外齒輪1 2 0和內齒輪1 3 0之節距點的距離 -17- 201231842 η * L之比表示。 [數6]2n{R-L) + AL^- (20) N*k = sin(^ - θ) 2n{R-L) η Next, the parameter Gp (referred to as the pin-type pitch coefficient) is introduced. Wherein, the intersection of the straight line passing through the eccentric shaft B and the rotation axis Fc with the common normal of the contact points generated by the engagement of the external gear 120 (outer teeth 124) and the internal gear 130 (the internal teeth 128) is referred to as external A node of the gear 120 and the internal gear 130. The pin type pitch coefficient Gp is capable of easily grasping the relative positional relationship between the tooth shape physical position of each of the external gear 1 2 〇 and the internal gear 1 30 0 and the pitch point, and can easily adjust these parameters to each other and introduce them. Specifically, as shown in the formula (21), the pin-type pitch coefficient Gp is determined by the radius R1 (= RL) from the eccentric axis B to the pitch point based on the external gear 1 2 0 and the internal gear 1 3 0 - 17- 201231842 η * L ratio representation. [Number 6]

Gp =Gp =

n* L R-L …（2 1) m = N-(-'2Gph …（23) 當點Ph表示基於外齒輪12 0A和減速用內齒輪之節點時，在第14圖中表示外齒輪120之齒形實體 (R-L)與假想減速比nh之關係。依據式（21 )，將得到的銷型節距係數Gph (稱爲銷型減速側節距係數義成式（22)。若設爲參數i = l，並由式（19) (20 )整理式（22 )，則得出式（23 )。 [數7]n* L RL ... (2 1) m = N-(-'2Gph (23) When the point Ph represents a node based on the external gear 120A and the internal gear for deceleration, the tooth of the external gear 120 is shown in FIG. The relationship between the shape entity (RL) and the imaginary reduction ratio nh. According to the formula (21), the pin type pitch coefficient Gph (referred to as the pin type deceleration side pitch coefficient formula (22) is obtained. If it is set to the parameter i = l, and formula (22) is obtained by equation (19) (20), then equation (23) is obtained. [7]

Gph = ^i …（2 2)Gph = ^i ...(2 2)

R-L --Θ 2 sin(—-^) 當點Pi表示基於外齒輪120B和輸出用內齒輪之節點時，在第15圖中表示外齒輪120之齒形實體 (R-L )與假想減速比ηι之關係。依據式（2 1 )，將得到的銷型節距係數Gpl (稱爲銷型輸出側節距係數義成式（24 )。若設爲參數i = 0，並由式（19 ) (20 )整理式（24 )，則得出式（25 ) » 1 30A 半徑此時 )定、式 1 30B 半徑此時 )定、式 •18- 201231842 [數8] °ΡΙ = !^ …（2 4) Λ —厶 j- …（25) sin(|-^) 因此，若給出半徑R、減速比N、銷型減速側節距係數Gph及咬合角度Θ，則可以決定假想減速比nh、偏心量 L，接著可以求出銷型輸出側節距係數Gpl、假想減速比 η】〇本實施形態中，如第1 4圖、第1 5圖所示，代入銷型減速側節距係數Gph<l，求出銷型輸出側節距係數GP1>1 之値。本實施形態中，從求出各齒形之結果考慮，咬合角度Θ爲40〜65度且銷型減速側節距係數Gph之cos·1之値爲1 5〜3 0度之情況爲進一步更佳的條件。其次，決定內齒輪130之齒形的修正範圍。如第16圖所示，將連接內齒128之座標和外齒124 (銷）之中心0 c的直線與X軸所成的角度β約爲45度時的角度設爲β s。如此一來，角度β在從零至β s之間，有可能存在與外齒輪120之外齒124之干擾，故在其範圍內對內齒輪130之內齒128之齒根進行修正。又，將外齒 124之齒頂與內齒128之齒頂的距離δ爲銷的半徑pi之約15%時的角度β設爲pf。角度β在pf至π之間，有可能存在與外齒輪120之外齒124之干擾以及在與外齒輪 -19- 201231842 120之外齒124之咬合時成爲高面壓，因此在其範圍內對內齒輪130之內齒128之齒頂進行修正。亦即，未進行齒形修正的角度Ps〜Pf (未修正之齒形區域）成爲進行理論咬合的有效範圍。其次，求出同時咬合數Nph、Npl。同時咬合數Nph、Npl可藉由以外齒輪120之旋轉角度α所決定的有效範圍除以節角（271除以齒數NT之値）求出。其中，角度Pfh、Psh爲減速用內齒輪130A中的角度，角度Pfl、psl設爲輸出用內齒輪130B中的角度。從式（14 )之關係來看，由角度pfh、psh、βίΐ、psl求出的旋轉角度分別爲 afh、ash、afl、asl。亦即，藉由利用式 (1〇，可分別由式（26)、式（27)求出減速用內齒輪 130A之同時咬合數Nph、輸出用內齒輪130B之同時咬合數 N p 1 〇 [數9]RL - Θ 2 sin (--^) When the point Pi represents the node based on the external gear 120B and the output internal gear, the tooth profile (RL) of the external gear 120 and the imaginary reduction ratio ηι are shown in Fig. 15. relationship. According to the formula (2 1 ), the obtained pin-type pitch coefficient Gpl (referred to as the pin-type output side pitch coefficient formula (24). If the parameter is i = 0, and is sorted by the formula (19) (20) Equation (24), then the formula (25) » 1 30A radius at this time), formula 1 30B radius at this time), formula • 18- 201231842 [number 8] °ΡΙ = !^ ... (2 4) Λ —厶j- ...(25) sin(|-^) Therefore, if the radius R, the reduction ratio N, the pin type deceleration side pitch coefficient Gph, and the nip angle Θ are given, the imaginary reduction ratio nh and the eccentric amount L can be determined. Then, the pin type output side pitch coefficient Gpl and the virtual reduction ratio η can be obtained. In the present embodiment, as shown in Figs. 4 and 5, the pin type deceleration side pitch coefficient Gph < l is substituted. Find the pin type output side pitch coefficient GP1 >1. In the present embodiment, from the result of obtaining each tooth shape, the occlusion angle Θ is 40 to 65 degrees, and the case where the pin type deceleration side pitch coefficient Gph is 1 to 3 0 degrees is further increased. Good condition. Next, the correction range of the tooth profile of the internal gear 130 is determined. As shown in Fig. 16, the angle at which the angle β between the line connecting the coordinates of the internal teeth 128 and the center 0 c of the external teeth 124 (pin) and the X-axis is about 45 degrees is β s . As a result, the angle β is between zero and β s, and there is a possibility of interference with the external teeth 124 of the external gear 120, so that the root of the internal teeth 128 of the internal gear 130 is corrected within the range. Further, the angle β when the distance δ between the addendum of the external tooth 124 and the addendum of the internal tooth 128 is about 15% of the radius pi of the pin is pf. The angle β is between pf and π, there is a possibility of interference with the external teeth 124 of the external gear 120 and a high surface pressure when meshing with the external teeth 124 of the external gear -19-201231842 120, and therefore within the range thereof The crests of the internal teeth 128 of the internal gear 130 are corrected. That is, the angles Ps to Pf (uncorrected tooth-shaped regions) where the tooth shape correction is not performed become an effective range for performing theoretical occlusion. Next, the number of simultaneous occlusions Nph and Npl is obtained. At the same time, the number of nips Nph and Npl can be obtained by dividing the effective range determined by the rotation angle α of the external gear 120 by the pitch angle (271 divided by the number of teeth NT). Here, the angles Pfh and Psh are the angles in the internal gear 130A for deceleration, and the angles Pfl and ps1 are the angles in the internal gear 130B for output. From the relationship of the equation (14), the rotation angles obtained from the angles pfh, psh, βίΐ, and psl are afh, ash, afl, and asl, respectively. In other words, by using the equation (1), the simultaneous occlusion number Nph of the reduction internal gear 130A and the simultaneous engagement number N p 1 〇 of the output internal gear 130B can be obtained from the equations (26) and (27), respectively. Number 9]

Nph = a fh~ a sh 2π =m β fh-* β sh ~2π (N + \)*k (N + \)*k (26)Nph = a fh~ a sh 2π =m β fh-* β sh ~2π (N + \)*k (N + \)*k (26)

