Power transmitting apparatus

Abstract: In a power transmitting apparatus for transmitting power from the driving side to the driven side using a fluid joint means, a lock-up rolling bearing clutch includes an inner ring (8), an outer ring (9) and a plurality of inclined rollers (10) interposed between track surfaces (8a,9a) of the outer periphery of the inner ring (8) and the inner periphery of the outer ring (9) is disposed between the driving side and the driven side for performing ON/OFF operations by allowing the track surfaces of the inner ring (8) and the outer ring (9) to move near to each other or away from each other. The inner and outer rares (8,9) may be caused to move towards one another by a piston (7) acting via a ball bearing (23).

Description: The present invention relates to a power transmitting apparatus according to the preamble of patent claim 1. This power transmitting apparatus, such as a fluid coupling, a torque converter or the like for transmitting power from a driving side to a driven side using fluid joint means, can mechanically connect the driving side to the driven side with a clutch mechanism interposed therebetween and ensures that the power transmitting efficiency is improved without any slip loss by equalizing the rotational speed of the input from a prime mover with the rotational speed of a unit to be driven. Power transmitting apparatus such as a fluid coupling, a torque converter or the like for transmitting power from a driving shaft to a driven shaft via a fluid filled in a casing where a pump impeller is connected to a driving shaft and a turbine impeller is connected to a driven shaft is known. When the rotational speed of a unit or a driven shaft is to be controlled by means of a fluid coupling, the rotational speed on the loading side or of the driven shaft can be steplessly changed from a minimum rotational speed to a maximum rotational speed by shifting a scoop tube, i.e. the minimum rotational speed or the maximum rotational speed of the driven shaft can be attained by feeding working oil to the fluid coupling formed by an impeller, a runner and an impeller casing or by interrupting the feeding of the working oil to the fluid coupling. In any case, when the driven shaft is rotated at a maximum rotational speed, the maximum rotational speed of the driven shaft is lower than that of the driving shaft or the input rotational speed of the prime mover such as a motor or an engine usually by about 3 % owing to slippage. This slippage cannot be avoided when the rotational speed is controlled by a fluid coupling. Therefore, the fluid coupling efficiency is reduced owing to slippage by the slip percentage, and the larger the transmitting power, the larger the loss of power. To overcome the aforementioned problem the power transmitting apparatus includes a clutch mechanism which mechanically connects the driving side to the driven side and ensures that slip loss is eliminated by equalizing the rotational speed of the input from a prime mover with the rotational speed of the unit to be driven, resulting in an improved power transmitting efficiency. The Document DE-A-3144194 describes a hydraulic torque converter for transmitting torque by means of a hydraulic fluid and/or a mechanical overrunning clutch. The mechanical clutch has an annular driving clutch member with a conical track on the inner periphery and a driven clutch member with a conical track on the outer periphery, the two tracks being parallel to and facing each other. A roller cage supporting a plurality of shock-absorbing tapered cylindrical rollers of a resilient material is disposed between the two tracks. When the clutch is not in operation, each roller is at a predetermined angle of inclination to the generatrix of a conical surface having the same angle as the conical tracks and passing through the middle of the roller. The driven clutch member can be hydraulically actuated to be urged close to the driving clutch member against the force of a spring, bringing the rollers into pressure contact with the track surfaces. When the driving clutch member is driven in one direction relative to the driven member, the rollers are rotated and increase their angles of inclination to the generatrix, causing the two clutch members to be mechanically connected. When the driving clutch member is driven in the opposite direction, the rollers are rotated in a direction to decrease the angles of inclination. As the angles become zero, a clearance is formed between the rollers and the track surfaces, so that the transmission of torque is interrupted. It is the object of the invention to provide the power transmitting apparatus of the preamble of claim 1 with a clutch mechanism which immediately disconnects the mechanical coupling between the driving side and the driven side when the thrusting means are released from the thrust position. According to the invention this object is achieved with the characterizing features of patent claim 1. The rolling bearing clutch is constructed in such a way