CN116981503A

CN116981503A - Vibration unit and application thereof in upper and lower limb vibration dynamometer

Info

Publication number: CN116981503A
Application number: CN202280020836.0A
Authority: CN
Inventors: 奥利弗·邓克尔伯格
Original assignee: Swiss Human Brain Intelligent Operations Co ltd
Current assignee: Swiss Human Brain Intelligent Operations Co ltd
Priority date: 2021-03-12
Filing date: 2022-03-03
Publication date: 2023-10-31
Also published as: US20240149107A1; EP4304745A1; AU2022231940A1; WO2022189253A1; KR20230152757A

Abstract

The invention relates to a bicycle load cell comprising at least one pedal device for a user and having a vibration unit, wherein: the vibration unit has at least one spindle (12) which is driven directly or indirectly by a motor (54) and which has an eccentric disc (6) fastened thereto; the eccentric disc (6) is rotatably coupled to the connecting rod (1); and the connecting rod (1) transmits the vibration to the bearing (29) of the pedal device through a connecting rod head (1 a) provided opposite the eccentric disc (6), so that the vibration acts on the bearing (29) substantially only in the vertical direction.

Description

Vibration unit and application thereof in upper and lower limb vibration dynamometer

Technical Field

The present invention relates to a load cell having a vibration unit, to a method of operating such a load cell, to a method of producing such a load cell, and to the use of such a load cell.

Background

In order to be able to positively and effectively influence the individual performance structure of rehabilitation/elderly patients or athletes, it is necessary to convert as many as possible of the external training motivation measures in a manner that balances and adapts to the different structural levels of the human organs. In this process, the training apparatus should be applied in a range of conditions (strength, endurance, speed, flexibility) and coordination (neuromotor) factors.

The variety of vibration training devices provides a new training option for optimizing physiology by reactivating pathologically degenerated functional systems in the human body structure or improving the ability of intact functional systems. Although commercial application of Medical Vibration Training (MVT) has begun, scientific validation of this approach is still in the fundamental research stage.

Devices for delivering vibrational energy to a user are well known in a variety of publications.

In this way US4570927 discloses a device in which the legs of a paraplegic patient are moved and vibrated by means of a crank unit, driven by a motor for example.

NL1021619C describes a device for transmitting vibrational energy to an upper limb by means of a handle.

A device is known from DE10241340A1, in which a vibrator selectively transmits vibrations to an expanded muscle structure.

Another vibration device is also proposed in DE10225323B4, in which stochastic resonance is transmitted to the user by a complex mechanical structure.

DE19639477A1 shows a device with a seat, a handle and a vibration unit, by means of which the feet of a user are impacted.

None of the above-mentioned five devices disclose the use in combination with or as a load cell, for example by means of a brake unit connected to the crankshaft, and the details on how vibrations are generated are also largely not mentioned.

From DE10313524B3 a training device is known in which the point or points of contact with the person to be trained are isolated from vibrations by means of one or more damping elements, which points of contact can be influenced by vibrations, so that all modules for supporting the body part of the user are in a vibrating state.

From WO2006/69988A1 a vibration load cell is known, in which a chassis bearing is fixedly connected to a vibrating plate, which is vibrated by two vibration motors running in opposite directions. The disadvantage is that non-directional vibrations are generated and the amplitude of the vibration plate is reduced depending on the mechanical load on the pedal crank or the adjustment of the dynamometer brake. The connection between the pedal crank and the dynamometer brake can only be achieved with a bicycle chain with a chain tensioner to compensate for the difference in length and position between the undercarriage bearing and the dynamometer. Thus, unpleasant noise is generated and additional fixing measures are required to prevent the chain from jumping out of the front links.

EP2158944A2 describes a vibrating load cell having vibrations of variable amplitude. There is no disclosure of how the vibrations are generated in particular and how the variation of the amplitude is achieved.

EP-a-2008695 relates to a training device comprising a mechanism which is rotated by a user of the training device by means of a drive means which is rotated about a rotation axis, and a vibration means by means of which the drive means can be put into a vibration state, wherein the vibration means comprise an electric motor which is rotated about the rotation axis and which comprises at least one weight which is rotated about the rotation axis by means of the motor, wherein the weight is arranged eccentrically with respect to the rotation axis. The electric motor is freely pivotable about a fulcrum pin extending parallel to the rotational axis of the electric motor, wherein the fulcrum pin is arranged above the electric motor below the rotational axis of the drive means, and the electric motor is pivotably connected thereto by means of a bracket supporting the rotational axis of the drive means, wherein the bracket is connected to the frame of the exercise device by means of a spring means.