Npl = a fi~ a si 2π β fi_ j3 si m —2π— (2 7) N*k N*k 沿式（26)、式（27)求出同時咬合數。此時，將設爲k = 2時求出的減速用內齒輪130A之同時咬合數Nph、輸出用內齒輪130B之同時咬合數Npl分別示於第17圖、第1 8圖中。依該等同時咬合數_ Nph、同時咬合數Npl均實現2以 -20- 201231842 上的直徑（2 * R )和減速比（1 /N )之條件’決定本實施形態中的內齒輪130之齒形。亦即’當齒數差爲2時 (K = 2)，減速比（1/N)爲1/20而並不會成爲本實施形態之齒形，而由1 / 3 〇以下（比1 / 3 0更大地減速的減速比）決定本實施形態之內齒輪130之齒形。 <動作> 主要利用第2圖對撓性咬合式齒輪裝置100之動作進行說明。若震盪體104依未圖示的輸入軸之旋轉而旋轉，則按照其旋轉狀態，外齒輪120 Α透過震盪體軸承110Α而撓性變形。再者，此時，外齒輪120B亦透過震盪體軸承 1 1 0B以和外齒輪1 2〇A相同的相位撓性變形。外齒輪1 20之撓性變形按照震盪體1 04之曲率半徑 rl之形狀進行。於第4圖所示之震盪體1〇4之第1圓弧部FA部份中的位置，由於曲率一定，故撓性應力成一定。於第1圓弧部FA與第2圓弧部SA之連接部份A中的位*置，由於切線T相同，故防止連接部份處之急劇的撓性變形。同時，於連接部份 A，由於沒有滾子1 16 A、 1 16B之急劇的位置變動，故滾子1 16A、1 16B之滑動少，且扭矩之傳遞損失少。藉由外齒輪1 2 0由震盪體〗04的撓性變形，外齒1 2 4 於第1圓弧部FA (咬合範圍）部份向半徑方向外側移動，從而咬合於內齒輪130之內齒128。進行咬合時，由 -21 - 201231842 於外齒124爲可旋轉的銷，故於咬合面，外齒124進行近似滾動之運動，而於面壓低於咬合面之基礎構件122側，外齒124進行滑動。因此，傳遞效率之損失少。又，內齒 128之齒形相對於作爲圓筒形銷之外齒124，成爲依據次擺線曲線之齒形。因此，外齒124和內齒128爲完全的理論咬合，故可減少損失，實現高扭矩傳遞效率》進行咬合時，對外齒124A施加與外齒124B不同的負載（方向和大小）（不同於本實施形態的外齒輪1 20，參照第29圖）。但是，震盪體軸承110A、110B除內圈 112之外，在軸向Ο上分離爲相對於與減速用內齒輪 130A咬合的外齒124A的部份和相對於與輸出用內齒輪 13 0B咬合的外齒124B的部份。因此，可分別防止以減速用內齒輪130A與外齒124A之咬合爲原因的滾子116B之偏斜及以輸出用內齒輪130B與外齒124B之咬合爲原因的滾子1 16A之偏斜。又，由於滾子116A、116B爲圓筒形，故相對於具備相同大小的球的球軸承，耐負載大且與內圈112及外圈 118A、118B接觸的部份多，故可加大負荷容量。再者，外齒1 24係在軸向0上分割成減速用內齒輪 130A咬合的部份（外齒124A)和輸出用內齒輪130B咬合的部份（外齒124B)者。因此，當外齒輪12 0A和減速用內齒輪13 0A咬合時，即使在外齒124B上有變形等，亦不會因其變形在外齒1 24A上產生變形。同樣地，當外齒輪120B和輸出用內齒輪130B咬合時，即使在外齒 -22- 201231842 124A上有變形等，亦不會因其變形在外齒124B上產生變形。亦即，藉由分割外齒124，能夠防止因一方的外齒 124A(124B)之變形使另一方的外齒124B(124A)變形而使其其咬合關係惡化之類的傳.遞扭矩之下降。Npl = a fi~ a si 2π β fi_ j3 si m — 2π — (2 7) N*k N*k The number of simultaneous occlusions is obtained along the equations (26) and (27). In this case, the simultaneous engagement number Nph of the reduction internal gear 130A and the simultaneous engagement number Npl of the output internal gear 130B obtained when k = 2 are shown in Fig. 17 and Fig. 18, respectively. According to the simultaneous occlusion number _ Nph and the occlusion number Npl, the condition of the diameter (2 * R ) and the reduction ratio (1 /N) on -20-201231842 is determined to determine the internal gear 130 in the present embodiment. Tooth shape. That is, when the difference in the number of teeth is 2 (K = 2), the reduction ratio (1/N) is 1/20, and it does not become the tooth shape of the present embodiment, but is less than 1 / 3 ( (ratio than 1 / 3). 0 Reduction ratio of greater deceleration) The tooth profile of the internal gear 130 of this embodiment is determined. <Operation> The operation of the flexible snap gear device 100 will be mainly described with reference to Fig. 2 . When the vibrating body 104 is rotated by the rotation of the input shaft (not shown), the external gear 120 挠 is flexibly deformed by the vibrating body bearing 110Α in accordance with the rotation state. Further, at this time, the external gear 120B is also transmitted through the vibrating body bearing 1 10 B to be flexibly deformed in the same phase as the external gear 1 2A. The flexible deformation of the outer gear 110 is performed in accordance with the shape of the radius of curvature rl of the vibrating body 104. The position in the first arc portion FA portion of the oscillating body 1 〇 4 shown in Fig. 4 has a constant curvature, so that the flexibility stress is constant. The bit * in the connection portion A between the first circular arc portion FA and the second circular arc portion SA is set to be the same, so that the tangent T is the same, so that the sharp deformation at the joint portion is prevented. At the same time, in the connecting portion A, since there is no sharp positional change of the rollers 1 16 A and 1 16B, the sliding of the rollers 1 16A and 1 16B is small, and the torque transmission loss is small. The external gear 1 2 4 is moved outward in the radial direction from the first arc portion FA (the nip portion) by the flexible deformation of the external gear 1 20 from the oscillating body 04, thereby engaging the internal teeth of the internal gear 130. 128. When the occlusion is performed, the outer tooth 124 is a rotatable pin from -21 to 201231842. Therefore, the outer tooth 124 performs an approximately rolling motion on the occlusal surface, and the outer tooth 124 is performed on the side of the base member 122 whose surface pressure is lower than the occlusal surface. slide. Therefore, the loss of transmission efficiency is small. Further, the tooth shape of the internal teeth 128 is a tooth shape based on the minor cycloid curve with respect to the teeth 124 as the cylindrical pin. Therefore, the external teeth 124 and the internal teeth 128 are completely theoretically engaged, so that the loss can be reduced and the high torque transmission efficiency is achieved. When the occlusion is performed, the external teeth 124A are applied with different loads (direction and size) from the external teeth 124B (different from this). The external gear 1 20 of the embodiment is referred to Fig. 29). However, in addition to the inner ring 112, the slewing body bearings 110A, 110B are separated in the axial direction by a portion that is engaged with the external teeth 124A that mesh with the internal gear 130A for deceleration, and with respect to the portion that is engaged with the internal gear 13BB for output. Part of the outer teeth 124B. Therefore, the deflection of the roller 116B due to the engagement of the reduction internal gear 130A and the external teeth 124A and the deflection of the roller 1 16A due to the engagement of the output internal gear 130B and the external teeth 124B can be prevented, respectively. Further, since the rollers 116A and 116B have a cylindrical shape, the ball bearing having the same size of the ball has a large load resistance and a large number of contacts with the inner ring 112 and the outer rings 118A and 118B, so that the load can be increased. capacity. Further, the external teeth 1 24 are divided into a portion where the decelerating internal gear 130A is engaged (the external teeth 124A) and a portion where the output internal gear 130B is engaged (the external teeth 124B) in the axial direction 0. Therefore, when the external gear 120A and the reduction internal gear 130A are engaged, even if there is deformation or the like on the external teeth 124B, deformation of the external teeth 1 24A is not caused by the deformation. Similarly, when the external gear 120B and the output internal gear 130B are engaged, even if there is deformation or the like on the external teeth -22-201231842 124A, deformation is not caused on the external teeth 124B due to the deformation thereof. In other words, by dividing the external teeth 124, it is possible to prevent the external teeth 124B (124A) from being deformed by deformation of one of the external teeth 124A (124B), thereby deteriorating the transmission torque. .