that a contact angle at least one of contact portions of each of the rollers with track surfaces of the outer ring and the inner ring as seen in the axial direction of the rollers is set to be larger than the frictional angle corresponding to the static frictional coefficient at the contact portion, whereby when the inner and outer rings are relatively rotated in the locking direction, the rollers can slide at the contact portions with the track surfaces of the inner ring and the outer ring. The inner and outer rings are locked by imparting a thrust force to the inner or outer ring by way of thrusting means while the inner and outer rings are relatively rotated in the locking direction, whereby the rolling bearing clutch is shifted to the ON position. The rollers are allowed to slide at the contact portions on the track surfaces of the inner and outer rings by releasing the thrusting means from the thrust position, whereby the rolling bearing clutch is shifted to the OFF position. When the rolling bearing clutch is shifted to the ON position, the thrusting means may be driven by utilizing the centrifugal pressure of a fluid to displace the outer ring, whereby the track surfaces of the inner ring and the outer ring come close to each other. It should be noted that since the rolling bearing clutch is designed such that the track surfaces of the inner ring and the outer ring are moved away from each other when the thrusting force is released, the mechanical coupling between the driving side and the driven side can be immediately disconnected. The above and other objects, features and advantages of the present invention will become more apparent from the following description when taken in conjunction with the accompanying drawings in which preferred embodiments of the present invention are shown by way of illustrative examples. Fig. 1 is a sectional view which shows an embodiment of a power transmitting apparatus according to the present invention. Fig. 2 is a diagram of the working oil circuit used in the power transmitting apparatus shown in Fig. 1. Figs. 3(a) to 3(c) are illustrative views of a known rolling bearing clutch, wherein Fig. 3(a) is a partially sectioned perspective view of the rolling bearing clutch and Figs. 3(b) and 3(c) are diagrams showing the operation of the rolling bearing clutch. Figs. 4(a) to 4(c) are illustrative views of a rolling bearing clutch shown in Fig. 1, wherein Fig. 4(a) is a schematic sectional view showing the relationship between the inner ring, the outer ring and the rollers, and Figs. 4(b) and 4(c) are diagrams showing the relationship between the rollers and the track surfaces. Figs. 5(a) to 5(c) are illustrative views of track surfaces of the rolling bearing clutch, wherein Fig. 5(a) is a view showing track surfaces of the inner ring and the outer ring, Fig. 5(b) is a view showing the disposition of the rollers relative the center axis of the inner and outer rings, and Fig. 5(c) is a diagram showing the relationship between the track surfaces, a socket angle of the track surface, a twist angle of the rollers and a wedge angle of the rollers. Figs. 6(a) to 6(c) are illustrative views for explaining the wedge angle of rollers, wherein Fig. 6(a) is a view showing a screw thread of a taper screw, Fig. 6(b) is a view showing spirally sectioned inner and outer rings of the rolling bearing clutch, and Fig. 6(c) is a schematic view showing the relationship between the roller and the track surfaces of the inner and outer rings. The present invention will now be described in detail hereinafter with reference to the accompanying drawings which illustrate preferred embodiments thereof. In these embodiments, description will be made on the power transmitting apparatus exemplified by a fluid coupling. Fig. 1 is a sectional view which shows the entire structure of the fluid coupling. In Fig. 1 reference numeral 1 denotes a driving shaft, and driven shaft 4 is arranged adjacent to the driving shaft 1. The driving shaft 1 is rotatably supported in a bearing metal sleeve 13 disposed in a bearing casing 18, whereas the driven shaft 4 is rotatably supported in bearing metal sleeves 15, 16 disposed in a bearing casing 19. The bearing casings 18 and 19 are held in position by a bearing casing retainer 20. An impeller 2 is connected to the driving shaft 1, and a runner 3 is connected to the driven shaft 4. An impeller casing 5 is fixed to the impeller 2, and an outer annulus 24 is fixed to the impeller casing 5. An outer sleeve 25 is attached to the outer annulus 24, and an inner annulus 6 is attached to the outer sleeve 25. A rolling bearing clutch C L is disposed in the space surrounded by the impeller casing 5, the outer annulus 24 and the inner annulus 6. The rolling bearing clutch C L is constituted of an inner ring 8, an outer ring 9, a plurality of cylindrical rollers 10 disposed between the inner and outer rings 8 and 9, and a retainer 12 for holding the plurality of rollers 10 at predetermined positions. The outer ring 9 is connected to the outer sleeve 