WO-A-2019219653 provides A self-driven vibration mechanism which can be fitted to an existing pedal shaft but which operates independently of the existing pedal. Mechanical isolation or mechanical decoupling allows vibrational energy to be transferred to the foot rather than the pedal shaft and bicycle, respectively. In another embodiment, a fully detachable pedal with a self-driven vibration mechanism may replace an existing pedal.

US-se:Sup>A-2011152040 describes se:Sup>A training system for training se:Sup>A body part of se:Sup>A user, comprising se:Sup>A frame for positioning the training system on se:Sup>A surface during use, se:Sup>A bicycle device comprising at least one bicycle element configured to be rotatable about se:Sup>A bicycle axis, and se:Sup>A vibration device for moving the at least one bicycle element into vibration, as well as se:Sup>A method of training se:Sup>A system and se:Sup>A method of use.

US-se:Sup>A-2020054920 provides se:Sup>A training machine with pedals or foot plates by means of which se:Sup>A person can transfer kinetic energy to the training machine during use, wherein the training machine comprises means for vibrating the pedals or foot plates during training.

Disclosure of Invention

All of the aforementioned dynamometer systems are based on the principle of combining the user with a training device for use on a vibrating plate. All the components for supporting the person to be trained apply vibration energy to the body part or the corresponding body section, respectively, which is in contact with the components. This results in a full body vibration (WBV) that to some extent exceeds the occupational health threshold specified by the DIN ISO 2631 standard. Resonance collisions reduce the use time and thus (time-limited) minimize efficiency. The constructive isolation between the features of MVT devices and the unified neuromotor stimulation coordinated intramuscularly, while focusing on the components of conditional forces, results in a broad range of conditionally coordinated multifunctionality of WBV. MVT products of the prior art cover only a partial selective aspect of training therapies; these devices fail to implement the overall training concept. In combination with conservative training devices is mandatory (e.g. in combination with heart devices or assisted mechanical resistance training during the warm-up/relaxation phase).

The object of the present invention is to provide a load cell with a vibration unit, wherein the amplitude and frequency of the vibration is preferably adjustable, the vibration acting in only one direction, wherein the amplitude of the vibration is substantially independent of the load on the vibration unit and a vibration frequency of up to 50Hz (hertz) can be achieved. Another object of the invention is to use the vibrating unit of the invention in an upper and lower limb vibrating dynamometer.

In particular, the invention relates to a load cell, in particular a bicycle load cell, having at least one pedal device for a user and having a vibration unit according to claim 1. The characterizing part of the claims is not disclosed in the prior art mentioned above.

The present invention relates generally to bicycle load cells. However, the concepts described herein may also be used in a similar manner for an upper limb dynamometer, i.e., a hand dynamometer. The invention can also be used in combination of a bicycle and a hand dynamometer in two crank arrangements. If the proposed technique is used in a hand dynamometer, the chassis bearing used is therefore of course not the actual chassis bearing, but a crank bearing for such a hand dynamometer.

According to the invention, such a load cell is characterized in that the vibration unit has at least one spindle which is driven directly or indirectly by a motor and has an eccentric disc fastened thereto, wherein the eccentric disc is rotatably coupled to the connecting rod. The connecting rod transmits the vibration generated by the rotation of the motor and the eccentricity of the eccentric disc to the bearing of the crank or pedal device via a connecting rod head arranged opposite one of the eccentric discs, so that the vibration acts on this bearing substantially only in the vertical direction.

In this way, highly concentrated vibrations are generated on the undercarriage bearing, in which case the direction of such vibrations is also perfectly vertical and accordingly causes as little overall vibrations as possible. The connecting rod and eccentric disc used provide a very stable construction which is easy to control and which can withstand high loads for a long time without problems. Furthermore, this construction can be regarded as being designed to be easy to adjust the amplitude and frequency and to integrate the additional elements described below in a simple manner.

According to a preferred first embodiment, such a load cell is characterized in that the vibration unit is arranged below the bearing, the connecting rod eye being coupled directly to the bearing, preferably forming the bearing shell of the bearing. In this case, the connecting rod is preferably able to support substantially all of the load directed vertically downwards on the bearing alone, without any other guiding means.

In general, the axis of the spindle preferably extends parallel to the axis of the bearing.

The bearing of the pedal device can also be mounted in a vertical linear guide with a linear slide, wherein the linear slide is fixedly connected with the bearing at the top and with the connecting rod eye at the bottom, wherein the axis of the main shaft preferably runs parallel to the axis of the bearing.

In the case of such a load cell, a base plate can preferably also be provided, under which the spindle and the motor are preferably also arranged, and above which the pedal device is arranged, wherein a recess can be provided in the base plate, through which recess the connecting rod passes and through which the connecting rod eye is coupled directly to the bearing.

