KR100905238B1 - Autotensioner - Google Patents

Autotensioner Download PDF

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Publication number
KR100905238B1
KR100905238B1 KR20020005438A KR20020005438A KR100905238B1 KR 100905238 B1 KR100905238 B1 KR 100905238B1 KR 20020005438 A KR20020005438 A KR 20020005438A KR 20020005438 A KR20020005438 A KR 20020005438A KR 100905238 B1 KR100905238 B1 KR 100905238B1
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KR
South Korea
Prior art keywords
arm
friction member
wall surface
rocking
friction
Prior art date
Application number
KR20020005438A
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Korean (ko)
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KR20030011214A (en
Inventor
아유카와카즈마사
Original Assignee
게이츠 유닛타 아시아 가부시키가이샤
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Priority to JP2001227584 priority Critical
Priority to JPJP-P-2001-00227584 priority
Application filed by 게이츠 유닛타 아시아 가부시키가이샤 filed Critical 게이츠 유닛타 아시아 가부시키가이샤
Publication of KR20030011214A publication Critical patent/KR20030011214A/en
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Publication of KR100905238B1 publication Critical patent/KR100905238B1/en

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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16HGEARING
    • F16H7/00Gearings for conveying rotary motion by endless flexible members
    • F16H7/08Means for varying tension of belts, ropes, or chains
    • F16H7/10Means for varying tension of belts, ropes, or chains by adjusting the axis of a pulley
    • F16H7/12Means for varying tension of belts, ropes, or chains by adjusting the axis of a pulley of an idle pulley
    • F16H7/1209Means for varying tension of belts, ropes, or chains by adjusting the axis of a pulley of an idle pulley with vibration damping means
    • F16H7/1218Means for varying tension of belts, ropes, or chains by adjusting the axis of a pulley of an idle pulley with vibration damping means of the dry friction type
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16HGEARING
    • F16H7/00Gearings for conveying rotary motion by endless flexible members
    • F16H7/08Means for varying tension of belts, ropes, or chains
    • F16H2007/0802Actuators for final output members
    • F16H2007/081Torsion springs
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16HGEARING
    • F16H7/00Gearings for conveying rotary motion by endless flexible members
    • F16H7/08Means for varying tension of belts, ropes, or chains
    • F16H7/0829Means for varying tension of belts, ropes, or chains with vibration damping means
    • F16H2007/084Means for varying tension of belts, ropes, or chains with vibration damping means having vibration damping characteristics dependent on the moving direction of the tensioner
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16HGEARING
    • F16H7/00Gearings for conveying rotary motion by endless flexible members
    • F16H7/08Means for varying tension of belts, ropes, or chains
    • F16H7/0829Means for varying tension of belts, ropes, or chains with vibration damping means
    • F16H7/0831Means for varying tension of belts, ropes, or chains with vibration damping means of the dry friction type

Abstract

The automatic tensioning device includes a cup-shaped fixing member and an arm installed to be rotatable in the opening of the fixing member. An axial bore is formed on the bottom of the fixing member. A rocking wall surface extending to the bottom surface is provided on the arm. A first friction member is provided between the rocking wall surface and the inner wall surface of the opening. The first friction member is engaged between the rocking wall surface and the inner wall surface. A rocking shaft extending to the bottom surface is provided at the center of the cover portion of the arm. The rocking shaft is inserted into the axial bore. A second friction member is provided between the rocking shaft and the axial bore. The first friction member and the second friction member are formed of a synthetic resin mainly composed of PPS, and exhibit a high limit PV value and a small friction coefficient.

Description

Automatic Tensioner {AUTOTENSIONER}

The objects and advantages of the present invention will be more readily understood by the following description based on the accompanying drawings.

1 is a diagram showing a belt system of an automobile engine applied to the automatic tensioning device of the first embodiment.

Fig. 2 is a diagram showing the appearance of the automatic tensioning device of the first embodiment.

3 is a cross-sectional view of the automatic tensioning device of the first embodiment.

4 is a graph illustrating the limit PV value of the first friction member.

Fig. 5 is a sectional view of a tension device manufactured for measuring the damping characteristics of the automatic tension device of the first embodiment.