25 via a ball spline mechanism 26, and the inner ring 8 is connected directly to the runner 3. The ball spline mechanism 26 allows the axial relative movement between the outer ring 9 and the outer sleeve 25 while preventing relative rotational movement therebetween. The cylindrical rollers 10 are arranged in the twisted position between the outer ring 9 and the inner ring 8 at a certain angle relative to the center axis of the inner and outer rings. A space surrounded by the outer sleeve 25 and the inner annulus 6 constitutes a cylinder portion, and a piston 7 is axially displaceably disposed in the cylinder portion. A clutch ON/OFF working oil chamber 30 is defined between the inner surface of the cylinder portion and the piston 7 so that working oil is fed to the working oil chamber 30 via working oil feeding passages 31a, 31b, 31c and 31d, each formed in the bearing casing 19 and the inner annulus 6. The working oil feeding passage 31a is hydraulically connected via a pipe to a hydraulic pump (as will be described later). A plurality of balls 23 are interposed between the piston 7 and the outer ring 9 which allow the relative rotational movement therebetween. In addition, a plurality of compression coil springs 11 are interposed between the outer ring 9 and the outer annulus 24, and the outer ring 9 is normally thrust in the rightward direction by the resilient force of the springs 11. A radial bearing 32 is interposed between the outer annulus 24 and the inner ring 8. When working oil is fed to the working oil chamber 30, a certain centrifugal force is caused in the oil in the working oil chamber 30 depending on the quantity of oil fed to the working oil chamber 30, the rotational speed and so forth, whereby the piston 7 is thrust in the leftward direction and the inner annulus 6 is thrust in the rightward direction as seen in Fig. 1. Consequently, the force is exerted on the piston 7 which serves to thrust the outer ring 9 of the rolling bearing clutch C L in the leftward direction against the resilient force of the springs 11 and pull the inner ring 8 to the right-hand side as seen in Fig. 1. The pulling force is exerted on the inner ring 8 through the inner annulus 6, the outer sleeve 25, the outer annulus 24 and the radial bearing 32. Thus, the track surface of the outer ring 9 comes near to the track surface of the inner ring 8, causing the clutch to be shifted to the ON position (connected). On the contrary, when feeding of the working oil to the working oil chamber 30 is interrupted, the working oil which generates the centrifugal force is discharged outside of the working oil chamber 30 from a nozzle 28 which is communicated with the working oil chamber 30, whereby the rolling bearing clutch C L is brought in the OFF position (disconnected). The thrust force of the driving shaft 1 is borne by a thrust bearing 14, and the thrust force of the driven shaft 4 is borne by a thrust bearing 17. In addition, a plurality of nozzles 21 are disposed along the outer periphery of the impeller casing 5, and a dam 27 for establishing the minimum rotational speed of a unit to be driven is formed on the inner wall surface of the impeller casing 5 on the inside of the position where each nozzle 21 is disposed. A working oil chamber of the fluid coupling impeller is constituted by the impeller 2, the runner 3 and the impeller casing 5. The working oil is fed to the fluid coupling impeller working oil chamber via a working oil feeding passage 18a of the bearing casing 18 and an oil feeding hole 2a of the impeller 2. As shown in Fig. 2, the passage for feeding the working oil to the fluid coupling impeller working oil chamber comprises a passage I 1 which includes an opening/closing control valve f 1 adapted to quickly perform full opening and full closing and a maximum oil flow rate adjusting orifice j, and a passage I 2 which includes an opening/closing control valve f 2 disposed in parallel with the control valve f 1 to bypass the latter and a minimum oil flow rate adjusting orifice h. As an oil pump c is driven, the working oil is fed to the working oil chamber from an oil tank a via a strainer b through the passage I 1 or I 2 , the oil feeding passage 18a of the bearing casing 18 on the driving side and the oil feeding hole 2a of the impeller 2. The working oil in the working oil chamber is discharged to the outside via the nozzles 21. As desired, a hydraulic pressure adjusting relief valve d and an oil cooler e are arranged. A flow rate of working oil fed to the working oil chamber is determined by the maximum oil flow rate adjusting orifice j when the control valve f 1 is fully opened, and a minimum flow rate is determined by the minimum oil flow rate adjusting orifice h when the control valve f 1 is fully closed and the control valve f 2 is fully opened. Thus, the rotational speed of the unit is driven at a maximum rotational speed when the control valve f 1 is fully opened, and it is driven at a minimum rotational speed when the control valve f 1 is fully closed and the control