A further preferred embodiment is characterized in that the brake is preferably arranged substantially in the same horizontal position as the pedal device, the brake being coupled to the pedal device by means of a force transmission element, preferably in the form of a chain, a timing belt or a V-belt. In this case, the bearing of the pedal device is mounted so as to be pivotable about a horizontal rotation axis, which is preferably arranged in a horizontal position of the brake shaft, wherein the rotation axis is preferably arranged in such a way that a pivoting movement of the bearing position takes place substantially only in the vertical direction.

In this context, the rotation shaft base of the bearing may be realized by a substantially fork-shaped structure, in which the fork-shaped ends of the arms are mounted rotatable about the rotation shaft, while the opposite converging arms are connected to the bearing, preferably forming bearing seats for the pedal device bearing in the converging region. In this configuration, it is also possible to provide more struts for stability, either transversely to the axis of the bearing or parallel to the axis of the bearing.

The vibration unit can also be arranged below the brake, preferably above the floor, wherein the coupling of the connecting rod to the bearing is preferably achieved by at least one strut which extends obliquely upwards and connects the connecting rod head directly or indirectly to the bearing, and which preferably also can be rigidly connected to the rotating shaft base.

According to a further preferred embodiment, a further eccentric disc is provided on the main shaft, by means of which a counterweight can be provided to compensate for the vibrations, wherein this further eccentric disc is preferably provided on the main shaft in an eccentric manner opposite to the eccentric disc for the drive link. Thanks to this type of compensation means, it is ensured that the vibrations are accurately supported only where needed, in particular on the undercarriage bearings, and to the desired extent. In other words, the compensating device prevents vibrations from being transmitted to other elements of the load cell, such as the floor, the user's seat or the handle, and also to the extent that these other components are damaged as a result or the device is displaced during use, respectively.

A preferred first embodiment of such a compensating device is characterized in that the further eccentric disc drives a further connecting rod which is rotatably mounted on the further eccentric disc and is coupled to a counterweight whose vibration direction is substantially the same as that of the vibration device on the bearing, but which serves to compensate for the vibrations on the bearing, preferably the vibrations on the counterweight are offset 180 ° from the vibrations on the bearing.

Furthermore, the brake may preferably be arranged substantially at the same horizontal position as the pedal device, the brake being coupled to the pedal device by means of a force transmission element, preferably in the form of a chain, a timing belt or a V-belt, the counterweight being mounted pivotable about a horizontal rotation axis mount, preferably arranged at the horizontal position of the axle of the brake, wherein the rotation axis is preferably arranged such that the counterweight in the bearing area performs a pivoting movement substantially only in the vertical direction, wherein the counterweight in the bearing area preferably has a counterweight head, and wherein the counterweight head also preferably covers the bearing area at least partly in a fork-like shape at the top and bottom.

As an alternative or in addition to such a compensation device with counterweight, vibrations on components that do not actually need to vibrate can also be prevented, since the load cell is mounted on a weight plate, which usually has a weight of at least 50kg (kilograms), preferably more than 100kg, provided for example by sheet metal, sandboxes, water tanks and/or stone elements, which are for example arranged in a frame mounted on a platform, the height of which can be adjusted. Such a frame may preferably be height-adjustable and/or leveled, optionally even electrically adjustable, and moved to a desired position by means of rollers (e.g. the rollers can only be lowered when moved). The plate may additionally contain damping elements; damping elements of this type are preferably arranged at the corners of the frame and/or the weight plates and/or are provided with damping pads supported on the frame or frame element. Damping mats having a fine cellular elastic structure with a closed gas volume, for example based on polyether urethanes with a thickness in the range of 10-30mm (millimeters), are particularly suitable. The mechanical high-pass filter may be provided with a structure that largely prevents vibration of the floor on which the device is mounted and of components of the load cell that do not need to vibrate. The high pass filter effectively filters out vibrations below 25Hz, preferably below 20 Hz.

A further preferred embodiment of such a load cell is characterized in that the eccentric disc and/or the optional further eccentric disc is mounted on the spindle so as to be movable and adjustable in a direction perpendicular to the spindle axis of rotation, wherein such mounting is preferably effected by a door guide in which the at least one adjustment element, when moved along the spindle axis, causes the eccentric disc to move in a direction perpendicular to the spindle axis of rotation. Such eccentric control can be used to control the effective vibration amplitude of the vibration unit and the compensating device. Such control may be achieved by means of further actuators, or may be feedback controlled by a program according to a desired therapeutic or training procedure, optionally in coordination with the vibration frequency.