外齒輪120A和減速用內齒輪130A之咬合位置隨著震盪體104之長軸方向X之移動而旋轉移動。在此，若震盪體1 04旋轉1圏，則外齒輪1 20A之旋轉相位僅延遲與減速用內齒輪130A之齒數差。亦即，基於減速用內齒輪 13 0A之減速比可設爲（（外齒輪120A之齒數（N*k)-減速用內齒輪130A之齒數（（N + l) *k) ) /外齒輪120A 之齒數（N*k ) ) =-1/Ν )而求出。外齒輪120B和輸出用內齒輪130B之齒數（N*k)均相同，故外齒輪120B和輸出用內齒輪130B的相互咬合的部份不會移動，由相同的齒彼此相咬合。因此，從輸出用內齒輪130B輸出與外齒輪120B之自轉相同之旋轉。其結果，可從輸出用內齒輪13 0B取出依據基於減速用內齒輪13 0A之減速比1/N減速震盪體104之旋轉的輸出。本實施形態中，作爲其基本結構，具備使筒形外齒輪 120咬合於2個具有剛性的內齒輪130 (減速用內齒輪 130A和輸出用內齒輪130B )之結構，並且以使外齒輪 120和內齒輪130具備將外齒輪120與內齒輪130之同時咬合數Nph、Npl均設爲2以上之齒形之方式構成，再者，藉由利用次擺線曲線來實現理論咬合。因此，能夠提高耐衝擊性，分散施加於咬合齒面之面壓，並傳遞大扭 -23- 201231842 矩，尤其是和以往通常的杯形撓性咬合式齒輪裝置相比，能夠格外減少外齒輪1 2 0中產生之局部應力。亦即，在本實施形態之撓性咬合式齒輪裝置中，在不會因震盪體之撓性而產生圓錐形變形，且亦不會有杯底部之應力集中之狀態下，能夠謀求咬合面積之增大和面壓之分散，故可較大地增加負荷容量。又，在本實施形態中，如第Μ圖、第15圖、第19 圖所示，設爲銷型減速側節距係數Gph< 1、銷型輸出側節距係數 Gpl>l，故式（28 )成立。亦即，如式（29 )所示，在本實施形態中，從偏心軸B，外齒輪1 2 0之銷中心 (齒形實體）之位置配置於從偏心軸B至基於外齒輪 120A和減速用內齒輪13 0A之節點Ph的距離（nh*L)與從偏心軸B至基於外齒輪12 0B和輸出用內齒輪130B之節點P!的距離（n，L)之間》 [數 10] m*L . m*L . -< 1，->1 …（28) R-L R-L m* L<R-L<ni*L …（2 9) 因此，與減速用內齒輪130A咬合時施加於外齒輪 120A之外齒124A的負載和與輸出用內齒輪130B咬合時施加於外齒輪120B之外齒124B的負載具備相互反方向之成份，並且可以使施加於外齒輪120之該2個負載之區域於外齒輪120之外周方向上靠近。亦即，從軸向〇觀察時，進行咬合動作時，將2個內齒輪1 3 0能夠設爲僅夾入 -24- 201231842 少數外齒1 24之態樣。因此，可防止外齒輪1 2〇與內齒輪 130之咬合因過度扭矩而偏移之現象（棘輪現象）。亦即，能夠提高耐棘輪性。在利用了實際上已產品化的杯形外齒輪的撓性咬合式齒輪裝置（內齒輪之齒形實體半徑約爲26mm且減速比爲 1/50者（稱爲比較例））和具備相同程度的大小和相同減速比的本實施形態之撓性咬合式齒輪裝置1 00中，可以確認出相對於比較例之實測値已大幅度（約4倍以上）改善了耐棘輪性。.同時，通過理論計算及試驗可確認，比較例中額定扭矩爲3 . 3 kgfm，相對於此，本實施形態之撓性咬合式齒輪裝置100中，額定扭矩爲6.6kg fm。亦即，在理論計算上以及通過試驗都可以確認，額定扭矩均爲約2 倍。如此，在本實施形態中，能夠增大傳遞扭矩，並且能夠增大傳遞效率。再者，代替傳遞扭矩之提高，亦可使撓性咬合式齒輪裝置100進一步緊湊化。又’在本實施形態中，外齒輪1 20之齒形在分別與減速用內齒輪130A及輸出用內齒輪130B咬合的部份設成相同’因此容易加工外齒輪1 20，可較低地抑制加工成本’並且可實現高精度的形狀加工。亦即’根據本發明，藉由增加外齒輪.1 2 0與內齒輪 130之同時咬合數Nph、Npl，能夠增大傳遞扭矩及傳遞效率。 -25- 201231842 《第2實施形態》利用第20圖至第29圖對本發明之第2實施例進行詳細說明。本實施形態係針對外齒輪採用線曲線之齒形來代替第1實施形態之圓筒形銷，輪之外齒與基礎構件一體成型者（稱爲實心型）若與第1實施形態中使用的參數相同地定義，則態中使用的參數的標記亦設爲相同。對與第1實施形態不同的結構和齒形決定方明，對於其他部份，在後兩位數附加相同標記而說明。 <結構> 如第20圖、第21圖所示，外齒輪220A與齒輪230A內咬合。外齒輪220A包括基礎構件齒224A。基礎構件222爲具有可撓性的筒形構於震盪體軸承210A之外側並與外齒224A —體此，能夠縮小外齒224A，並且能夠進行高精度亦即，本實施形態之外齒輪220A適合於負荷容型撓性咬合式齒輪裝置。依據次擺線曲線f 2 2 4 A 〇如第20圖、第21圖所示，外齒輪220B與齒輪23 0B內咬合。而且，外齒輪2 2 0B與外齒輔同，包括基礎構件222和外齒224B。外齒224B 外齒224A相同並且以相同形狀成型。在此，如形態之一基於次擺並將外齒。再者，本實施形法進行說省略重複減速用內 2 2 2和外件，配置成型。因的加工。量小的小它型外齒輸出用內 ί 220A 相的數量與第20圖 -26- 201231842 所示，外齒 224A和外齒 224B爲在軸向上被態，但基礎構件222是通用的。亦即，外齒_ 2 2 0B爲相同形狀之齒形。震盪體204之偏心量位傳至外齒 224A和外齒 224B。以後’將外έ 224Β統稱爲外齒224。 <齒形決定方法> 對外齒輪220、減速用內齒輪23 0Α及輸出 2 3 0Β之齒形決定方法進行說明。首先，以下槪略說明齒形之求法。最開始將內齒輪之內齒假想爲圓筒形銷’由線公式表示設爲銷半徑ρ2 = 0時的內齒輪之齒形之軌跡，利用其次擺線曲線公式定義外齒輪220 其次，求出外齒輪之齒形實體位置之軌跡’並從義內齒輪之齒形。其次，由外齒輪220和內齒輪小和齒數，對定義外齒輪220和內齒輪230之齒個參數建立相互關聯。其次，決定內齒輪230之頂和齒根之修正範圍。其次，用已建立關聯之參正範圍以外的齒形部份，並由其齒形部份求出數。而且，以將同時咬合數均設爲2以上之方式參數。在決定參數時，以同時滿足扭矩、齒面壓、各部位之主應力 '軸承壽命等目標値之方索0 以下，進行詳細說明分斷之形 [220A ' L以同相 | 224A ' 用內齒輪次擺線曲實體位置之齒形。其軌跡定 23 0之大形的多數齒形的齒數求出修同時咬合決定最佳之容許面式不斷摸 -27- 201231842 最開始定義外齒輪220之齒形》假想配置半徑P2之圓筒形銷作爲減速用內齒輪230A 之內齒228 A (方便起見，設爲減速用內齒輪230A，但亦可配置於輸出用內齒輪230B )，從而求出銷半徑p2 = 0 (與銷中心的意思相同）的減速用內齒輪230A之齒形實體位置之軌跡。而且，之後將向內側（外齒輪220側）僅移動銷的半徑p2者設爲外齒輪220之齒形。以下，進一步具體說明。再者，假想減速比n(nh，ni)與第1實施形態之定義相同。與第1實施形態相同，外齒輪220爲雙圓弧形狀，由式（3)、式（4)表示半徑R1、R2之關係。外齒輪220與假想具備銷的減速用內齒輪230A進行理論咬合》因此，如第22圖所示，由外次擺線曲線公式 (外擺線曲線公式）表示減速用內齒輪230A之銷中心在以偏心軸B爲中心的靜止空間從座標（x4，y4 )向座標 (x5，y5 )移動時所描繪出的軌跡座標（Xp，yp)作爲外齒輪220之齒形實體位置之座標。亦即，若利用以偏心軸 B爲中心固定的基圓BB之半徑b2、沿基圓BB之圓周不滑動地旋轉的旋轉圓AB之半徑a2、描繪點之半徑L2及旋轉角β2，則由式（30)、式（31 )表示外齒輪220之齒形實體位置的座標（xp，yp)。 -28- 201231842The meshing position of the outer gear 120A and the reduction internal gear 130A is rotationally moved in accordance with the movement of the long axis direction X of the vibrating body 104. Here, when the oscillating body 104 rotates by 1 圏, the rotational phase of the external gear 1 20A is delayed only by the difference in the number of teeth from the internal gear 130A for deceleration. In other words, the reduction ratio based on the internal gear 13A for deceleration can be set to ((number of teeth of the outer gear 120A (N*k) - number of teeth of the internal gear 130A for deceleration ((N + l) * k)) / external gear 120A The number of teeth (N*k)) = -1 / Ν ) is obtained. The number of teeth (N*k) of the external gear 120B and the output internal gear 130B are the same, so that the mutually engaged portions of the external gear 120B and the output internal gear 130B do not move, and the same teeth mesh with each other. Therefore, the same rotation as the rotation of the external gear 120B is output from the output internal gear 130B. As a result, the output of the rotation of the oscillating body 104 in accordance with the reduction ratio 1/N based on the reduction internal gear 130A can be taken out from the output internal gear 130B. In the present embodiment, as a basic configuration, the cylindrical external gear 120 is configured to be engaged with two rigid internal gears 130 (the internal gear 130A for deceleration and the internal gear 130B for output), and the external gear 120 and The internal gear 130 is configured to have a tooth shape in which both the number of meshes Nph and Npl of the external gear 120 and the internal gear 130 are two or more, and further, theoretical occlusion is achieved by using a trochoidal curve. Therefore, it is possible to improve the impact resistance, disperse the surface pressure applied to the occlusal tooth surface, and transmit the large twist -23-201231842 moment, especially in comparison with the conventional cup-shaped flexible snap-in gear device, the external gear can be particularly reduced. Local stress generated in 1 2 0. In other words, in the flexible snap-in gear device of the present embodiment, the conical deformation is not caused by the flexibility of the vibrating body, and the occlusal area can be achieved without stress concentration at the bottom of the cup. The increase and the dispersion of the surface pressure can greatly increase the load capacity. Further, in the present embodiment, as shown in the drawings, Fig. 15, and Fig. 19, the pin type deceleration side pitch coefficient Gph < 1, the pin type output side pitch coefficient Gpl > l, 28) Established. That is, as shown in the formula (29), in the present embodiment, from the eccentric shaft B, the pin center (toothed body) of the external gear 1 0 0 is disposed from the eccentric shaft B to the external gear 120A and decelerated. The distance (nh*L) between the node Ph of the internal gear 130A and the distance (n, L) from the eccentric shaft B to the node P! based on the external gear 12BB and the output internal gear 130B" [10] m*L . m*L . -<1,->1 (28) RL RL m* L<R-L<ni*L (2 9) Therefore, when engaged with the internal gear 130A for deceleration The load applied to the teeth 124B outside the external gear 120B when the load of the external gear 120A and the external gear 120A are engaged with the output internal gear 130B have components opposite to each other, and the two loads applied to the external gear 120 can be made. The area is close to the outer circumferential direction of the outer gear 120. In other words, when the occlusion operation is performed from the axial 〇, the two internal gears 1 30 can be set to sandwich only a small number of external teeth 1 24 of -24-201231842. Therefore, it is possible to prevent the engagement of the external gear 1 2 〇 and the internal gear 130 from being shifted due to excessive torque (ratchet phenomenon). That is, the ratchet resistance can be improved. A flexible snap-in gear device that utilizes a cup-shaped external gear that has been actually produced (the internal gear has a toothed solid radius of about 26 mm and a reduction ratio of 1/50 (referred to as a comparative example)) and has the same degree In the flexible snap-in gear device 100 of the present embodiment having the same size and the same reduction ratio, it was confirmed that the ratchet resistance was improved substantially (about four times or more) with respect to the actual measurement of the comparative example. At the same time, it was confirmed by theoretical calculation and test that the rated torque in the comparative example was 3.3 kgfm, whereas the rated biting torque of the flexible snap gear device 100 of the present embodiment was 6.6 kg fm. That is, both the theoretical calculation and the test can confirm that the rated torque is about 2 times. As described above, in the present embodiment, the transmission torque can be increased and the transmission efficiency can be increased. Further, instead of the increase in the transmission torque, the flexible snap gear device 100 can be further compacted. Further, in the present embodiment, the tooth profile of the external gear 120 is set to be the same as the portion that is engaged with the reduction internal gear 130A and the output internal gear 130B, respectively, so that the external gear 126 can be easily processed, which can be suppressed low. Processing cost' and high-precision shape processing. That is, according to the present invention, the transmission torque and the transmission efficiency can be increased by increasing the number of simultaneous engagement of the external gear .120 and the internal gear 130, Nph, Npl. -25-201231842 <<Second Embodiment>> A second embodiment of the present invention will be described in detail with reference to Figs. 20 to 29. In the present embodiment, the external gear is formed by a tooth profile of a line curve instead of the cylindrical pin of the first embodiment, and the outer tooth of the wheel and the base member are integrally formed (referred to as a solid type). The parameters are defined identically, and the tags of the parameters used in the state are also set to be the same. The configuration and the tooth shape determination that are different from the first embodiment will be described with the same reference numerals in the other two portions. <Structure> As shown in Figs. 20 and 21, the external gear 220A is engaged with the gear 230A. The outer gear 220A includes a base member tooth 224A. The base member 222 has a flexible cylindrical shape on the outer side of the oscillating body bearing 210A and is integrated with the external teeth 224A, so that the external teeth 224A can be reduced, and the high precision can be achieved, that is, the gear 220A is suitable for the present embodiment. For load-bearing flexible snap-in gear units. According to the trochoidal curve f 2 2 4 A 〇 As shown in Figs. 20 and 21, the external gear 220B is engaged with the gear 23 0B. Moreover, the external gear 2 2 0B is complementary to the external teeth, including the base member 222 and the external teeth 224B. External teeth 224B The outer teeth 224A are identical and shaped in the same shape. Here, one of the forms is based on the secondary pendulum and the external teeth. Further, in the present embodiment, the inside of the deceleration 2 2 2 and the outer member are omitted, and the molding is arranged. Due to the processing. The small amount is small. The external member 222A and the external teeth 224B are axially aligned, but the base member 222 is common, as shown in Fig. 20-26-201231842. That is, the external teeth _ 2 2 0B are tooth shapes of the same shape. The eccentricity of the oscillating body 204 is transmitted to the external teeth 224A and the external teeth 224B. Later, the outer 224 is called the external tooth 224. <Tooth shape determination method> The tooth profile determination method of the external gear 220, the reduction internal gear 23 0Α, and the output 2 3 0Β will be described. First, the following is a brief description of the method of determining the tooth shape. Initially, the internal gear of the internal gear is assumed to be a cylindrical pin. The trajectory of the tooth profile of the internal gear when the pin radius ρ2 = 0 is expressed by the line formula, and the external gear 220 is defined by the second cycloid curve formula. The trajectory of the toothed solid position of the outer gear 'and the tooth shape of the inner gear. Secondly, the external gear 220 and the internal gear are small and the number of teeth is used to establish a correlation between the tooth parameters defining the external gear 220 and the internal gear 230. Next, the correction range of the top and the root of the internal gear 230 is determined. Secondly, the toothed portion outside the reference range that has been associated is used, and the number is determined from the toothed portion. Further, the parameters are set such that the number of simultaneous occlusions is 2 or more. In the case of determining the parameters, the target of the torque, the tooth surface pressure, the principal stress of each part, the bearing life, etc., is below 0, and the shape of the breaking is further described [220A 'L with the same phase | 224A ' internal gear The tooth shape of the position of the oscillating line. The number of teeth of most of the tooth shapes whose trajectory is set to a size of 23 0 is determined by the simultaneous occlusion. The optimum allowable surface type is constantly touched. -27- 201231842 The tooth profile of the external gear 220 is initially defined. The imaginary cylindrical shape with a radius P2 is configured. The pin is used as the internal tooth 228 A of the internal gear 230A for deceleration (for convenience, it is set as the internal gear 230A for deceleration, but it can also be disposed in the internal gear 230B for output) to obtain the pin radius p2 = 0 (with the center of the pin) The trajectory of the toothed solid position of the internal gear 230A for deceleration. Further, the radius p2 at which only the pin is moved to the inner side (the outer gear 220 side) is set to the tooth shape of the outer gear 220. The following is further explained in detail. Further, the virtual reduction ratio n (nh, ni) is the same as the definition of the first embodiment. Similarly to the first embodiment, the external gear 220 has a double arc shape, and the relationship between the radii R1 and R2 is expressed by the equations (3) and (4). The external gear 220 is theoretically engaged with the decelerating internal gear 230A having the pin. Therefore, as shown in Fig. 22, the pin center of the decelerating internal gear 230A is expressed by the outer cycloid curve formula (external cycloid curve formula). The track coordinates (Xp, yp) drawn by the stationary space centered on the eccentric axis B from the coordinates (x4, y4) to the coordinates (x5, y5) serve as the coordinates of the toothed physical position of the external gear 220. That is, when the radius b2 of the base circle BB fixed around the eccentric axis B2, the radius a2 of the rotation circle AB that does not slide along the circumference of the base circle BB, the radius L2 of the drawing point, and the rotation angle β2 are used, Equations (30) and (31) represent coordinates (xp, yp) of the position of the toothed body of the external gear 220. -28- 201231842