valve f 2 is fully opened. On the other hand, the feeding of the working oil to the clutch ON/OFF working oil chamber 30 is achieved by the oil pump c through a passage I 3 , a control valve k and communication passages 31a to 31d (see Fig. 1). A known rolling bearing clutch is shown in Fig. 3. Specifically, the rolling bearing clutch includes an inner ring 100, an outer ring 101 and a number of rollers 104 interposed between the outer periphery of the inner ring 100 and the inner periphery of the outer ring 101 as shown in Fig. 3(a). The rollers 104 are arranged with a predetermined twist angle β relative to a center axis X of the inner and outer rings 100 and 101 as shown in Fig. 3(b). A track surface 102 of the outer periphery of the inner ring 100 is convergent in the form of a substantially circular conical taper surface, and a track surface 103 of the inner periphery of the outer ring 101 is designed also in the form of a substantially circular conical taper surface corresponding to the track surface 102 of the inner ring 100. More precisely, since the track surfaces 102 and 103 are under a condition that the rollers 104 come into line contact with them, they respectively assume a single hyperboloid contour that is a locus which is scribed by turning the rollers 104 about the center axis X. The track surfaces 102 and 103 of the inner and outer rings 100 and 101 are inclined at a predetermined socket angle  relative to the center axis X. The rollers 104 come into contact with the track surfaces 102 and 103 of the inner and outer rings 100 and 101 with a predetermined twist angle β so that as the inner and outer rings 100 and 101 are rotated, the rollers 104 spirally roll between the inner and outer rings 100 and 101. As shown in Fig. 3(b), when the inner ring 100 is rotated in such a direction that it is squeezed or advanced in the outer ring 101, a clutch component F1 which serves to allow the inner ring 100 and the outer ring 101 to move near to each other in the axial direction due to the rolling of the rollers 104 appears so that the rollers 104 between the inner ring 100 and the outer ring 101 are brought into a locked position, enabling torque to be transmitted. On the other hand, when the inner ring 100 is rotated in the reverse direction as shown in Fig. 3(c), a component F2 which serves to allow the inner and outer rings 100 and 101 to be parted away from each other appears so that the rollers 104 can rotate freely. The known rolling bearing clutch explained above is utilized as a one-way clutch which transmits torque in one direction and freely rotates in the other direction by utilizing the aforementioned characteristics. It should be noted that a plurality of preliminary pressure springs 105 for imparting light preliminary pressure between the inner ring 100 and the outer ring 101 are arranged between the inner ring 100 and the outer ring 101 so as to allow the rollers 104 to come into initial contact with the inner and outer rings 100 and 101. Next, the relationship between the inner ring 8, the outer ring 9 and the rollers 10 in the rolling bearing clutch C L employed for the present invention will be described below in detail. Fig. 4 is a diagrammatic illustration which shows the relationship between the inner ring 8, the outer ring 9 and the rollers 10, wherein Fig. 4(a) is a schematic sectional view which shows the inner ring 8, the outer ring 9 and the rollers 10, and Figs. 4(b) and 4(c) are diagrammatical illustrations which show the relationship between the rollers 10 and the track surfaces. As shown in Figs. 4(a) and 4(b), the rollers 10 are arranged with a predetermined twist angle β relative to the center axis X. A track surface 8a of the outer periphery of the inner ring 8 is contoured in the form of a substantially convergent circular conical taper surface, and also a track surface 9a of the inner periphery of the outer ring 9 is designed in the form of a substantially circular conical surface corresponding to the track surface 8a of the inner ring 8. Here, the track surfaces 8a and 9a of the inner and outer rings 8 and 9 satisfactorily function only when the rollers 10 come into line contact with the track surfaces 8a and 9a, and, therefore, as shown in Figs. 5(a) and 5(b), the track surfaces 8a and 9a are loci of the outer periphery and the inner periphery of the rollers 10 when the rollers 10 are turned about the center axis X, and thus, the track surfaces 8a and 9a respectively exhibit a hyperboloid contour. As shown in Figs. 4(b) and 5(c), the track surfaces 8a and 9a of the inner and outer rings 8 and 9 are inclined with a predetermined socket angle  relative to the center axis X. As shown in Fig. 4(b), the socket angle  is the angle between the center axis X and a line a extending tangentially to a track surface at a roller contact point P(x, y) and through the center axis X of the inner ring 8 and the outer ring 9. The rollers 10 are retained by the retainer 12 in the equally spaced relationship. The twist angle β of each roller 10 is