Fig. 6 is a diagram showing the static hysteresis of the damping force of the automatic tensioning device of the first embodiment.

Fig. 7 is a diagram showing dynamic hysteresis of the damping force of the automatic tensioning device of the first sealing type.

Fig. 8 shows measurement results of changes in the forward rotation damping force and the reverse rotation damping force with respect to the rocking frequency of the arm.

9 is a graph showing the results shown in FIG. 8.                 

Fig. 10 is a sectional view of the automatic tensioning device of the second embodiment.

Fig. 11 is a sectional view of the automatic tensioning device of the third embodiment.

The present invention relates to an automatic tensioning device used in a belt system for transmitting the driving force of an automobile engine to a driven pulley by means of a transmission belt.

Typically, the automatic tensioning device is installed in the driven device to transmit the driving force of the automobile engine to the plurality of mechanisms through the transmission belt, and the automatic tensioning device which reliably transmits the driving force to each of the mechanical devices by applying tension to the transmission belt is It is already known. Such an automatic tensioning device is provided with, for example, a fixed member that can be fixed to the engine, an arm that can swing relative to the fixed member, and a pulley rotatably attached to the arm. For example, a torsion coil spring is fitted to the stationary member for tensioning the transmission belt through the pulley.

In this automatic tensioning device, the arm swings when the transmission belt vibrates, and a load acts between the arm and the fixing member. In order to eliminate this load, dampen the vibration of the belt and prevent damage caused by contact between the arm and the fixing member, for example, a friction member made of synthetic resin is fixed to the arm, and when the arm swings, Sliding with respect to the fixing member. It is known to use a C spring to push the friction member from the interior to a substantially constant pressure for engagement of the friction member. For example, such a configuration is shown in Japanese Patent Laid-Open No. 8-338487.

However, the C spring must be set to the material and shape according to the required pressure. In addition, it is necessary to have a structure for engaging the C spring and the friction member. Therefore, when using a C spring, there is a problem that the structure is complicated and the manufacturing cost increases.

It is therefore an object of the present invention to provide an automatic tensioning device in which a friction member is fixed with a simple structure for generating the necessary damping force without using a C spring.
According to the present invention, there is provided a cup-shaped fixing member, an arm, a pulley, and a first friction member.

The cup-like fastening member has a bottom surface with an opening and an axial bore formed therein. The arm is attached to the opening. The arm has a rocking shaft extending to the bottom surface and inserted into the axial bore so that the arm swings about the rocking shaft. The arm has a stub shaft which is eccentric from the swing shaft and extends in a direction opposite to the swing shaft. The pulley rotates about the stub axis and tensions the transmission belt. The first friction member is provided between an annular wall surface of the fixing member located near the opening and a rocking wall formed on the arm to generate a first frictional resistance force caused by rocking of the arm.

Damping force is generated by a simple configuration in which the friction member is sandwiched between the circumferential wall surface and the rocking wall surface.

The automatic tensioning device may be provided with a second friction member interposed between the axial bore and the swing shaft to generate a second frictional resistance force due to the swing of the arm. By having this second friction member together with the first friction member, the rocking of the arm is attenuated.

Preferably, the first friction member has a friction surface in which the first frictional resistance is generated between the rocking wall surfaces due to rocking of the arm, and the area of the friction surface is at a maximum load acting on the first friction member. It is recommended to set the size accordingly.

The area of the friction surface of the first friction member may be determined by the following formula.

A = {(a + b) / a} × F / P

A is the area of the friction surface of the first friction member, and a is the distance from the first peak position at which the maximum load acts on the second friction member to the second peak position at which the maximum load acts on the first friction member. And b is the distance from the second peak position to the third peak position where the maximum load acts on the pulley, F is the maximum load acting on the pulley, and P is the internal pressure of the first friction member.

Preferably, the first friction member is made of a synthetic resin mainly composed of polyphenyl sulfone, wherein the synthetic resin is substantially 2.0 MPa · m when sliding against the arm at a rate of 0.5 m / sec. Limit PV value exceeding / sec. By making the first friction member with a material having a high limit PV value, it can exhibit sufficient durability against the arm swing.