The adjustment is characterized in that at least one adjustment element is mounted in a recess or through hole of the spindle for adjustable displacement by means of an actuating device, the door in or on the adjustment element adjusting the eccentricity of the eccentric disc by interaction with a slider on the eccentric disc.

The eccentric disc for generating the required vibrations and the further eccentric disc for the counterweight can be mounted on the spindle, or an adjusting element can be provided by means of which the eccentricities of the two eccentric discs are adjusted in a correlated manner so as to be offset by 180 °, or two separate adjusting elements can be provided for each eccentric disc by means of which the eccentricities of the eccentric discs are individually adjusted.

The load cell of this type is preferably configured and operated at a frequency of 1-50Hz, respectively, with vibration amplitudes on the bearings of between 1-10mm, preferably between 3-7mm, these values being understood to be the variables generated by the vibration unit on the bearings of the pedal device. These values are preferably combined in the load range of 50-500W (watts), in particular 100-300W.

The invention also relates to the operation of such a load cell, or to the use of such a load cell as described above in therapy and/or in a shaping therapy, respectively, wherein the frequency of the bearing is preferably adjusted in the range of 5-50Hz, preferably in the range of 7-25Hz, and/or the amplitude is adjusted in the range of 1-10mm, preferably in the range of 3-7 mm.

Further embodiments are set forth in the dependent claims.

Drawings

The preferred embodiments of the present invention will now be described by way of the accompanying drawings, which are for illustrative purposes only and are not limiting. In the drawings:

fig. 1 shows in an exploded view the main elements of a vibration unit for a load cell according to a first embodiment;

fig. 2 shows in a sectional view in fig. 2 a) the vibration unit according to fig. 1, and in fig. 2 b) a detail of part a according to fig. 2 a);

Fig. 3 shows in an exploded view the main elements of a vibration unit for a load cell according to a second embodiment;

fig. 4 shows the vibration unit according to fig. 3 in a sectional view;

fig. 5 shows in an exploded view the main elements of a vibration unit for a load cell according to a third embodiment;

fig. 6 shows the vibration unit according to fig. 5 in a sectional view;

fig. 7 shows a different arrangement of the vibration unit, wherein:

a) Is an embodiment in which the chassis bearing is directly mounted from below by a connecting rod through a swing arm,

b) Is an embodiment in which the chassis bearing is mounted in the linear mounting portion without the swing arm, the vibration unit is coupled to the linear mounting portion from below, and

c) The vibration unit is arranged below the brake, the underframe bearing is arranged through the swing arm, and the balance weight is arranged;

FIG. 8 shows a side view of the embodiment according to FIG. 7 b);

fig. 9 shows a view of the embodiment according to fig. 7c, wherein for the sake of clarity of the individual components, a suspension without counterweight is shown in a) and only a suspension with counterweight is shown in b);

fig. 10 shows a different view of another embodiment with a vibrating unit and counterweight coupled to a swing arm, wherein a right side view is shown in a), a left side view is shown in b), a top view is shown in c), an exploded view is shown in d), a right oblique upper view is shown in e), and a right oblique lower view is shown in f).

Detailed Description

Fig. 1 shows the main elements of the vibration unit in an exploded view. The actual spindle 12 is mounted by means of two bearings 11 and is rotated by means of a motor (not shown). The coupling to the motor may be direct or indirect, for example by means of a V-belt. The motor is preferably a servomotor with an output power of between 300 and 1600W. The spindle 12 has a structure and has a region on the left side 40, in which the spindle 12 is mounted by means of the bearing 11. Two ball bearings 11 are used to mount the spindle 12 to the bearing housing 19 and prevent axial displacement of the spindle 12. A shoulder surface 12a is provided on the right side. The shoulder surface 12a prevents the eccentric disc 6 from being axially displaced, as shown on the right, thereby preventing the entire connecting rod 1 from being axially displaced. The eccentric disc 6 is movably mounted on the sliding surface 12b of the main shaft. The housing 9 is held in a form-fitting manner in the eccentric disk 6 and biases the eccentric disk 6 out of adjustment from the axis of rotation of the spindle 12. The rotational force of the main shaft 12 is transmitted to the eccentric disk 6 via the sliding surface 12b and then to the connecting rod 1 via the housing 9. The eccentric disk 6 is not directly supported on the sliding surface 12b of the spindle, but on the housing 9 located between them, the housing 9 being divided into two parts, as shown in the figure, or being a part. The contact surface 41 on the inner side of the eccentric disk 6 is correspondingly in contact with the outer side of the housing 9, while the contact surface 42 on the inner side of the housing 9 is in contact with the sliding surface 12b of the spindle 12.