not maintained by the retainer 12 but is automatically maintained by the hyperboloid contour of the track surfaces 8a and 9a of the inner and outer rings 8 and 9. The retainer 12 serves to hold the rollers 10 to prevent them from being scattered away when the inner and outer rings 8 and 9 are disassembled. Next, a squeeze-in angle ψ will be described below with reference to Fig. 6. Fig. 6(a) is a diagrammatic illustration which shows a taper screw. As shown in the drawing, a screw thread 34 of the taper screw 33 is spirally wound around an imaginary cylindrical surface 35 with the predetermined lead angle β, and at the same time, it is located on the outer periphery of a circular cone 36 of which diameter is upwardly increased with the predetermined socket angle  relative to the center axis X. Accordingly, the outer diameter of the screw thread 34 exhibits a wedge-shaped contour which gradually enlarges with a predetermined angle  as it goes upwardly relative to the imaginary cylindrical surface 35. This angle  is called a rolling wedge angle. This rolling wedge angle  is an increasing or decreasing angle which increases or decreases as a male thread and a female thread are twisted in the case of a taper screw. However, when the male thread is screwed into the female thread, a sum of both the rolling wedge angles becomes an actual wedge angle. This wedge angle is defined as a squeeze-in angle ψ. In the case of the rolling bearing clutch C L of the invention, since each roller 10 interposed between the inner ring 8 and the outer ring 9 has a twist angle β, when the inner ring 8 and the outer ring 9 are rotated via the rollers 10, the inner ring is screwed into the outer ring 9 in one direction of rolling of the rollers 10 as if there are screws, and the rollers 10 are squeezed between the track surfaces 8a and 9a of the inner ring 8 and the outer ring 9, whereby entirely the same relationship as in the taper screw 33 is established. Fig. 6(b) shows a schematic view which spirally sections the inner and outer rings 8 and 9 of the rolling bearing clutch 1 in the direction of rolling of the rollers 10. As shown in the drawing, the track surface 8a on the inner ring 8 is inclined in such a way that it is gradually increased in diameter with a rolling wedge angle i in one spiral direction (arrow A), while the track surface 9a of the outer ring 9 is inclined in such a way that it is gradually decreased in diameter with a wedge angle o in the other spiral direction (arrow B). Fig. 6(c) is a schematic view in which the roller 10 interposed between the track surfaces 8a and 9a is viewed in the axial direction. When the inner ring 8 and the outer ring 9 are moved in such a direction that the distance between the track surfaces 8a and 9a of the inner ring 8 and the outer ring 9 is narrowed, the roller 10 is squeezed therebetween. Since the squeezing of the roller 10 is affected by both of the rolling wedge angle i on the inner ring 8 and the rolling wedge angle o on the outer ring 9, the resultant summed angle (i+o) is defined as a squeeze-in angle ψ. In the conventional roller bearing clutch, when the roller 10 is locked as it rolls between the track surfaces 8a and 9a of the inner ring 8 and the outer ring 9, a slippage should not be caused at the contact portions between the roller 10 and the track surfaces 8a and 9a. However, in the present invention, slippage is caused at the contact portions between the roller 10 and the track surfaces 8a and 9a as they rotate in the locking direction in order to prevent the roller 10 from being squeezed therebetween. This condition is established by the fact that at least one of the rolling wedge angles i and o at the contact portions between the roller 10 and the track surfaces 8a and 9a is made larger than the frictional angles λi and λo corresponding to the static frictional coefficients µsi and µso at the contact portions. Here, the frictional angle λ is an angle at which an article starts to slip in the case that the article is placed on a flat surface and then the latter is gradually inclined, where equations of tan λi=µsi and tan λo=µso are established. Since any squeezing-in does not occur at the contact point of the roller 10 when there is a slippage at least at one of the contact portions on the inner ring 8 and the outer ring 9, an inequality of i>λi or o>λo should be established. It of course is obvious that inequalities of i>λi and o>λo may be established. In other words, at least one of two conditions of tan i>µsi and tan o>µso should be established. Fig. 4(c) represents a squeeze-in angle ψ which is defined as a summed angle (i+o) of i and o. Since it is assumed that the inner and outer rings 8 and 9 are made of a same material, the static frictional coefficients at the contact portions between the roller 10 and both the track surfaces 8a and 9a are equal to each other and the rolling contact angles i and o are substantially equal to each other, the squeeze-in angle (ψ) may be determined such that the relationship represented by the