The rocking wall surface and the annular wall surface face each other in parallel, and the first friction member has a bearing portion formed in a pipe shape between the rocking wall surface and the annular wall surface. Such a first friction member is easy to form.

The swinging wall surface is inclined to face the annular wall surface, and the first friction member may have a bearing portion formed in a taper between the swinging wall surface and the annular wall surface. The first friction member may exhibit high durability against radial load by adjusting the thickness of the shaft member according to the distribution of the load acting on the bearing part.

Preferably, the forward rotational damping force acting on the arm when the arm moves in the first direction in which the transmission belt is loosened acts on the arm when the arm moves in a second direction in which the transmission belt is subjected to tension. Greater than the reverse rotation damping force.

Also preferably, the dynamic damping force acting on the arm is greater than the static damping force acting on the arm. In this case, the dynamic damping force is at least two times the static damping force.

The present invention will be described below in accordance with the embodiment shown in the drawings.

FIG. 1 is a diagram showing a belt system of an automobile engine to which an automatic tension device of the first embodiment is applied, and FIG. 2 is a view showing the appearance of the automatic tension device.

The automatic tensioning device 10 is installed in the belt system shown in FIG. The belt system comprises a drive pulley (11) attached to the output shaft of the engine, an air conditioner pulley (12), a power steering system pulley (13), an alternator pulley (14), an idler pulley 15 and 16, and an automatic tension device or tension device 10 are provided. An endless transmission belt 17 wraps around the pulley. The rotational driving force of the drive pulley 11 is transmitted to the other pulley by the transmission belt 17. This transmission belt 17 is driven clockwise in the figure. The tension device 10 pushes the transmission belt 17 from the outside to apply tension to the transmission belt 17.

As shown in Fig. 2, the tension device 10 has a cup-shaped fixing member 20 fixed to an engine block (not shown). The arm 30 is attached to the fixing member 20 so as to be swingable or swingable, and the pulley 40 is supported to be rotatable by the arm 30. As shown in FIG. 1, the transmission belt 17 is wound around the outer periphery of the pulley 40, and the pulley 40 rotates with the rotation of the transmission belt 17. A torsion coil spring (not shown) is received in the holding member 20, and as a result, the pulley 40 is pushed in the direction of applying tension to the transmission belt 17 by this biasing force. The arm 30 oscillates or moves in the I direction in which the transmission belt 17 is loose, and oscillates or moves in the J direction in which the transmission belt 17 is tensioned.

3 shows a cross section of the tension device 10. The fixing member 20 has an installation portion 21 and a cup 22, which has an opening 26 and a bottom 27. The mounting portion 21 has a mounting hole 211 for fixing the fixing member 20 to the engine block.

The cup 22 has a bearing engagement portion 222 extending from the center of the bottom 27 toward the opening 26. This bearing engagement portion 222 has an axial bore 223.

The arm 30 is provided in the opening part 26, and can rock with respect to the axis of the cup 22 or the axis of the oscillation axis 311 and the oscillation wall surface 312 demonstrated later. The arm 30 has a cover portion 310, a swing shaft 311, and a stub shaft 320. The lid 310 is rotatably supported by the opening 26 through the first friction member 50 described later. The swing shaft 311 extends toward the bottom surface 27, and the stub shaft 320 is eccentric from the swing shaft 311 and extends in the opposite direction to the swing shaft 311.

Two tubular portions extending toward the bottom surface 27 are formed on the cover portion 310. This tubular portion is a rocking shaft 311 and a rocking wall surface 312, and the diameter of the rocking wall surface 312 is larger than the rocking shaft 311.