Xp = (62 + α2) * cosβ 2-L2*cos b2 + a2 a2 β2 ) » = (62 + a2)*sinj3 2-L2*sini b2+al /32 l a2 y 其中，若利用式（32 )〜（34)之關係，即 (3 5 )、式（3 6 )之關係。 [數 12] a2 = —R ··· (3 2) m + 1 …（30) …（3 1) 可求出式 b2 = -^R -.-(3 3) m + 1 2 = j3、L2 = L ---(34) xP = R*cosβ-L*cos((m +1)*j3) ... (3 5) » = Λ * sin ]3 - Z * sin((m + \)* β) ...(36) 其次，使外齒輪220之齒形實體位置之g yP)向內側（外齒輪220側）僅移動假想爲內磨半徑p2。如此一來，可由式（37)〜（39) 0 軸Fc爲原點的外齒輪220之齒形的座標（xkfc， [數 13] Xk/c = Xp-p2* COS0 +L ytrfc = yP-p 2*s'm<i) ,i? * sin ]3 - Z, * (m +1) * sin((m + \)* β) φ =tan -- R * cos j3 - L * (m +1) * cos((m +1) * j3 ) …（37) …（38) 標（χΡ ’ 2 2 8之銷示以旋轉 y k fc)。 …（3 9) -29- 201231842 亦即，藉由代入半徑R、p2、偏心量L、假想減速 nh並使角度β變化，可求出外齒輪220之齒形的座 (xkfc 5 Ykfc )。其次，定義內齒輪230之齒形。求出外齒輪220之形實體位置的座標（xp，yp )之包絡線，使其包絡線向側（內齒輪23 0側）僅移動半徑p2而設爲內齒輪230 齒形軌跡。亦即，關於減速用內齒輪2 3 0 A，重新求出齒形。以下，進一步具體說明。外齒輪220之以偏心軸B爲中心的Xd-yd座標上的齒輪22〇之齒形軌跡Q (第23圖中所示之2個虛線份）在旋轉角度α時，如第23圖所示描繪包絡線（第圖中所示之實線部份）。因此，利用式（3 0 ) 、（ 3 1 ) 由式（40) 、（41)表示以旋轉軸Fc爲原點的內齒 230之齒形實體位置的座標（xpf(：，ypfe)。其中，藉由用作爲包絡線條件式的式（42)，由式（43)表不角比標齒內之其外部 23 輪利度 α、β之關係· [數 14] Ap/c=^>*cos—+>j»*sin—+L*cosa …（4 0) γφ = -xp * sin—+,y/. * cos—+Z, * sin a …（4 1) , dyufc _ dypfe ^ dxp/c 1J = U …（4 2) =_L* sin-1 r \ m i?*sin(nfc*3) n+l \A*cos0-I*(ra+l)*cos((m+I)*0)j …（4 3) -30- 201231842 其次，藉由使內齒輪230之齒形實體位置的座標 (Xpfc，yPf。）向內側（內齒輪23 0側）僅移動假想爲內齒 228的銷半徑 p2，可由式（44) 、（45)求出以旋轉軸 F c爲原點的內齒輪2 3 0之齒形的座標（X fe ’ y f。）。 [數 15]Xp = (62 + α2) * cosβ 2-L2*cos b2 + a2 a2 β2 ) » = (62 + a2)*sinj3 2-L2*sini b2+al /32 l a2 y where, if using equation (32) The relationship of ~(34), that is, the relationship between (3 5 ) and (3 6 ). [12] a2 = —R ··· (3 2) m + 1 (30) (3 1) Equation b2 = -^R -.-(3 3) m + 1 2 = j3, L2 = L ---(34) xP = R*cosβ-L*cos((m +1)*j3) ... (3 5) » = Λ * sin ]3 - Z * sin((m + \ *β) (36) Next, the g yP) of the tooth-shaped solid position of the external gear 220 is moved to the inner side (the outer gear 220 side) by the imaginary inner grinding radius p2. In this way, the coordinates of the tooth profile of the external gear 220 which can be originated by the equation (37) to (39) 0 axis Fc (xkfc, [13] Xk/c = Xp-p2* COS0 + L ytrfc = yP- p 2*s'm<i) ,i? * sin ]3 - Z, * (m +1) * sin((m + \)* β) φ =tan -- R * cos j3 - L * (m +1) * cos((m +1) * j3 ) ...(37) ...(38) Mark (χΡ ' 2 2 8's pin to rotate yk fc). (3 9) -29- 201231842 That is, by substituting the radii R, p2, the eccentric amount L, the imaginary deceleration nh, and changing the angle β, the tooth-shaped seat (xkfc 5 Ykfc ) of the external gear 220 can be obtained. Next, the tooth profile of the internal gear 230 is defined. The envelope of the coordinate (xp, yp) of the physical position of the external gear 220 is obtained, and the envelope side (the inner gear 23 0 side) is moved by the radius p2 to be the tooth profile of the internal gear 230. That is, the tooth profile is re-determined with respect to the internal gear 2 3 0 A for deceleration. Hereinafter, it will be described in further detail. The tooth profile trajectory Q of the gear 22 〇 on the Xd-yd coordinate centered on the eccentric axis B of the outer gear 220 (the two broken lines shown in FIG. 23) is at the rotation angle α as shown in FIG. Describe the envelope (the solid line portion shown in the figure). Therefore, the coordinates (xpf(:, ypfe) of the tooth-shaped solid position of the internal tooth 230 whose origin is the rotation axis Fc are expressed by the equations (40) and (41) by the equations (30) and (31). By using equation (42) as the envelope conditional expression, the relationship between the outer angles of the equations (43) and the outer 23 rounds of the scales α and β in the standard teeth is selected. [Number 14] Ap/c=^&gt ;*cos—+>j»*sin—+L*cosa ...(4 0) γφ = -xp * sin—+, y/. * cos—+Z, * sin a ...(4 1) , dyufc _ Dypfe ^ dxp/c 1J = U ...(4 2) =_L* sin-1 r \ mi?*sin(nfc*3) n+l \A*cos0-I*(ra+l)*cos((m +I)*0)j (4 3) -30- 201231842 Next, by moving the coordinates (Xpfc, yPf.) of the toothed solid position of the internal gear 230 to the inner side (the inner gear 23 0 side), only the hypothesis is The pin radius p2 of the internal teeth 228 can be obtained by the equations (44) and (45) for the coordinate (X fe ' yf.) of the internal gear of the internal gear 203 with the rotation axis F c as the origin. [Number 15]

XfC = xPfc-p2*cosiφ~— …（4 4)XfC = xPfc-p2*cosiφ~— (4 4)