inequality of tan (ψ/2)>µs is established. Even though the squeeze-in angle is determined in that way, there would be no practical problem. In the drawing, N denotes a resistant force applied to the roller 10 from the contact surface, F denotes a frictional force and P denoted a synthesized force. However, it is in practice difficult to measure and form such a rolling contact angle  and, therefore, the contact angle  is determined in association with the relationship between the roller twist angle β and the socket angle . The contact angle , the roller twist angle β and the socket angle  have a geometrically constant relationship as shown in Fig. 5(c). From the taper screw model shown in Fig. 6(a), the relationship represented by the following equation is established. tan = sinβ · tan Namely, an equation of Δ=1 tanβ·tan is given from the relationship h=1 tanβ and Δ=h tan. Also, in Fig. 6, l 1 =1/cosβ and tan=Δ/l 1 is given. Thus, an equation of tan =cosβ·tanβ·tan=sinβ·tan is given. Accordingly, it is sufficient if the roller twist angle β and the socket angle  are determined in such a manner as to establish an inequality of sin β · tan >µs. The value µs could vary depending on the material employed for the inner ring 8, the outer ring 9 and the rollers 10, the lubricating condition or the like. In the case that µs is set about to 0.05, it is preferable that the roller twist angle β be set to 21° to 24° and the socket angle  be set to 8° to 10°. When the roller twist angle β is set to 21° or less, there readily occurs a locked state, and when the roller twist angle is set to 25° or more, rolling becomes difficult. When the socket angle  is set to 8° or less, there readily occurs a locked state, and when the socket angle  is in excess of 10°, there readily arises slippage, whereby the stability is deteriorated. When sin β·tan is calculated, in the case of β:24° and :8°, it amounts to about 0.057, in the case of β:21° and :10°, it amounts to about 0.063 and in the case of β:21° and :8°, it amounts to 0.0503. Thus, it is larger than the maximum static frictional coefficient µs and satisfies the sliding conditions. At this time, the squeeze-in angle ψ is represented by about (ψ/2)=3°. On the contrary, in a conventional rolling bearing clutch, while the roller twist angle β is set to 15 to 18°, the socket angle  is used within the range of 4 to 4.5°. When sin β·tan in the conventional case is calculated, in the case of β:15° and :4°, it amounts to about 0.018, in the case of β:18° and :4.5°, it amounts to about 0.024, and in the case of β:18° and :4°, it amounts to about 0.021, thus it is smaller than the maximum static frictional coefficient and provides locking conditions. It of course is obvious that the roller twist angle β and the socket angle  are relative values. Thus, it is sufficient that the roller twist angle β is conventionally set to 15° to 18° and the socket angle  is correspondingly selected so that sin β·tan is set to the range of 0.05 or more. On the contrary, it is sufficient that when the socket angle is set to the range of 4 to 4.5°, sin β·tan lies within the range of 0.05 or more. The present invention is applicable to the case where the roller twist angle β is set to a value other than the above-noted value, i.e., the range of 15° or less, the range of 18 to 21° or the range of 24° or more. In addition, the present invention is applicable to the case that the socket angle  is set to the range of 4° or less, the range of 4.5 to 8° or the range of 10° or more. Although the maximum static frictional coefficient µ is set to about 0.05 in the example, it is possible to adjust the static frictional coefficient µ to assume various values such as 0.1, 1.5 or the like. While the roller twist angle β and the socket angle  are set to conventional angles, sliding conditions can be obtained by changing the maximum static frictional coefficient. Namely, even though the conventional roller twist angle is set to the range of 15 to 18° and the socket angle  is set to the range of 4 to 4.5°, sliding conditions will be established if the maximum static frictional coefficient ps is set to be smaller than 0.02. The static frictional coefficient depends on the relative relationship between the static frictional coefficient µ and the rolling frictional angle  or the squeeze-in angle ψ of the roller 10, and the static frictional coefficient is not restricted to 0.05. As shown in Fig. 5(c), there is a geometrical constant relationship between the roller twist angle β and the socket angle . The following relationship is given from Fig. 5(c). Arithmetical formula 1 tan  = x·tan 2 β F 2 +x 2 ·tan 2 β tan = sinβ·tan = sinβ· x·tan 2 β F 2 +x 2 ·tan 2 β In Fig. 5(c), the contour of the track surfaces 8a, 9a of the inner and outer rings 8, 9 is represented by the following equations. y i 2 /a i 2 - x i 2 /b i 2 = 1 y o 2 /a o 2 - x o 2 /b o 2 = 1 In Fig. 5(c), F 1 and F 2 are focal points of hyperbolic curves representing the track surfaces. Here, since the squeeze-in angle ψ of the roller 10 is a