The oscillating wall surface 312 is inserted into the opening 26 facing the annular inner wall surface 224 of the fixing member 20 which is located close to the opening 26 and is substantially parallel. The first friction member 50 is provided between the inner wall surface 224 and the rocking wall surface 312. The first friction member 50 has a bearing 510 and a flange 520 protruding in the horizontal direction from the outer surface of the bearing 510. The bearing 510 extends along the rocking wall surface 312 and the inner wall surface 224 and has a tubular shape. This bearing 510 functions as a bearing for radial loads. The bearing 510 is engaged between the inner wall surface 224 and the rocking wall surface 312. When the arm 30 swings, frictional resistance is caused between the swinging wall surface 312 and the bearing 510. The flange 520 serves as a thrust bearing to cause smooth rocking of the arm 30.
The swing shaft 311 becomes smaller as the outer diameter goes toward the bottom surface 27. The outer diameter of its tip is smaller than the inner diameter of the axial bore 223. A female screw is formed on the inner wall surface of the tip of the swing shaft 311.

The swing shaft 311 is inserted into the axial bore 223. The tubular second friction member 60 is provided between the oscillation shaft 311 and the axial bore 223. The second friction member 60 is tapered in diameter toward the opening 26. The second friction member 60 has a bearing 61 for supporting a radial load and a flange 62 formed along the bottom surface 221 of the cup 22. When the swing shaft 311 swings with respect to the axis, frictional resistance is caused between the bearing 61 and the swing shaft 311. The movement of the swing shaft 311 in the axial direction is limited by the flange 62.

A disk 24 having a diameter substantially the same as the flange 62 is provided on the underside of the second friction member 60. The coupling bolt 23 is screwed into the tip of the swing shaft 311 through the disk 24.

The torsion coil spring 25 is accommodated in the space defined by the lid 310 and the cup 22. The torsion coil spring 25 is formed by winding a metal material having a predetermined spiral coil length. One end of the torsion coil spring 25 is engaged with the cover portion 310, while the other end is engaged with the bottom surface 221. This torsion coil spring 25 always pushes the arm 30 in the I direction (see Fig. 1).

The arm 30 has a columnar hole 321 formed in the stub shaft 320. The female screw is formed on the inner wall surface of the columnar hole 321. The pulley 40 is rotatably mounted to the stub shaft 320 via a ball bearing 42. The pulley bolt 41 is screwed into the columnar hole 321 so that the pulley 40 is fixed to the stub shaft 320. A dust shield 43 is provided between the pulley bolt 41 and the ball bearing 42.

When the arm 30 swings, the first friction member 50 and the second friction member 60 slide between the fixing member 20 and the arm 30. In other words, the first friction member 50 slides between the inner wall surface 224 and the rocking wall surface 312, and the second friction member 60 slides between the axial bore 223 and the rocking shaft 311. Lose. In other words, the tension device 10 is supported by the first friction member 50 and the second friction member 60 against the swing of the arm 30. The first friction member 50 and the second friction member 60 should be formed to have sufficient durability against sliding with the arm 30. There will be a description of the first friction member 50 and the second friction member 60.

The first friction member 50 is formed using synthetic resin, which is mainly composed of polyphenyl sulfone (PPS) and includes some aromatic nylon (PA-6T) described in Japanese Patent No. 2972561, and Japanese Patent No. 2951321 The polyether sulfone (PES) etc. which were described in these are included.

4 is a graph showing the limit PV value of the first friction member 50 formed of the bearing members J1, J2 and the synthetic resin G formed of a known material. In the figure, the abscissa represents the velocity (m / sec) under the conditions of use while the ordinate represents the PV value (MPa · m / sec).                     

Bearing member J1 is made of PA-6T, while bearing member J2 is made of PES. As can be seen from Fig. 4, when the arm 30 (see Fig. 3) slides with respect to the bearing members J1 and J2 at a speed of substantially 0.5 m / sec, the limit PV value of J1 is approximately 1.6 Mpa.m / sec. On the other hand, the limit PV value of J2 is approximately 2.0 MPa · m / sec. In contrast, when the arm 30 slides with respect to the first friction member 50 under the same conditions, the first friction member 50 exhibits a limit PV value of approximately 4.0 MPa · m / sec. Therefore, the first friction member 50 made of synthetic resin G has a value approximately twice that of the limit PV values of the bearing members J1 and J2, and thus has a relatively high limit PV value.