The housing 9 is preferably made of a material having friction properties, for example a plastic material having friction properties, such as PTFE (polytetrafluoroethylene), while the main shaft 12 is made of metal, in order to achieve an optimal friction pairing on the sliding surface 12 b.

The eccentric disk 6 is provided in its axial groove 43 with a slide 5, which slide 5 extends obliquely and transversely to the axis of the groove 43 and which determines the deflection of the eccentric disk 6 and thus the travel of the connecting rod 1. The slide 5 spans the recess 43 and is secured by the screw 7. The mounting screw 7 secures the slide 5 not only to the eccentric disk 6 in a force-fitting manner but also in a form-fitting manner. When the connecting rod 1 is mounted, the ball bearings are fixed to the eccentric disk 6 via the bearing ring 3. For this purpose, the ball bearing screws 2 are screwed to the eccentric disk via the bearing ring 3. The other side is provided with a clamping ring 8, and the outer ring of the ball bearing 4 is fixed on the connecting rod 1 in a force fit manner through a screw 10. The screw 10 clamps the ball bearing 4 to the connecting rod 1 via the clamping ring 8.

The force of the connecting rod 1 is transmitted through the eccentric disc 6 via the housing 9 to the main shaft 12 and through the bearing assembly 11 to the bearing housing 19. The connecting rod head 1a is adapted to receive a bearing to movably fix it to a linear unit or a swing arm (see below for details).

The threaded cylindrical adjusting element 13 is displaceably axially engaged in an axial blind bore 38 of the spindle 12. The adjusting element 13 is connected to the bearing block 15 by means of mounting screws 14 in a force-fitting and form-fitting manner. The bearing housing 15 receives a bearing assembly 16 in the form of two ball bearing rings. The trapezoidal thread nut 17 is mechanically coupled (i.e., rotationally fixed) to a bearing housing 19 (not shown in fig. 1) and is mounted on the bearing assembly 16. Also shown in fig. 1 are 6 holes for screws mounted to the bearing housing 19. The bearing assembly 16 is adjustable in the axial direction without play and is fastened to the trapezoidal spindle 18 by means of a spindle clamping nut 20 and a locking ring 21 (neither shown in fig. 1, see fig. 2). The trapezoidal spindle 18 moves the adjusting element 13 in the axial direction to change the stroke of the connecting rod 1. Due to the bearing assembly 16, the trapezoidal shaped spindle 18 does not rotate with the main shaft 12.

The adjusting element 13 is preferably made of a material having friction properties, for example a plastic material having friction properties (such as PTFE), while the slider 5 is made of metal for optimal friction pairing.

A door in the form of a cut-out region 13a extends transversely over the adjusting element. The cut-out region has a width substantially the same as the thickness of the slider 5, but a longer length. When the adjusting element 13 is pushed into the blind hole 38, the cutout region is aligned with the larger opening 39. In other words, the slider 5 passes through the openings 39 and 13a. Thus, the cut-out region 13a is part of the adjustment element 13. The slide 5 is located in the cut-out region 13 a; the deflection of the eccentric disk 6 is effected in a form-fitting manner by the plane of the slide 5 and the cutout region 13a of the adjusting element 13.

In this way, the eccentric disc 6 is eccentrically mounted on the main shaft 12. The lower ring of the connecting rod 1 is rotatably mounted on the eccentric disc 6 by means of the bearing ring 4. When the spindle 12 rotates, the eccentric disc 6 performs an eccentric movement and is transferred to the lower ring of the connecting rod 1, thereby being converted into a translation or oscillation at the connecting rod head 1 a. The frequency of these oscillations is determined by the rotational frequency of the spindle 12 and therefore also by the frequency of the motor driving the spindle. The amplitude of the oscillation can be adjusted by means of a trapezoidal spindle 18. The deeper the adjustment element 13 is pushed into the blind hole 31, the greater the distance of the eccentric disk 6 from the axis of the spindle 12 via the slide 5, and thus the greater the amplitude of the eccentricity and thus of the movement of the connecting rod eye 1 a. Accordingly, the vibration generated on the connecting rod head 1a can be fine-tuned and controlled in terms of frequency and amplitude. Furthermore, the connecting rod has a high mechanical stability and a high directional stability, i.e. the generated vibrations extend completely in the direction of the connecting rod, i.e. the proposed device allows the generation of frequency-adjustable, amplitude-adjustable quasi-unidirectional vibrations in a well-defined direction.