sum of the rolling contact angles i and o, the following formula is derived from the formula (1) and the relationship ψ=i+o. Arithmetical formula 2 ψ = tan -1 (y 1 +y o )Xtan 2 βsinβ y 1 ·y o -(Xtan 2 βsinβ) 2 F : arbitrary value given at the time of track designing y 1 : track radius of inner ring at X (F 1 2 +X 2 ·tan 2 β) 1/2 y o : track radius of outer ring at X (F o 2 +X 2 ·tan 2 β) 1/2 F 1 = F - r F o = F + r r = radius of roller As shown in Fig. 5(c), the track sectional shape is determined by the frictional angle  and the twist angle β. The socket angle  is determined depending on which range of the track section is practically used as a track surface. Namely, it is determined by the X coordinates of the roller contact point P (x0 in the foregoing formula) and a value of preset F. The roller contact point is located at the central position as seen in the axial direction of the roller. Next, operation of the fluid coupling constructed as above will be explained. When the control valve f 1 and a control valve K are closed and the control valve f 2 in Fig. 2 is opened while the prime mover is rotated at a rated rotational speed, working oil is fed at a specific flow rate via a minimum oil flow rate adjusting orifice h to the fluid coupling impeller working oil chamber formed by the impeller 2, the runner 3 and the impeller casing 5, which causes a unit to be driven at a minimum rotational speed. When the control valve f 1 is opened and the control valve k is closed, working oil is fed at a fixed rate defined by a maximum oil flow rate adjusting orifice j to the fluid coupling impeller working oil chamber so that the rotational speed of the unit is driven at a maximum rotational speed while including a slippage caused by the fluid coupling. When the control valve k shown in Fig. 2 is opened while the unit is driven at a maximum rotational speed inclusive of a slippage caused by the fluid coupling, the working oil is fed to the clutch ON/OFF working oil chamber 30 formed by an outer sleeve 25, an inner annulus 6 and a piston 7 so that pressure is caused by the centrifugal force in the working oil contained in the working oil chamber 30. The piston 7 is thus thrust in the leftward direction by the pressure, and the track surfaces of the outer ring 9 and the inner ring 8 move near to each other and the rolling bearing clutch C L is shifted to the ON position. Then, the rotational speed of the unit to be driven becomes a rotational speed coincident with the rotational speed on the driving side without including a slippage caused by the fluid coupling. At this time, the working oil continuously flows out from the nozzles 28 communicated with the working oil chamber 30 disposed in the outer sleeve 25. Thus, when the feeding of the working oil to the working oil chamber 30 is interrupted, the working oil in the working oil chamber 30 is discharged from the nozzles 28 by the centrifugal force and the piston 7 is released from the thrusting force. Thus, the rollers 10 start to slide at the contact portions with the track surfaces 8a and 9a of the inner ring 8 and the outer ring 9 so that only the outer ring 9 is rotated but the inner ring 8 is not rotated, whereby the rolling bearing clutch C L is immediately shifted to the OFF position. It is possible that the rotational speed of the unit is shifted to the minimum rotational speed by performing reverse operations to the aforementioned ones after the rolling bearing clutch C L is shifted to the OFF position. The centrifugal force Px for actuating the rolling bearing clutch C L is represented by the following formula. Px = γω 2 (r 2 -r x 2 )/2g Here, γ denotes a specific weight of the working oil, and ω denotes a rotational angular speed. When the working oil is fed by a flow rate Q o , an oil film having an innermost peripheral radius r x and an outermost peripheral radius r is formed in the working oil chamber 30. The outermost peripheral radius r is determined by the outer peripheral wall of the working oil chamber 30. When the working oil feeding flow rate Q o is increased, r x is decreased, and when Q o is decreased, r x is increased. As is apparent from the above description, it is possible for the necessary centrifugal pressure P x to be established by controlling the working oil feeding flow rate Q o . When the fluid coupling shown in Figs. 1 and 2 is applied to a pump for a descaling apparatus adapted to repeatedly perform high speed operation (load operation) and low speed operation (no load operation) for removing scale on the surface of a steel in a steel plant, operation can be performed at a minimum rotational speed at the time of no load, and operation can be performed at the same rated rotational speed as that of a prime mover without including slip loss owing to the use of fluid coupling at the time of load. Consequently, a large saving in energy can be obtained by simple control. Note: Performance of the present invention involves performance of the invention of the copending application EP 96109321.8 of even date.