The first friction member 50 is pressurized by the rocking wall surface 312 by a load acting in a constant direction from the transmission belt 17 (see FIG. 1). In addition, the first friction member 50 slides with respect to the rocking wall surface 312 due to the rocking of the arm 30. If the tension device 10 is used for a long time, the first friction member 50 may be worn due to the pressure and the slip, and the arm 30 may be inclined. In contrast, in the embodiment, durability is improved by using the first friction member 50 made of synthetic resin G having a high limit PV value and a small wear value. Similarly, the second friction member 60 is also made of a material having a small wear value.

Note that the wear value (k) is defined by the following formula:

Δw = k p v t

Where? W is the amount of wear of the friction member, p is the pressure acting on the friction member, v is the relative speed of the friction member relative to the arm 30, and t is the amount of time the friction member slides with the arm 30. to be.

Since the first friction member 50 is subjected to a relatively strong load compared to the second friction member 60, the first friction member 50 is a material having a high limit PV value, that is, a high pressure resistance value in order to exhibit high durability It is made of a material having The load acting on the first friction member 50 and the second friction member 60 may be calculated as described later. The first peak position where the maximum load occurs in the longitudinal direction of the first friction member 50 is denoted by D1, and the second peak position where the maximum load occurs in the longitudinal direction of the second friction member 60 is denoted by D2. The third peak position at which the maximum load occurs on the outer surface 411 of the pulley 40 on which the transmission belt 17 travels is denoted by K.

When the transmission belt 17 vibrates, the load acts on the outer surface 411 of the pulley in a constant direction. At this time, the maximum load acting on the third peak position K is represented by F. When the distance from the second peak position D2 to the first peak position D1 is a and the distance from the first peak position D1 to the third peak position K is b, the load fa acting on the first peak position D1 is expressed by the formula [1]. Is expressed:

fa = {(a + b) / a} × F [1]

Similarly, the load fb acting on the second peak position D2 is represented by the formula [2]:

fb = (b / a) × F [2]

As can be seen from the formulas [1] and [2], as the distance a becomes smaller, the load fa acting on the first friction member 50 and the load fb acting on the second friction member 60 become larger. In other words, the friction surface 51 of the first friction member 50 extends to the bottom surface 27 as shown by the broken line H in Fig. 3, and the first peak position D1 is lowered to the bottom surface 27 (in Fig. 3). By positioning at D'1) it is possible to increase the loads fa and fb. When the load fa and fb increase, the frictional resistance which arises in the 1st friction member 50 and the 2nd friction member 60 also increases, and it is possible to increase the damping force of the tension apparatus 10. FIG. At this time, the first friction member 50 and the second friction member 60 should be formed to sufficiently withstand the loads fa and fb acting on the peak positions D1 and D2.

The area A necessary for the member forming the first friction member 50 to withstand the load fa is expressed by the formula [3], where a value including a margin value for ensuring safety for the pressure value (withstand pressure value) to be endured is obtained. Is P.

A = fa / P [3]

As can be seen from formula [3], the area required for the friction surface 51 becomes smaller as the internal pressure value P increases. In the embodiment, since the first friction member 50 is made of a material having a high internal pressure value P, the required area A can be made relatively small. Also, according to formula [3], the larger the load fa, the larger area A should be made. Therefore, when the first friction member 50 is extended to increase the load fa and the first peak position D1 is moved to the bottom 27 side (that is, D'1), the area A is made large and the first friction member ( It is possible to improve the durability of 50).

By substituting Formula [1] into Formula [3], area A of friction surface 51 is represented by Formula [4]:

A = {(a + b) / a} × F / P [4]

Similarly, the area B of the friction surface of the second friction member 60 is represented by the formula [5]:                     

B = (b / a) × F / P [5]

As described above, by extending the friction surface 51 of the first friction member 50 in the direction of the bottom surface 27, the load fa acting on the first friction member 50 and the second friction member 60. The acting load fb increases. Further, by forming the first friction member 50 and the second friction member 60 to withstand these loads, the frictional resistance generated due to the rocking of the arm 30 becomes larger.

The damping force of the tension device 10 is calculated from the sum of the frictional forces generated by the first friction member 50 and the second friction member 60. That is, the damping force DF of the tension device 10 is expressed by the formula [6]. Here, μ1 is the friction coefficient of the material forming the first friction member 50, while μ2 is the friction coefficient of the material forming the second friction member 60.