I nJ (a\ . λ y/c = ypfc-ρ 2*sm Φ-— ··· (4 5) 、n) 亦即，藉由代入半徑R、p2、偏心量L、假想減速比 nh、〜並使角度β變化，可求出減速用內齒輪23 0A和輸出用內齒輪230Β之齒形的各個座標（Xfc，yfc )。其次，對規定外齒輪220和內齒輪230的參數建立相互關聯。如上述，與第1實施形態相同，外齒輪220之形狀爲由半徑R1、R2規定的雙圓弧形狀。亦即，式（19)、式 (20 )之關係在本實施形態中亦成立。其次，導入參數 Gs (稱爲實心型節距係數）。其中，將穿過偏心軸B和旋轉軸Fc的直線與由外齒輪22〇 (之外齒224)和內齒輪230(之內齒228)之咬合產生的接觸點之共同法線之交點稱爲基於外齒輪2 2 0和內齒輪 230之節點（亦即，節點之定義與第1實施形態相同）。實心型節距係數Gs與銷型節距係數Gp相同，係爲了可容易掌握外齒輪220和內齒輪230各自的齒形實體位置與 -31 - 201231842 節點的相對位置關係且可容易調整這些者。具體而言，如式（46 )所示，實心些從旋轉軸Fc至基於外齒輪220和減速用點的距離（n+1) *L與半徑R之比表示* 參數彼此而導入節距係數Gs由內齒輪230之節I nJ (a\ . λ y/c = ypfc-ρ 2*sm Φ-- ··· (4 5) , n) that is, by substituting the radius R, p2, the eccentric amount L, the virtual reduction ratio nh, When the angle β is changed, the respective coordinates (Xfc, yfc) of the tooth profile of the reduction internal gear 23 0A and the output internal gear 230 可 can be obtained. Secondly, the parameters defining the external gear 220 and the internal gear 230 are related to each other. As described above, the shape of the external gear 220 is a double arc shape defined by the radii R1, R2, as in the first embodiment. That is, the relationship between the equations (19) and (20) is also established in the present embodiment. Second, import the parameter Gs (called the solid pitch coefficient). Wherein, the intersection of the straight line passing through the eccentric shaft B and the rotation axis Fc with the common normal of the contact points generated by the engagement of the external gear 22 〇 (outer teeth 224 ) and the internal gear 230 (the internal teeth 228 ) is called The node based on the external gear 220 and the internal gear 230 (that is, the definition of the node is the same as that of the first embodiment). The solid pitch coefficient Gs is the same as the pin type pitch coefficient Gp in order to easily grasp the relative positional relationship between the respective toothed physical positions of the external gear 220 and the internal gear 230 and the -31 - 201231842 nodes and can be easily adjusted. Specifically, as shown in the formula (46), the distance from the rotation axis Fc to the distance based on the external gear 220 and the deceleration point (n+1) *L and the radius R are expressed as * parameters are introduced to each other and the pitch coefficient is introduced. Gs by the inner gear 230 section

[數 16] Gs = (^A R …（4 6)[Number 16] Gs = (^A R ... (4 6)

在第24圖中表示內齒輪230之齒形想減速比IU之關係。依據式（46 )，蔣型節距係數Gsh (稱爲實心型減速側節 (47 )。若設參數i=l，並由式（19 )、 (47)，則得出式（48)。 [數 17] G^ = ^ + 1)*L …（4 7) R 實體半徑R與假此時得到的實心係數）定義成式式（20 )整理式 x*m1+ 2Gsh —f *Ν+ {>i(\-Gsh) + 2Csh —j , *m-[1 ΚΗ j 羽[ 在第25圖中表示內齒輪230之齒形想減速比n i之關係。依據式（4 6 )，將型節距係數Gsl (稱爲實心型輸出側節进 (49 ) ^若設參數i = 0，並由式（19 )、 (49 )，則能夠得出式（5 0 )、式（5 1 ) (1_GjA)jA^ = 0 …（4 8) 實體半徑R與假 :此時得到的實心 ί係數）定義成式式（20)整理式 -32- 201231842 [數 18]Fig. 24 shows the relationship between the tooth profile of the internal gear 230 and the reduction ratio IU. According to the formula (46), the Jiang type pitch coefficient Gsh (referred to as the solid type deceleration side section (47). If the parameter i = l, and by the equations (19), (47), the equation (48) is obtained. [Number 17] G^ = ^ + 1)*L ...(4 7) R The solid radius R and the solid coefficient obtained at this time are defined as the formula (20). The formula x*m1+ 2Gsh —f *Ν+ { >i(\-Gsh) + 2Csh —j , *m-[1 ΚΗ j Feather [In Fig. 25, the relationship between the tooth profile of the internal gear 230 and the reduction ratio ni is shown. According to the formula (4 6 ), the type pitch coefficient Gsl (referred to as the solid type output side pitch (49) ^, if the parameter i = 0, and by the equations (19), (49), the formula can be obtained ( 5 0 ), (5 1 ) (1_GjA)jA^ = 0 (4 8) The radius of the solid R and the false: the solid ί coefficient obtained at this time is defined as the formula (20) finishing -32- 201231842 18]

{R-L)n (4 9){R-L)n (4 9)

--Θ {R-L)n + 2L-^-~ sin -τ-θ …（5 0) (5 1)--Θ {R-L)n + 2L-^-~ sin -τ-θ (5 0) (5 1)

因此，若給出半徑R、減速比N、實心型減速側節距係數G s h及咬合角度Θ ’則可以決定假想減速比n h、偏心量L，接著可以求出實心型輸出側節距係數G s 1、假想減速比…。本實施形態亦與第1實施形態相同’如第24圖、第 2 5圖所示，代入實心型減速側節距係數Gsh< 1，求出實心型輸出側節距係數Gsl> 1之値。本實施形態亦與第1實施形態相同，從求出各齒形之結果考慮，咬合角度Θ爲40 〜65度且銷型減速側節距係數Gph之cos·1之値爲15〜 3 〇度之情況爲進一步更佳的條件。其次，決定內齒輪230之齒形之修正範圍。與第1實施形態相同，修正內齒228之齒頂和齒根。因此未進行齒形修正的角度ps〜pf (未修正之齒形區域）成爲進行理論咬合的有效範圍。 -33- 以外出。 (53 用內 Nsh = Nsl = 將設 N sh 圖、上的態中 (K = 之齒定本 …（5 2) …（5 3) 201231842 其次，求出同時咬合數Nsh、Nsl。與第1實施形態相同，同時咬合數Nsh、Nsl可齒輪220之旋轉角度α決定的有效範圍除以節距亦即，利用式（43 )之關係，分別由式（52 ) )求出減速用內齒輪23 0Α之同時咬合數Nsh、齒輪23 0B之同時咬合數Nsl。 [數 19] a fh_ a sh _ α (]3 fh) - α (j3 sh) 2π 2π (N + \)*k (N + \)*k α π~α si _ ct (η)-〇! (]3 si) In ~ 2π N*k N*k 沿式（52 )、式（53 )求出同時咬合數。此時，爲k = 2時求出的減速用內齒輪23 0A之同時咬 •輸出用內齒輪23 0B之同時咬合數Nsl示於負第27圖中。依這些同時咬合數Nsh、同時咬合數Nsl均實現直徑（2*R)和減速比（1/N)之條件，決定本實的內齒輪 230之齒形。亦即，當齒數差爲 =2 )，減速比（1/N )爲1/30而不會成爲本實施形，由1/50以下（比1/50更大地減速的減速比實施形態之內齒輪之齒形。在本實施形態中，與基礎構件222 —體成型故容易加工外齒輪220，且可實現高精度的加工關於其他內容，在本實施形態中亦可得到與第1 藉由角求、式輸出分別合數 I 26 2以施形 2時形態 )決外齒〇實施 -34- 224 201231842 形態幾乎相同的作用效果。例如，在本實施形態中亦與第1實施形態相同’如第 24圖、第25圖、第28圖所示’由於設爲實心型減速側節距係數G s h < 1、實心型輸出側節距係數G s 1 > 1 ’故式 (54)成立。亦即，如式（55)所示’將內齒輪230之內齒2 2 8假想爲銷時的銷中心（齒形實體）之位置配置於從旋轉軸Fc至基於外齒輪220A和減速用內齒輪230A之節點Ph爲止的距離（（nh+l ) *L )與從旋轉軸Fc至基於外齒輪220B和輸出用內齒輪230B之節點Pi爲止的距離 ((n〗+l ) *L)之間。 [數 20] + (»/ + !)*! ^ ... (54)Therefore, if the radius R, the reduction ratio N, the solid type deceleration side pitch coefficient G sh and the nip angle Θ ' are given, the virtual reduction ratio nh and the eccentric amount L can be determined, and then the solid type output side pitch coefficient G can be obtained. s 1, imaginary reduction ratio... The present embodiment is also the same as the first embodiment. As shown in Fig. 24 and Fig. 25, the solid type deceleration side pitch coefficient Gsh < 1, is obtained, and the solid type output side pitch coefficient Gsl > 1 is obtained. Also in the present embodiment, as in the first embodiment, the occlusion angle Θ is 40 to 65 degrees from the result of obtaining each tooth profile, and the cos·1 of the pin type deceleration side pitch coefficient Gph is 15 to 3 degrees. The situation is further better conditions. Next, the correction range of the tooth profile of the internal gear 230 is determined. As in the first embodiment, the addendum and the root of the inner teeth 228 are corrected. Therefore, the angle ps to pf (uncorrected tooth-shaped region) in which the tooth profile correction is not performed becomes an effective range for performing theoretical occlusion. -33- Outside. (53) Nsh = Nsl = will be set to the N sh diagram, the state above (K = the tooth version... (5 2) ... (5 3) 201231842 Next, find the number of simultaneous occlusions Nsh, Nsl. In the same form, the effective range determined by the rotation angle α of the nip number Nsh and the Nsl gear 220 is divided by the pitch, that is, the internal gear for deceleration 23 0 is obtained from the equation (52) by the relationship of the equation (43). At the same time, the number of occlusions Nsh and the number of teeth of the gears 23 0B are Nsl. [19] a fh_ a sh _ α (]3 fh) - α (j3 sh) 2π 2π (N + \)*k (N + \) *k α π~α si _ ct (η)-〇! (]3 si) In ~ 2π N*k N*k Find the number of simultaneous occlusions along equations (52) and (53). In this case, k = 2 at the same time as the deceleration internal gear 23 0A, the simultaneous biting and output internal gear 23 0B at the same time the number of occlusions Nsl is shown in the negative 27th diagram. According to these simultaneous occlusion numbers Nsh, the number of occlusions Nsl both achieve the diameter ( The condition of 2*R) and the reduction ratio (1/N) determines the tooth profile of the actual internal gear 230. That is, when the difference in the number of teeth is =2), the reduction ratio (1/N) is 1/30. Will become this embodiment, from 1/50 or less (decelerated more than 1/50) In the present embodiment, the outer gear 220 is easily formed by the basic member 222, and the high-precision machining can be realized. Other things can be obtained in the present embodiment. In the first embodiment, the shape is obtained by the angle I and the output of the equation I 26 2, and the shape is almost the same. For example, in the present embodiment, it is the same as in the first embodiment. 'As shown in Fig. 24, Fig. 25, and Fig. 28', it is set as a solid type deceleration side pitch coefficient G sh < 1 and a solid type output side. The pitch coefficient G s 1 > 1 'Formula (54) holds. That is, as shown in the formula (55), the position of the pin center (toothed body) when the internal tooth 2 2 8 of the internal gear 230 is assumed to be a pin is disposed from the rotation axis Fc to the external gear 220A and the deceleration The distance ((nh+l)*L) from the node Ph of the gear 230A to the distance ((n+1+l)*L) from the rotation axis Fc to the node Pi based on the external gear 220B and the output internal gear 230B between. [Number 20] + (»/ + !)*! ^ ... (54)