Claims: A power transmitting apparatus for transmitting power from a driving side to a driven side, and having a fluid joint means and a rolling bearing clutch (C L ) provided in parallel therebetween for transmitting power from said driving side to said driven side via said fluid joint means and/ or said bearing clutch means, said rolling bearing clutch (C L ) including an inner ring (8), an outer ring (9), and a plurality of rollers (10) interposed between the outer periphery of said inner ring (8) and the inner periphery of said outer ring (9), and being disposed between the driving side and the driven side for performing ON/OFF operations by allowing the track surfaces (8a, 9a) of said inner ring (8) and said outer ring (9) to move near to each other or away from each other, the driving side and the driven side being mechanically connectable to each other not via fluid joint means by shifting said rolling bearing clutch (C L ) to an ON position, wherein said power transmitting apparatus is provided with thrusting means for shifting the track surfaces (8a, 9a) of said inner ring (8) and said outer ring (9) in said rolling bearing clutch (C L ) to move near to each other, and means for biasing the rolling bearing clutch (C L ) to an OFF position, wherein said rolling bearing clutch (C L ) is constructed in such a way that said rollers (10) are inclined with a predetermined twist angle (β) relative to the center axis (X) of said inner and outer rings (8, 9), each of the track surfaces (8a, 9a) of said inner ring (8) and said outer ring (9) exhibits a hyperboloid contour that is a revolving locus scribed when said rollers (10) are turned about the center axis (X), and when said inner and outer rings (8, 9) are relatively rotated in one direction, said rolling bearing clutch (C L ) can be locked while being freely rotatable in the reverse direction, characterized in that a contact angle () of at least one of the contact portions (P) of each of said rollers (10) with the track surfaces (8a, 9a) of said outer ring (9) and said inner ring (8) as seen in the axial direction of said rollers (10) is set to be larger than the frictional angle (λ) corresponding to the static frictional coefficient (µs) at the contact portion (P), whereby when said inner and outer rings (8, 9) are relatively rotated in the locking direction, said rollers (10) can slide at the contact portions (P) with the track surfaces (8a, 9a) of said inner ring (8) and said outer ring (9), and said inner and outer rings (8, 9) are locked by thrusting said inner or outer ring (8, 9) with thrusting means while said inner ring (8) and said outer ring (9) are relatively rotated in said locking direction, whereby said rolling bearing clutch (C L ) is shifted to the ON position, and said rollers (10) are allowed to slide at the contact portions (P) on the track surfaces (8a, 9a) of said inner ring (8) and said outer ring (9) by releasing said thrusting means from the thrust position, whereby said rolling bearing clutch (C L ) is shifted to the OFF position. The power transmitting apparatus as claimed in claim 1, wherein as a condition that said rollers (10) slide at the contact portions (P) on said inner ring (8) and said outer ring (9) while said inner and outer rings (8, 9) are relatively rotated in the locking direction, the relationship represented by an inequality of tan (ψ/2) > µs is established, wherein (ψ) is a squeezing-in angle defined between said rollers (10) and track surfaces (8a, 9a) of said inner and outer rings (8, 9) as seen in the axial direction of said rollers (10), and (µs) is a static frictional coefficient at the contact portion (P). The power transmitting apparatus as claimed in claim 2, wherein a maximum static frictional coeficient (µs) is set to about 0.05. The power transmitting apparatus as claimed in claim 3, wherein a socket angle () is defined by an angle between said center axis (X) and the tangential line of said rollers (10) with the track surface (9a) of said outer ring (9), a twist angle (β) of said rollers (10) is set to the range of 21° to 24° and said socket angle () is set to 8° to 10°. The power transmitting apparatus as claimed in claim 1 wherein said thrusting means comprises a piston (7) that is engageable with said outer ring (9) and means for applying fluid pressure to the pressure receiving surface of said piston (7). The power transmitting apparatus as claimed in claim 5, wherein said means for applying fluid pressure comprises a working oil chamber (30) defined between the pressure receiving surface of said piston (7) and a cylinder portion adapted to receive said piston (7), a working oil feeding portion for feeding working oil to said working oil chamber (30), and a nozzle (28) for discharging the working oil from said working oil chamber (30), said fluid pressure being caused by the centrifugal force applied to the working oil in said working oil chamber (30). The power transmitting apparatus as claimed in claim 5, wherein a plurality of balls (23) are interposed between said outer ring (9) of said rolling bearing clutch (C L ) and said piston (7).

Inventors: EBISUI MASAHIKO, HATTORI KAZUO, KIMURA KATSUMI, OGATA HIROSHI, OTSUKA MICHIO, EBISUI, MASAHIKO, HATTORI, KAZUO, KIMURA, KATSUMI, OGATA, HIROSHI, OTSUKA, MICHIO
Applicants: EBARA CORP, EBARA CORPORATION
Application year: 1996
Main classification: F16D41
IPC codes: F16H45, F16D47, F16D41, F16D33