DF = μ1 × fa + μ2 × fb [6]

Therefore, by increasing the loads fa and fb acting on the first friction member 50 and the second friction member 60, the tension device 10 exhibits a high damping force. Therefore, the damping force of the tension device 10 can be adjusted by adjusting the loads fa and fb acting on the first friction member 50 and the second friction member 60. At this time, it is also possible to adjust the damping force of the tension device 10 by adjusting the friction coefficients µ1 and µ2 of the first friction member 50 and the second friction member 60. For example, it is possible to change the coefficient of friction by mixing PTFE or other materials with synthetic resin G, which is mainly composed of PPS.

Next, the experimental results regarding the damping force of the tension device 10 are described below.

The tension device 70 shown in Fig. 5 is produced for detecting the damping performance of the tension device 10 in the embodiment. The structure of the tension device 70 is provided with a ball bearing 71 in place of the first friction member 50 (see FIG. 3), and a ball bearing 72 in place of the second friction member 60 (see FIG. 3). This is different from the tension device 10 in that it is provided. The other part of the tension device 70 is the same as the tension device 10.

6 shows a static hysteresis of the damping force of the tension device 10. Solid lines L1 and L2 are related to the tension device 10. The solid line L1 indicates the load acting on the arm 30 when the arm 30 moves or swings in the forward direction in the J direction (see FIG. 1), and the solid line L2 indicates the arm when the arm 30 moves in the reverse direction in the I direction. The load acting on 30 is indicated (see FIG. 1). The solid line L3 is related to the tension device 70 and indicates a load applied to the arm 30 when the arm 30 moves in the forward or reverse direction. The rocking frequency of the arm 30 is 0.02 Hz.

As can be seen from Fig. 6, the load increases linearly when the arm 30 moves in the forward direction, and the load decreases linearly when the arm 30 moves in the reverse direction. In the tension device 70 using the ball bearings 71 and 72, since the frictional force does not substantially act on the arm 30, the damping force acting on the arm 30 is constant regardless of the direction of movement of the arm 30. (See solid line L3). On the contrary, in the tension device 10 of the embodiment, due to the first and second friction members 50 and 60, the forward rotational load (solid line L1) is greater than the reverse rotational load (solid line L2), and the forward rotation damping force ( The absolute value of S1) is greater than the absolute value of the reverse rotation damping force S2. In other words, the static hysteresis of the tension device 10 is inequality.

Figure 7 shows the dynamic hysteresis of the damping force of the tension device. The solid line L4 shows the dynamic characteristics of the tension device 10 of the embodiment. In other words, the solid line L4 indicates the relationship between the angular position of the arm 30 and the load acting on the arm 30 when the arm 30 swings. On the other hand, the solid line L5 indicates the dynamic characteristics of the tension device 70 using the ball bearings 71 and 72. The rocking frequency of the arm 30 is 20 Hz.

As can be understood in FIG. 7, in the tension device 70 using the ball bearings 71 and 72, the damping force acting on the arm 30 is reduced by the friction force does not substantially act on the arm 30. It is constant irrespective of the direction of movement in 30) (see solid line L5). In contrast, in the tension device 10 of the embodiment, the forward rotational load is larger than the reverse rotational load and the hysteresis appears as described above. In other words, the absolute value of the forward rotation damping force S3 which is the difference between the forward rotational load and the load applied to the tension device 70 is the reverse rotation damping force that is the difference between the reverse rotation load and the load applied to the tension device 70. It is larger than the absolute value of (S4). Therefore, the dynamic hysteresis of the tension device 10 has an inequality similar to the static hysteresis.

FIG. 8 shows measurement results of changes in the forward rotational damping force and the reverse rotational damping force with respect to the swing frequency of the arm 30, and FIG. 9 is a graph showing the results shown in FIG. As can be seen from the figure, as the rocking speed or rocking frequency increases, the damping force increases from the slow rocking speed (i.e. 0.02 Hz), and the damping force is approximately constant when the rocking frequency exceeds 10 Hz. do. In other words, the dynamic damping force acting on the arm 30 is greater than the static damping force acting on the arm 30, and when the oscillation frequency is 20 Hz, for example, the dynamic damping force is approximately static damping force. 2.3 times.