R 5 R (m + \)*L < R < ^ηι + {)*L ..-(5 5) 因此，與減速用內齒輪230A咬合時施加於外齒輪 220A之外齒224A的負載Fd和與輸出用內齒輪230B咬合時施加於外齒輪220B之外齒224B的負荷F〇具備相互反方向之成份，並且可以使施加於外齒輪220的該2個負載Fd、Fo之區域在外齒輪220之外周方向上靠近。亦即，如第29圖所示，從軸向0觀察，進行咬合動作時，能夠使負載Fd和負載Fo之區域靠近而將2個內齒輪230 設爲僅夾入少數外齒224之態樣。因此，與第1實施形態相同，能夠提高耐棘輪性。再者，式（29)和式（55)均可變形爲式（56)。 -35- 201231842 [數 21] -<L< 今·· …（56) ni + \ m + \ 亦即，在上述實施形態中，當將外齒輪120之外齒 124設爲圓筒形銷時，或者將內齒輪230之內齒228設爲 (假想爲）圓筒形銷時，該銷中心0 c配置於穿過旋轉軸 Fc和偏心軸B的直線與由外齒輪120、220與內齒輪 130、230之咬合產生的各個接觸點之共同法線之交點亦即節點Ph、P,之間，故可以提高耐棘輪性。舉出上述實施形態對本發明進行了說明，但本發明並不限於上述實施形態。亦即，在不脫離本發明的宗旨之範圍內可進行改良及設計之變更，這不言而喻的。例如，在上述實施形態中，將同時咬合數Nph、 Npl、Nsh、Nsl設爲2以上時，依據次擺線曲線求出了外齒輪或內齒輪之齒形，但本發明不限於此。例如，可以按同一道理從所求出的內齒輪之齒形的座標求出由外齒輪與內齒輪之咬合產生的接觸點之軌跡亦即接觸線，故亦可利用該接觸線。以下具體說明第1實施形態時的內齒輪1 3 0 之齒形的座標與接觸線之同理關係。接觸線CL成爲將內齒輪130之齒形的座標（xfc， yfc)旋轉α角度的從第30圖所示的X-Y座標系觀察的軌跡。因此，由將內齒輪130之齒形的座標（xfe，yfe )旋轉 α角度的式（5 7 )、式（5 8 )給出接觸線的座標 (Xcfc，ycfc) 0 -36- 201231842 [數 22] …（5 7) …（5 8)R 5 R (m + \)*L < R < ^ηι + {)*L .. - (5 5) Therefore, the load applied to the teeth 224A outside the external gear 220A when engaged with the internal gear 230A for deceleration When the Fd and the output internal gear 230B are engaged with each other, the load F〇 applied to the teeth 224B outside the external gear 220B has a component opposite to each other, and the two loads Fd and Fo applied to the external gear 220 can be placed on the external gear. 220 is close in the outer circumferential direction. That is, as shown in Fig. 29, when the occlusion operation is performed from the axial direction 0, the area of the load Fd and the load Fo can be brought close to each other, and the two internal gears 230 can be set to sandwich only a small number of external teeth 224. . Therefore, as in the first embodiment, the ratchet resistance can be improved. Furthermore, both equations (29) and (55) can be modified to equation (56). -35- 201231842 [Number 21] - <L< Now... (56) ni + \ m + \ That is, in the above embodiment, when the external gear 120 is made of a cylindrical pin When the internal teeth 228 of the internal gear 230 are set to (imaginary) cylindrical pins, the pin center 0 c is disposed on a straight line passing through the rotating shaft Fc and the eccentric shaft B and by the external gears 120, 220 and The intersection of the common normals of the respective contact points generated by the engagement of the gears 130 and 230, that is, between the nodes Ph and P, can improve the ratchet resistance. The present invention has been described by way of the above embodiments, but the present invention is not limited to the above embodiments. That is, it is self-evident that modifications and design changes can be made without departing from the spirit of the invention. For example, in the above-described embodiment, when the number of simultaneous nips Nph, Npl, Nsh, and Nsl is 2 or more, the tooth profile of the external gear or the internal gear is obtained from the trochoid curve, but the present invention is not limited thereto. For example, the contact line of the contact point generated by the engagement of the external gear and the internal gear, that is, the contact line, can be obtained from the coordinates of the tooth profile of the internal gear obtained in the same manner, and the contact line can also be utilized. The similarity between the coordinates of the tooth profile of the internal gear 1 300 and the contact line in the first embodiment will be specifically described below. The contact line CL is a trajectory observed from the X-Y coordinate system shown in Fig. 30, in which the coordinates (xfc, yfc) of the tooth shape of the internal gear 130 are rotated by an angle α. Therefore, the coordinate of the contact line (Xcfc, ycfc) is given by the equation (5 7 ) and the equation (5 8 ) which rotates the coordinate of the tooth profile of the internal gear 130 (xfe, yfe ) by an angle (0cfc, ycfc) 0 -36- 201231842 22] ...(5 7) ...(5 8)

Xc/c = x/c * cos a - 3//C * sin a y〇fc - x/c * sin a +y/c * cos a 將由上式得出的接觸線C L示於第3 1圖中。可知，接觸線C L描繪於外齒輪1 2 0和內齒輪1 3 0的多數個齒頂與齒根之中間，可確保多數個同時咬合數Nph、Npl。因此，也可利用該接觸線設想可確保多數個同時咬合數Nph、Npl的接觸線並由此求出內齒輪之齒形。又，在上述實施形態中，使減速側節距係數Gph、Xc/c = x/c * cos a - 3//C * sin ay〇fc - x/c * sin a +y/c * cos a The contact line CL obtained by the above formula is shown in Fig. 31 . It can be seen that the contact line C L is drawn between the outer end of the external gear 1 120 and the internal gear 1 30 and the root of the tooth to ensure a plurality of simultaneous occlusions Nph, Npl. Therefore, it is also possible to use the contact line to ensure that a plurality of contact lines of Nph and Npl are simultaneously engaged and thereby obtain the tooth profile of the internal gear. Further, in the above embodiment, the deceleration side pitch coefficient Gph,