Fig. 2 shows the vibration unit in a) in a schematic cross-sectional view through the axis and in b) details about a in a). It can be seen from the figure how this type of vibration unit is mounted under a base plate 28, which serves as a central fastening seat for the vibration unit. The base plate is provided with a recess 44 from which the connecting rod 1 can freely protrude upwards. On the underside of the base plate 28, a left bearing housing 19 for mounting the spindle is provided, on the one hand, and a right bearing housing 19a for mounting the trapezoidal threaded nut 17 is provided, on the other hand.

The main shaft 12 is mounted in the right bearing housing 19 by means of the above-mentioned bearing 11, wherein a shaft clamping nut 20 is provided which fixedly clamps the bearing assembly 11 for minimizing the axial and radial play of the main shaft 12. Further, a lock ring 21 is provided to prevent the shaft clamping nut 20 from inadvertently loosening.

The bearing assembly 11 in fig. 2 is embodied, for example, as an O-bearing assembly. The forces act outside the bearing assembly 11. The radial and axial play of the spindle 12 is thus adjusted.

The only expected oscillation in the present invention is the deflection of the linkage head 1, which is substantially perpendicular to the floor.

Fig. 3 shows a second exemplary embodiment of a vibration unit in an exploded view, this time with two eccentric discs 6 mounted on the same shaft. In this case, two links 1 with substantially shorter link arms are coupled to the two eccentric discs 6; one of the links is used to generate a practically effective vibration for the user and the other link is used to generate a counter-motion of the counterweight, as will be further explained below. The two eccentric discs 6 are mounted on the same spindle 12, but in this context each eccentric disc 6 on the spindle 12 is provided with a separate sliding surface 12b, and the adjusting element 13 is provided with two correspondingly assigned and obliquely opposed cut-out areas 13a. In principle, however, the two eccentric discs 6 are mounted on the main shaft 12 and the eccentricity of the two eccentric discs is controlled by the adjusting element 13 in a manner similar to that described in the first exemplary embodiment. It is important here that the eccentricity of the two eccentric discs 6 is configured to be phase-shifted by 180 °, which is ensured by the relative inclination of the cutout regions 13a and the corresponding relative inclinations of the two slide blocks 5 of the respective eccentric disc 6. If the adjustment element 13 is moved in the recess 38 of the main shaft 12 by activating the trapezoidal spindle 18, in which case the trapezoidal spindle 18 is fastened by the locking ring 23, which prevents the shaft clamping nut 22 from inadvertently loosening, the shaft clamping nut 22 fixedly clamps the bearing assembly 16 in the bearing housing 15 so that the trapezoidal threaded spindle 18 is mounted axially and radially without play, with one eccentric disc being quasi-offset in a first direction and the other eccentric disc being quasi-offset in a direction opposite to the main axis. This results in a phase shift of 180 ° of the eccentricity of the two eccentric discs 6, in particular in a completely correlated manner, i.e. the adjustment by means of the individual adjusting elements 13 having the cutout regions 13a with opposite inclinations automatically results in a phase shift of exactly 180 °, independently of the amplitude of the vibrations to be adjusted. This structurally ensures that the two links always remain optimally phase shifted, so that the counterweight can be optimally compensated for at any adjustment and at any vibration amplitude.

The second exemplary embodiment also differs from the first exemplary embodiment in that the main shaft 12 is coupled in a slightly different manner. Here, a V-belt pulley 24 is also provided for coupling the servomotor to the main shaft via a V-belt. The V-pulley 24 is secured by a clamping nut, for example in the form of a conical lock sleeve.

Therefore, the second embodiment is different from the first embodiment in that unintended oscillations can be compensated. The term "unintended oscillation" particularly refers to oscillations of the base plate 28 in a direction opposite to the intended oscillation, as well as other oscillations that are not perpendicular to the base plate 28. The unintended oscillations are caused by an unbalanced eccentricity, which is significantly caused by the adjustability of the connecting rod and the structure of the connecting rod, which cannot achieve a static balance due to the amplitude modulation of the stroke.

FIG. 4 shows a second exemplary embodiment in cross-section; it can be seen in particular from the figure how the two links are mounted parallel to each other on the same main shaft 12 by means of two eccentric discs, how the V-pulley 24 for coupling the servomotor is located on the left side and how the trapezoidal spindle for adjusting the eccentricity is located on the right side. It follows that a very compact solution is provided in which two links absorbing high loads are mounted in a stable manner.

The bearing surface of the connecting rod eye 26 in fig. 4 is designed to be larger than the connecting rod eye 27 in order to absorb the larger forces that occur under load during operation, for example under the influence of the weight of the human body.

The adjusting element 13 extends the respective slide for the crank or the compensating weight in opposite directions. The two eccentric discs must be rotated 180 ° relative to each other in the axial direction in order to deflect in opposite directions. This offset arrangement of the eccentric disc 6 can be seen more clearly in fig. 5.