In automotive engines, the idle frequency is between 20 and 30 Hz. In the tension device 70 of the embodiment, the dynamic damping force is hardly changed when the rotational speed is changed under the condition of the rotational speed or more. In other words, the speed dependence of the damping force under the use condition in the tension device 10 is small, and even when the engine speed is changed, the tension of the transmission belt is always constant.

As described above, according to the first embodiment, the damping force generated by the tensioning device is kept constant without using the C spring.

A second embodiment will next be described with reference to FIG. The same components as in the first embodiment are given the same reference numerals.

The cover portion 81 of the tension device 80 has a rocking wall surface 82 extending along the direction of the bottom surface 27. The rocking wall surface 82 is inclined to face the inner wall surface 224 adjacent to the opening 26 of the cup 22. In other words, the distance between the oscillation wall surface 82 and the inner wall surface 224 is narrower as it is closer to the bottom surface 27.

The first friction member 90 is provided between the rocking wall surface 82 and the inner wall surface 224. The bearing portion 910 of the first friction member 90 extends along the oscillating wall surface 82 and the inner wall surface 224 and exhibits a tapered shape that becomes narrower in width toward the bottom surface 27. Since the tapered bearing portion 910 is thicker than the tubular bearing 510 of the first embodiment (see Fig. 3), the tapered bearing portion 910 acts on the bearing portion 910 in the radial direction. High durability against load The configuration of the second friction member 60, the fixing member 20, the torsion coil spring 25 and the pulley 40 is similar to that in the first embodiment.

According to the second friction member, it is possible to provide the first friction member 90 to the cup 20 without using a C spring or the like in the same manner as in the first embodiment. Further, according to the second embodiment, it is possible to form the bearing portion 910 having high durability against the radial load acting on the first friction member 90.

A third embodiment will next be described with reference to FIG. The same components as in the first embodiment are given the same reference numerals.

In the tension device 100, the installation part 111 of the fixing member 110 is formed on the outer circumference of the bottom surface 27. A bolt hole 113 recessed toward the opening 26 is formed in the center of the bottom surface 112. The engaging bolt 23 and the disk 24 are provided in this bolt hole 113. The coupling bolt 23 is screwed to the tip of the swing shaft 311 of the cover portion 30 through the disk 24. The second friction member 60 is inserted between the axial bore 223 and the swing shaft 311.

The fixing member 110 is fixed to the engine block while the bottom surface 112 and the engine block (not shown) are joined. The coupling bolt 23 and the disk 24 are provided in the bolt hole 113 and do not interfere with the engine block. The configuration of the first friction member 50, the second friction member 60, the torsion coil spring 25, the arm 30 and the pulley 40 is similar to that of the first embodiment.

According to the third embodiment, the present invention can be applied even when it is necessary to provide the mounting portion 111 on the bottom surface 27 due to the formation of the belt system. Therefore, according to the third embodiment, it is possible to provide the function of the tension device without including the C spring or the like in the configuration including the first friction member 50.

In the first to third embodiments, the materials and shapes of the first friction member 50 and the second friction member 60 are provided based on the formulas [1] to [5], but the friction member ( Other modifications may be used, taking into account the distribution of loads applied to them (50, 60) or their use or other factors.
While embodiments of the present invention have been described with reference to the accompanying drawings, it will be apparent that many improvements and modifications can be made by those skilled in the art without departing from the scope of the present invention.

According to the present invention as described above, it is possible to obtain an automatic tensioning device having excellent vibration damping property exhibiting dynamic hysteresis of unequality by a simple configuration without using a C-type spring.