Gsh小於1且使輸出側節距係數Gpl、Gsl大於1，但本發明未必一定限定於這種關係。例如，亦可使減速側節距係數G p h、G s h大於1且使輸出側節距係數G p 1、G s 1小於 1。又，並非否定使所有節距係數均大於1或者使所有節距係數均小於1之類的情況。這是因爲，不僅是規定節距係數的參數，並且還藉由不斷摸索並決定多數個參數之調整來求出外齒輪和內齒輪之齒形。 (產業上之實用性）本發明之撓性咬合式齒輪裝置可使用於各種用途中，例如可較佳使用於產業用機械手之關節（手腕）驅動裝置或工作機械等精密控制用途中。【圖式簡單說明】第1圖爲表不本發明之第1實施形態之撓性咬合式齒 -37- 201231842 輪裝置之整體結構之一例的分解立體圖。第2圖爲表示該裝置之整體結構之一例的剖視圖。第3圖爲表示該裝置之震盪體的圖。第4圖爲表示該裝置之震盪體的圖。第5圖爲組合該裝置之震盪體和震盪體軸承的槪略圖。第6圖爲該裝置之外齒輪與內齒輪之咬合圖。第7圖爲該裝置之外齒輪與減速用內齒輪及輸出用內齒輪之咬合放大圖。第8圖爲表示該裝置之外齒輪與減速用內齒輪及輸出用內齒輪之齒形實體位置的圖。第9圖爲定義該裝置之外齒輪之齒形的圖。第1〇圖爲定義該裝置之減速用內齒輪、輸出用內齒輪之齒形的圖。第11圖爲定義該裝置之減速用內齒輪、輸出用內齒輪之齒形的圖。第12圖爲定義該裝置之減速用內齒輪、輸出用內齒輪之齒形的圖。第13圖爲表示該裝置之減速用內齒輪、輸出用內齒輪及外齒輪之周長、齒數及節距之關係的表。第14圖爲表示該裝置之節點與外齒輪之實體位置之關係的圖。第15圖爲表示該裝置之節點與外齒輪之實體位置之關係的圖。 -38- 201231842 第16圖爲表示該裝置之減速用內齒輪、輸出用內齒輪之齒形修正的圖。第17圖爲表示第1實施方式中變更減速比和內齒輪之直徑時減速用內齒輪中的同時咬合數的表。第1 8圖爲表示第1實施方式中變更減速比和內齒輪之直徑時輸出用內齒輪中的同時咬合數的表。第19圖爲表示第1實施方式中外齒輪之實體位置與節點之關係的圖。第20圖爲表示本發明之第2實施方式之撓性咬合式齒輪裝置之整體結構之一例的分解立體圖。第2 1圖爲表示該裝置之整體結構之一例的剖視圖。第22圖爲定義該裝置之外齒輪之齒形的圖。第23圖爲定義該裝置之減速用內齒輪、輸出用內齒輪之齒形的圖。第24圖爲表示該裝置之節點與內齒輪之實體位置之關係的圖。第25圖爲表示該裝置之節點與內齒輪之實體位置之關係的圖。第20圖爲表示第2實施形態中變更減速比和內齒輪之直徑時減速用內齒輪中的同時咬合數的表。第27圖爲表示第2實施形態中變更減速比和內齒輪之直徑時輸出用內齒輪中的同時咬合數的表。第28圖爲表示第2實施形態中內齒輪之實體位置與節點之關係的圖。 -39- 201231842 第29圖爲表示第2實施形態中的棘輪防止效果的圖。第3 0圖爲用以求出第1實施形態中外齒輪與減速用內齒輪及輸出用內齒輪之接觸線的圖。第31圖爲表示該裝置之接觸線的圖。【主要元件符號說明】 100、200:撓性咬合式齒輪裝置 104 、 204 :震盪體 1 10A、110B、210A、2 10B :震盪體軸承 114A、 114B、 214A、 214B:保持器 116A ' 116B、 216A ' 216B ：滾子 120、 120A、 120B、 220、 220A、 220B ：外齒輪 122、222:基礎構件 124、 124A ' 124B、 224、 224A、 224B ：外齒 128、128A ' 128B、228、228A、228B ：內齒 130、 130A、 130B ' 230、 230A、 230B ：內齒輪 al、a2 :旋轉圓之半徑 AA、AB :旋轉圓 B :偏心軸 bl、b2 :基圓之半徑 B A、BB :基回 C L :接觸線 FA:第1圓弧部（咬合範圍） -40- 201231842 F c :旋轉軸 F d、F ο :負載 Gp、Gph、Gpl、Gs、 L :偏心量 η、n h、n t :假想減速 N :減速比（之倒數） Nph、Npl ' Nsh、Nsl 〇 :軸向 Oc :銷中心 P h、P !:節點 R:內齒輪之齒形實儀 R 1 :外齒輪之咬合範 R2 :外齒輪之非咬合 SA :第2圓弧部（非 pi、p2 :圓筒形銷半=Gsh is less than 1 and the output side pitch coefficients Gpl, Gsl are greater than 1, but the present invention is not necessarily limited to this relationship. For example, the deceleration side pitch coefficients G p h, G s h may be greater than 1 and the output side pitch coefficients G p 1 , G s 1 may be made smaller than 1. Further, it is not a case of negating that all the pitch coefficients are greater than one or that all the pitch coefficients are less than one. This is because not only the parameters of the pitch coefficient are specified, but also the tooth profile of the external gear and the internal gear is obtained by continuously groping and determining the adjustment of a plurality of parameters. (Industrial Applicability) The flexible snap-in type gear device of the present invention can be used in various applications, and can be preferably used, for example, in a precision control application such as an industrial robot joint (wrist) drive device or a work machine. [Brief Description of the Drawings] Fig. 1 is an exploded perspective view showing an example of the overall configuration of a flexible snap-in tooth-37-201231842 wheel device according to the first embodiment of the present invention. Fig. 2 is a cross-sectional view showing an example of the overall configuration of the apparatus. Fig. 3 is a view showing a oscillating body of the device. Fig. 4 is a view showing a vibrating body of the device. Figure 5 is a schematic diagram of the combination of the oscillating body and the oscillating body bearing of the device. Figure 6 is a bite view of the outer gear and the internal gear of the device. Fig. 7 is an enlarged view showing the engagement of the external gear of the device with the internal gear for reduction and the internal gear for output. Fig. 8 is a view showing the position of the toothed body of the external gear of the device, the internal gear for deceleration, and the internal gear for output. Figure 9 is a diagram defining the tooth profile of the gears outside the device. The first drawing is a view for defining the tooth profile of the internal gear for deceleration and the internal gear for output of the device. Fig. 11 is a view showing the tooth profile of the internal gear for deceleration and the internal gear for output of the apparatus. Fig. 12 is a view showing the tooth profile of the internal gear for deceleration and the internal gear for output of the apparatus. Fig. 13 is a table showing the relationship between the circumference, the number of teeth, and the pitch of the internal gear for deceleration, the internal gear for output, and the external gear of the device. Figure 14 is a diagram showing the relationship between the node of the device and the physical position of the external gear. Figure 15 is a diagram showing the relationship between the node of the device and the physical position of the external gear. -38- 201231842 Fig. 16 is a view showing the tooth profile correction of the internal gear for deceleration and the internal gear for output of the device. Fig. 17 is a table showing the number of simultaneous occlusions in the internal gear for deceleration when the reduction ratio and the diameter of the internal gear are changed in the first embodiment. Fig. 18 is a table showing the number of simultaneous occlusions in the output internal gear when the reduction ratio and the diameter of the internal gear are changed in the first embodiment. Fig. 19 is a view showing the relationship between the physical position of the external gear and the node in the first embodiment. Fig. 20 is an exploded perspective view showing an example of the overall configuration of the flexible snap gear device according to the second embodiment of the present invention. Fig. 21 is a cross-sectional view showing an example of the overall configuration of the apparatus. Figure 22 is a diagram defining the tooth profile of the gears outside the device. Fig. 23 is a view showing the tooth profile of the internal gear for deceleration and the internal gear for output of the apparatus. Figure 24 is a diagram showing the relationship between the node of the device and the physical position of the internal gear. Figure 25 is a diagram showing the relationship between the node of the device and the physical position of the internal gear. Fig. 20 is a table showing the number of simultaneous occlusions in the internal gear for deceleration when the reduction ratio and the diameter of the internal gear are changed in the second embodiment. Fig. 27 is a table showing the number of simultaneous engagements in the output internal gear when the reduction ratio and the diameter of the internal gear are changed in the second embodiment. Fig. 28 is a view showing the relationship between the physical position of the internal gear and the node in the second embodiment. -39-201231842 Fig. 29 is a view showing the ratchet preventing effect in the second embodiment. Fig. 30 is a view for obtaining a contact line between the external gear, the internal gear for reduction, and the internal gear for output in the first embodiment. Figure 31 is a view showing the contact line of the device. [Description of main component symbols] 100, 200: Flexible snap-in gear device 104, 204: Oscillator 1 10A, 110B, 210A, 2 10B: Oscillator bearings 114A, 114B, 214A, 214B: Retainers 116A '116B, 216A ' 216B : Rollers 120 , 120A , 120B , 220 , 220A , 220B : external gears 122 , 222 : base member 124 , 124A ' 124B , 224 , 224A , 224B : external teeth 128 , 128A ' 128B , 228 , 228A , 228B : internal teeth 130, 130A, 130B '230, 230A, 230B: internal gears a1, a2: radius AA of the rotating circle, AB: rotation circle B: eccentric axis bl, b2: radius of the base circle BA, BB: base back CL : contact line FA: first circular arc (biting range) -40- 201231842 F c : rotation axis F d, F ο : load Gp, Gph, Gpl, Gs, L: eccentricity η, nh, nt: hypothetical deceleration N : reduction ratio (reciprocal) Nph, Npl ' Nsh, Nsl 〇: axial Oc : pin center P h, P !: node R: internal gear tooth profile R 1 : external gear bite range R2 : outside Non-occluded gear SA: 2nd circular arc (non-pi, p2: cylindrical pin half =

Gsh、Gsl :節距係數比（之倒數） :同時咬合數丨半徑圍之齒形實體半徑範圍之齒形實體半徑咬合範圍） -41 -Gsh, Gsl: pitch coefficient ratio (reciprocal): simultaneous occlusion number 丨 radius circumference of the toothed solid radius range of the toothed solid radius occlusion range) -41 -