Fig. 5 shows a third exemplary embodiment of a vibration unit in an exploded view, in contrast to the second exemplary embodiment, the eccentricities of the two connecting rods 1 or of a designated eccentric disc can be individually adjusted. For this purpose, the spindle 12 is no longer mounted on one side and the other side is opened for control by the adjusting element 13, but is mounted on both ends by the bearing ring 11, as is shown in particular in the sectional view of fig. 6. The spindle is no longer provided with blind holes but with axial through holes, so that now a single adjusting element 13 for adjusting the eccentricity of each eccentric disc 6 can be inserted from both sides. Thus, there is a trapezoidal spindle 18 on both sides, which controls the designated adjusting element 13 respectively. However, both adjusting elements are likewise provided with cutout regions 13a with a relative inclination, so that in principle the eccentricity can be adjusted separately, but still be phase-shifted by 180 °. In this way it can be ensured that the phase offset is always 180 °, but the vibration amplitudes of the two links can be set differently. In this way, a finer adjustment, in particular with respect to environmental or user parameters, is possible by means of the vibration compensation of the counterweight, so that an optimum compensation effect is always ensured.

Therefore, the third embodiment is different from the second embodiment in that the amplitudes of the two links can be controlled in a mutually independent manner. According to this embodiment, the unintended compensation is performed by oscillation balancing. The essential difference in relation to the embodiment according to fig. 3 and 4 is that the adjusting element 13 is configured in two parts. Both adjustment elements 13 require a separate O-shaped mounting and actuation by a motor. The left-hand trapezoidal mandrel 18 controls deflection of the compensating weights; the right-hand trapezoidal spindle 18 controls deflection of the crankshaft. In the present embodiment, the driving of the spindle 12 takes place centrally between the two connecting rods 1.

The adjustment of the compensation may also be performed manually, but it is also possible to actuate the trapezoid spindle or spindles by means of further actuators. For example, the actuator may be actuated by feedback control, e.g. by a vibration sensor or vibration sensors and a corresponding control unit. Thus, in particular, such control may also be feedback controlled by a self-learning algorithm, such that the vibrations measured by the vibration sensor are minimized where no vibrations are generated (e.g. on the floor) and maximized or well within a desired range where vibrations are generated (e.g. at the chassis bearings).

Fig. 6 is a sectional view of fig. 5 exploded. The two adjusting elements 13 differ in length: the link travel in fig. 6 is zero. For changing the stroke, the right adjusting element 13 is moved to the right, the left adjusting element 13 is likewise moved to the right by the rotation of the trapezoidal spindle 18; as a result, the deflection of the eccentric disc changes, which, as can be seen in fig. 6, is due to the different gap positions of the shells 9 (the gap of the right shell at the top and the gap of the left shell at the bottom).

Fig. 7 now shows various possibilities for arranging such a vibrating unit on a (bicycle) load cell.

A first possibility, as shown in the side views of fig. 7b and 8, is to arrange the vibration unit under the bottom plate 28, with the connecting rod 1 passing vertically upwards through this bottom plate, through a recess in this bottom plate. The chassis bearing 29 of the load cell is selectively mounted on a linear slide 34 for movement in a strictly vertical direction, the linear slide being moved by a linear guide 35 mounted on the base plate. The linear slide 34 is fixedly connected at the top with the ball bearing 29 and coupled at the bottom with the connecting rod head 1 a.

In this way a structure is provided which enables vibrations to be selectively achieved only in a strictly vertical direction and the entire suspension and load of the vibration unit is handled by the front area under the undercarriage bearing. Such a vibration unit may be combined with a conventional brake 30, the brake 30 being coupled by a force transmitting element, such as a chain, belt, timing belt.

In this structure, the vibration unit in the first exemplary embodiment described above, that is, only a single link for vibration on the chassis bearing may be used. However, the vibration unit in the second or third exemplary embodiment may also be used. As shown in fig. 8, in particular a counterweight 36 may be mounted so that a 180 ° phase shift is made in such a housing by means of a further link, so that vibrations are transmitted to the chassis bearing by means of the first link shown in the front part of fig. 8 with the desired frequency and amplitude, but vibrations, for example, which are related to the environment and in particular to the chassis 28, are cancelled in a manner quasi-similar to noise cancellation. In practice, such devices do present significant problems due to artificially generated vibrations. On the one hand, artificially generated vibrations can lead to uncomfortable noise emissions, in particular because the floor or the corresponding leg connected thereto can transmit vibrations to floors, buildings, etc.; on the other hand, however, vibrations of other components, in particular brakes and the like, can also produce uncomfortable noise emissions. Furthermore, vibration problems can also occur because such devices are prone to shifting due to rocking and wandering. Last but not least, the vibrations can cause mechanical damage to the device itself and to other components of the device, as well as to other devices disposed nearby to which the vibrations are transmitted unexpectedly.