Claims (10)

  1. A cup-shaped fixing member having an axial bore and having an opening and a bottom;
    An arm having a rocking shaft installed in the opening and extending to the bottom surface and inserted into the axial bore, the arm rocking about the rocking shaft and eccentric from the rocking shaft and being opposite to the rocking shaft; An arm having a stub axis extending into the arm;
    A pulley rotating about the stub axis to tension the transmission belt;
    A first friction member provided between an annular wall surface of the fixing member positioned near the opening and a rocking wall surface formed in the arm such that a first frictional resistance is generated by rocking of the arm; And
    And a second friction member inserted between the axial bore and the swing shaft to produce a second frictional resistance by swinging the arm.
  2. delete
  3. The frictional surface of claim 1, wherein the first friction member has a friction surface which generates the first frictional resistance with the oscillation wall surface by oscillation of the arm, and the area of the friction surface is the maximum acting on the first friction member. Automatic tensioning device characterized in that the size is set according to the load.
  4. The automatic tensioning device according to claim 3, wherein an area of the friction surface of the first friction member is determined by the following equation.
    A = {(a + b) / a} × F / P
    (A is the area of the friction surface of the first friction member, and a is the second peak at which the maximum load acts on the second friction member from the first peak position at which the maximum load acts on the first friction member. The distance to the position, b is the distance from the first peak position to the third peak position at which the maximum load acts on the pulley, F is the maximum load acting on the pulley, and P is the internal pressure of the first friction member. to be.)
  5. The method of claim 1, wherein the first friction member is made of a synthetic resin mainly composed of polyphenyl sulfone, and substantially 2.0 MPa · m / sec when the synthetic resin slides against the arm at a speed of 0.5 m / sec. An automatic tensioning device, characterized by exceeding the limit PV value.
  6. 2. The automatic as claimed in claim 1, wherein the rocking wall surface and the annular wall surface face each other and are substantially parallel, and wherein the first friction member has a first bearing portion formed tubularly between the rocking wall surface and the annular wall surface. Tension device.
  7. The automatic tensioning device according to claim 1, wherein the oscillating wall surface is inclined to face the annular wall surface, and the first friction member has a second bearing portion tapered between the oscillating wall surface and the annular wall surface. .
  8. The forward damping force acting on said arm when said arm moves in a first direction in which said transmission belt is slack, wherein said arm acts on said arm when said arm moves in a second direction in which said transmission belt is tensioned. Automatic tensioning device, characterized in that greater than the acting reverse rotation damping force.
  9. 2. The automatic tensioning device of claim 1 wherein the dynamic damping force acting on the arm is greater than the static damping force acting on the arm.
  10. 10. The automatic tensioning device according to claim 9, wherein the dynamic damping force is at least two times the static damping force.
KR20020005438A 2001-07-27 2002-01-30 Autotensioner KR100905238B1 (en)

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KR101340892B1 (en) 2013-01-21 2013-12-13 배정식 Usb keyboard apparatus and method capable of supporting n-key rollover over 62 keys
KR101374801B1 (en) 2013-01-21 2014-03-13 배정식 Usb keyboard apparatus and method capable of supporting n-key rollover with bios compatibility

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US7588507B2 (en) 2001-04-13 2009-09-15 Unitta Company Thin autotensioner
GB2410310B (en) * 2001-04-13 2005-11-09 Unitta Co Ltd Autotensioner
DE10328900A1 (en) * 2003-06-26 2005-01-13 Ina-Schaeffler Kg Clamping system with a Drechstab as spring means
ES2322054T3 (en) 2004-05-14 2009-06-16 Bayerische Motoren Werke Aktiengesellschaft Torsion bar tensioner of a belt transmission of a vehicle with an improved amortiguation device.
DE102010019054A1 (en) * 2010-05-03 2011-11-03 Schaeffler Technologies Gmbh & Co. Kg jig

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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
KR101340892B1 (en) 2013-01-21 2013-12-13 배정식 Usb keyboard apparatus and method capable of supporting n-key rollover over 62 keys
KR101374801B1 (en) 2013-01-21 2014-03-13 배정식 Usb keyboard apparatus and method capable of supporting n-key rollover with bios compatibility

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KR20030011214A (en) 2003-02-07
GB2377981B (en) 2005-02-09
GB0202267D0 (en) 2002-03-20
GB2377981A (en) 2003-01-29
CA2369327A1 (en) 2003-01-27

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