Claims

201231842 七、申請專利範圍 1. 一種撓性咬合式齒輪裝置，具備：震盪體：筒形外齒輪’配置於該震盪體之外周並具有依該震盪體之旋轉而撓性變形的可撓性；第1內齒輪，具有該外齒輪內咬合的剛性；及第2內齒輪，軸向上與該第1內齒輪並設’且具有與前述外齒輪內咬合的剛性，其特徵爲：前述外齒輪分別與前述第1內齒輪及前述第2內齒輪咬合的部份之齒形相同，前述外齒輪、第1內齒輪及第2內齒輪分別具備如該外齒輪與第1內齒輪之同時咬合數及該外齒輪與第2內齒輪之同時咬合數均成爲2以上的齒形。 2-如申請專利範圍第1項記載之撓性咬合式齒輪裝置，其中，前述外齒輪、第1內齒輪或第2內齒輪之齒形爲依據次擺線曲線之形狀。 3.如申請專利範圍第1或2項記載之撓性咬合式齒輪裝置，其中，前述外齒輪之外齒成爲圓筒形銷。 4-如申請專利範圍第3項所述之撓性咬合式齒輪裝置，其中，前述外齒輪、第1內齒輪及第2內齒輪之齒形是藉由給出在與該第1內齒輪或第2內齒輪咬合時從該外齒輪的咬合半徑之中心亦即偏心軸至前述銷中心之位置爲止的距離與從該偏心軸至節點的距離，亦即從該偏心軸穿過前述震盪體之旋轉軸和該偏心軸的直線與由該外齒輪與該第1內齒輪或第2內齒輪之咬合產生的接觸點之共同$ 線的交點即節點爲止的距離來求出。 -42- 201231842 5 如申請專利範圍第3或4項記載之撓性咬合式齒輪裝置’其中，由前述第1內齒輪之齒數與該外齒輪之齒數之比及該第2內齒輪之齒數與該外齒輪之齒數之比求出的減速比爲1/30以下。 6. 如申請專利範圍第1或2項記載之撓性咬合式齒輪裝置’其中，前述外齒輪、第1內齒輪及第2內齒輪之齒形是藉由給出從前述震盪體之旋轉軸至將該第1內齒輪或第2內齒輪之內齒假想爲圓筒形銷時的該銷中心之位置爲止的距離與從該旋轉軸至節點爲止的距離，亦即從該旋轉軸穿過該旋轉軸和前述外齒輪之偏心軸的直線與由該外齒輪與該第1內齒輪或第2內齒輪之咬合產生的接觸點之共同法線的交點即節點爲止的距離的比來求出。 7. 如申請專利範圍第1至6項中任一項記載之撓性咬合式齒輪裝置，其中，前述外齒輪、第1內齒輪及第2 內齒輪之齒形依據由該第1內齒輪之齒數與該外齒輪之齒數之比及該第2內齒輪之齒數與該外齒輪之齒數之比求出的減速比決定。 8 · —種撓性咬合式齒輪裝置，具備：震盪體；筒形外齒輪，配置於該震盪體之外周並具有依該震盪體之旋轉而撓性變形的可撓性；第1內齒輪，具有該外齒輪內咬合的剛性；及第2內齒輪，軸向上與該第1內齒輪並設，且具有與前述外齒輪內咬合的剛性，其特徵爲：當將該外齒輪之外齒設爲圓筒形銷時或假想爲圓筒形銷時，或者，將該第1內齒輪或第2內齒輪之內齒設爲圓 -43- 201231842 筒形銷時或假想爲回筒形銷時，於穿過前述震盪體之旋轉軸和與前述第1內齒輪或第2內齒輪咬合時前述外齒輪的咬合半徑之中心亦即偏心軸的直線與由該外齒輪與該第1 內齒輪及第2內齒輪之咬合產生的接觸點之各個共同法線之交點亦即節點之間配置該銷中心。 9. 一種撓性咬合式齒輪裝置之齒形決定方法，撓性咬合式齒輪裝置，具備：震盪體；筒形外齒輪，配置於該震盪體之外周並具有依該震盪體之旋轉而撓性變形的可撓性；第1內齒輪，具有該外齒輪內咬合的剛性；及第2內齒輪，軸向上與該第1內齒輪並設，且具有與前述外齒輪內咬合的剛性，其特徵爲，包括：使前述外齒輪分別與前述第1內齒輪及前述第2內齒輪咬合的部份之齒形相同，定義該外齒輪、第1內齒輪及第2內齒輪之齒形的步驟；由各個齒輪的大小和齒數，對定義該外齒輪、第1內齒輪及第2內齒輪之齒形的多數個參數建立關聯的步驟；決定該第1內齒輪和第2內齒輪的各個齒形之齒頂和齒根之修正範圍的步驟；利用前述多數個參數求出該第1內齒輪和第2內齒輪各自的該修正範圍以外的齒形部份，從而求出各自的同時咬合數的步驟；及以該同時咬合數均成爲2以上爲條件決定前述多數個參數’從而決定前述外齒輪 '第丨內齒輪及第2內齒輪之 -44- 201231842 齒形的步驟。 1 0 ·如申請專利範圍第9項記載之撓性咬合式齒輪裝置之齒形決定方法，其中，由次擺線曲線定義前述外齒輪或前述第1內齒輪和第2內齒輪之齒形，前述次擺線曲線由以前述震盪體之旋轉軸爲中心固定的基圓和沿該基圓之圓周不滑動地旋轉的旋轉圓規定。 11. 如申請專利範圍第1 〇項記載之撓性咬合式齒輪裝置之齒形決定方法，其中，前述外齒輪之外齒成爲圓筒形銷，將前述次擺線曲線設爲內次擺線曲線，並使該次擺線曲線平行移動與該銷半徑相應的量來定義前述第1內齒輪和第2內齒輪之齒形。 12. 如申請專利範圍第1 1項記載之撓性咬合式齒輪裝置之齒形決定方法，其中，爲了對前述多數個參數建立關聯，考慮如下距離之比，即與前述第1內齒輪或第2內齒輪咬合時從前述外齒輪的咬合半徑之中心亦即偏心軸至前述銷中心之位置爲止的距離，與從該偏心軸至穿過前述震盪體之旋轉軸和該偏心軸的直線與由該外齒輪與該第1 內齒輪或第2內齒輪之咬合產生的接觸點之共同法線的交點即節點爲止的距離的比。 13. 如申請專利範圍第9項記載之撓性咬合式齒輪裝置之齒形決定方法，其中，將前述第1內齒輪或第2內齒輪之內齒假想爲圓筒形銷，並將前述次擺線曲線設爲外次擺線曲線，並使該次擺線曲線平行移動與前述銷半徑相應的量來定義前述外齒輪之齒形，又，求出該次擺線曲線之 -45- 201231842 包絡線並使該包絡線平行移動與該銷半徑相應的量來定義前述第1內齒輪和第2內齒輪之齒形。 14.如申請專利範圍第11項記載之撓性咬合式齒輪裝置之齒形決定方法，其中，爲了對前述多數個參數建立關聯，考慮如下距離之比，即從前述震盪體之旋轉軸至前述銷中心之位置爲止的距離，與從該旋轉軸至節點的距離之比’前述節點爲穿過該旋轉軸和與前述第1內齒輪或第 2內齒輪咬合時前述外齒輪的咬合半徑之中心亦即偏心軸的直線與由該外齒輪與前述第1內齒輪之咬合產生的接觸點之共同法線之交點。 -46-201231842 VII. Patent application scope 1. A flexible snap-in gear device, comprising: an oscillating body: a cylindrical external gear gear disposed outside the oscillating body and having flexibility flexibly deformed according to the rotation of the oscillating body; a first internal gear having a rigidity that is engaged in the external gear; and a second internal gear that is axially coupled to the first internal gear and having a rigidity that engages with the external gear, wherein the external gear is respectively The external gear, the first internal gear, and the second internal gear respectively have the same number of meshes as the external gear and the first internal gear, respectively, in the same shape as the tooth shape of the first internal gear and the second internal gear. Both the external gear and the second internal gear have a tooth shape of 2 or more at the same time. The flexible snap-in gear device according to the first aspect of the invention, wherein the tooth shape of the external gear, the first internal gear or the second internal gear is a shape according to a trochoid curve. 3. The flexible snap-in gear device according to the first or second aspect of the invention, wherein the external gear of the external gear is a cylindrical pin. [4] The flexible snap-in gear device of claim 3, wherein the external gear, the first internal gear, and the second internal gear are toothed by being given with the first internal gear or The distance from the center of the occlusion radius of the external gear, that is, the eccentric axis to the position of the pin center, and the distance from the eccentric shaft to the node, that is, the eccentric shaft passes through the oscillating body The straight line of the rotating shaft and the eccentric shaft is obtained by the distance from the node which is the intersection of the common line of the contact point generated by the engagement between the external gear and the first internal gear or the second internal gear. The flexible snap-in gear device of claim 3, wherein the ratio of the number of teeth of the first internal gear to the number of teeth of the external gear and the number of teeth of the second internal gear are The reduction ratio of the number of teeth of the external gear is 1/30 or less. 6. The flexible snap-in gear device according to claim 1 or 2, wherein the external gear, the first internal gear, and the second internal gear are toothed by giving a rotation axis from the shock body The distance from the position of the center of the pin when the internal tooth of the first internal gear or the second internal gear is assumed to be a cylindrical pin, and the distance from the rotation axis to the node, that is, the distance from the rotation axis The ratio of the straight line of the rotating shaft and the eccentric shaft of the external gear to the distance from the joint of the common point of the contact point between the external gear and the first internal gear or the second internal gear . 7. The flexible snap-in gear device according to any one of claims 1 to 6, wherein the external gear, the first internal gear, and the second internal gear are shaped according to the first internal gear. The ratio of the number of teeth to the number of teeth of the external gear and the ratio of the number of teeth of the second internal gear to the number of teeth of the external gear are determined by the reduction ratio. 8) a flexible snap-in gear device comprising: an oscillating body; a cylindrical external gear disposed on the outer circumference of the oscillating body and having flexibility flexibly deformed according to the rotation of the oscillating body; the first internal gear, The rigidity of the external gear is engaged; and the second internal gear is axially disposed with the first internal gear, and has rigidity to engage with the external gear, and is characterized by: When it is a cylindrical pin or assuming a cylindrical pin, or when the internal gear of the first internal gear or the second internal gear is a round-43-201231842 cylindrical pin or when it is assumed to be a cylindrical pin a straight line passing through the rotating shaft of the oscillating body and the center of the occlusion radius of the external gear when engaging with the first internal gear or the second internal gear, and the external gear and the first internal gear and The intersection of the respective common normals of the contact points generated by the engagement of the second internal gear, that is, the center of the pin is disposed between the nodes. A method for determining a tooth shape of a flexible snap-in gear device, the flexible snap-in gear device comprising: an oscillating body; a cylindrical external gear disposed outside the oscillating body and having flexibility according to the rotation of the oscillating body a deformable flexible; a first internal gear having a rigidity engaged in the external gear; and a second internal gear axially coupled to the first internal gear and having rigidity engaged with the external gear, characterized in that The method includes the steps of: defining a tooth shape of a portion of the external gear that meshes with the first internal gear and the second internal gear, and defining a tooth profile of the external gear, the first internal gear, and the second internal gear; Determining, by the size and the number of teeth of each gear, a plurality of parameters defining a tooth profile of the external gear, the first internal gear, and the second internal gear; determining respective tooth shapes of the first internal gear and the second internal gear a step of correcting the range of the crest and the root; determining the tooth portions of the first inner gear and the second inner gear that are outside the correction range by using the plurality of parameters, thereby obtaining the respective simultaneous occlusion numbers step; In the same time, a number average nip becomes 2 or more parameter conditions determine the most -44-201231842 the second internal gear of the first gear tooth profile steps Shu 'external to decide the gear'. The tooth shape determining method of the flexible snap gear device according to the ninth aspect of the invention, wherein the outer gear or the tooth shapes of the first inner gear and the second inner gear are defined by a trochoidal curve, The trochoidal curve is defined by a base circle fixed around the rotation axis of the oscillation body and a rotation circle that does not slide along the circumference of the base circle. 11. The method of determining a tooth profile of a flexible snap-in gear device according to the first aspect of the invention, wherein the external gear of the external gear is a cylindrical pin, and the trochoidal curve is set as an inner trochoidal line. The profile of the first internal gear and the second internal gear is defined by a curve that causes the trochoidal curve to move in parallel by an amount corresponding to the radius of the pin. 12. The method for determining a tooth profile of a flexible snap-in gear device according to claim 1 wherein, in order to correlate the plurality of parameters, a ratio of distances to the first internal gear or the first 2 the distance from the center of the occlusion radius of the external gear, that is, the eccentric axis to the position of the pin center, and the straight line from the eccentric shaft to the rotating shaft passing through the oscillating body and the eccentric shaft The ratio of the distance between the external gear and the common point of the contact point of the contact between the first internal gear and the second internal gear, that is, the node. 13. The tooth shape determining method of the flexible snap gear device according to claim 9, wherein the inner tooth of the first internal gear or the second internal gear is assumed to be a cylindrical pin, and the foregoing The cycloid curve is set as the outer trochoid curve, and the trochoid curve is parallelly moved corresponding to the pin radius to define the tooth shape of the external gear, and the curve of the trochoid is obtained -45-201231842 The shape of the first internal gear and the second internal gear is defined by an envelope and moving the envelope in parallel by an amount corresponding to the radius of the pin. 14. The method for determining a tooth profile of a flexible snap-in gear device according to claim 11, wherein in order to correlate the plurality of parameters, a ratio of distances from the rotation axis of the oscillation body to the aforementioned The ratio of the distance from the position of the pin center to the distance from the rotation axis to the node is the center of the occlusion radius of the external gear when passing through the rotation axis and the first internal gear or the second internal gear That is, the intersection of the straight line of the eccentric shaft and the common normal of the contact point generated by the engagement of the external gear and the first internal gear. -46-