The frequency range of such vibrations suitable for the device is typically up to 50Hz. The low frequency of 7-12Hz and the amplitude of 7-10mm have proven to be particularly suitable for neural stimulation, with loads typically ranging around 100W. For example, higher frequencies in the range of 15-25Hz may also be used for athletes, in which case the amplitude of vibration is typically somewhat lower, at most 3-4mm. In this case, a load with a brake output power between 200-300W may be used. In this way, the range of vibrations and amplitudes is critical to the mechanical properties of the other components, and thus the compensation of one or more counterweights is of paramount importance.

The crank bearing in fig. 7b is fastened by means of a slide 34 to a linear bearing 35, wherein the linear bearing 35 is arranged perpendicular to the bottom plate 28. The links are connected to the linear slides to produce movement in a direction that is exactly perpendicular to the floor 28. Also, this structure may be implemented in an oscillation-compensating manner by the second link and the counterweight 36 as a second sliding portion on the linear guide portion.

Fig. 7a shows another possibility of providing such a vibration unit on a load cell. Here too, vibrations extending strictly in the vertical direction are generated by the connecting rod 1 (see arrow). However, the link 1 serves as the only mounting portion of the chassis bearing in the vertical direction, thereby providing a very slim structure. To make this structure possible, a swing arm 32 is now added. The swing arm 32 is a second mounting portion of the undercarriage bearing, generally about the brake axle 45. The swing arm 32 is provided with two arms 46, a first arm 46' and a second arm 46". The two arms engage the shaft 45 at its different ends and pivotally mount the undercarriage bearing 29. Since the axle of the undercarriage bearing 29 and the axle 45 of the brake are arranged substantially at the same level, it is ensured that the swing arm 32 enables mobility of the undercarriage bearing 29 on the undercarriage bearing substantially only in the vertical direction, thereby ensuring strict vertical vibrations. If in such a load cell the brake is arranged closer to the floor or substantially below the undercarriage bearing, the swing arm 32 should not be mounted to the axle of the brake, but rather on a separate axial bearing which is approximately horizontal to the undercarriage bearing, to just ensure that the undercarriage bearing is only capable of generating vibrations in the vertical direction.

In fig. 7a, the center of the connecting rod eye 1a is the same as the center of the crank bearing. The crank bearing is mounted only through the connecting rod and the swing arm. Here, all forces are absorbed by the swing arm, except those in the direction of the link. The adjustable braking force of the brake 30 is transmitted to the crank 33 via the force transmission element 31. The braking effect may be adjusted by suitable means known to those skilled in the art, such as translation between the crankshaft and the brake.

Fig. 7c shows another possibility of providing such a vibration unit on a load cell. The vibration unit is arranged below the brake, while the undercarriage bearing is quasi-free floating. A particularly compact and elegant construction pattern is thus formed. The swing arm 32 is again mounted on the shaft 45 of the brake and the undercarriage bearing 29 is mounted such that it can only move in the vertical direction. In this structure, the undercarriage bearing 29 is supported only in the vertical direction, because the swing arm 32 has a strut directed obliquely downward toward the vibration unit and is coupled to one of the two links of the vibration unit by the link seat 37. In other words, the swing arm 32 comprises means for coupling the vibrations of the vibrating means, ensuring, by means of the geometrical design and the lever used, that the vibrations on the chassis bearings are converted into strictly vertical vibrations, although said vibrations on the means act on the connecting rod in an oblique direction. Referring now particularly to fig. 9a, this configuration is shown with only the swing arm 32 having the strut 46 shown for greater clarity.

Fig. 7c thus shows a variant in which the oscillating drive is not arranged below the ball bearing, but outside the crankshaft region. Therefore, the components below the crankshaft are omitted, thereby achieving a very compact structural mode. On the link seat 37 of the swing arm, a link is movably connected with the swing arm.

In this configuration, the respective weights 36 are advantageously mounted in a very similar manner and actuated by a second link that is 180 ° out of phase. Referring now specifically to fig. 9b, this configuration of the counterweight is shown and the swing arm of the undercarriage bearing is omitted. In this case, the counterweight 36 or the counterweight head 50 of the counterweight is mounted on the shaft 45 of the brake by means of the first strut 47 in a similar manner to a swing arm. On the other side, between the counterweight head 50 and the counterweight's rod seat 37a, there is an additional strut 49, said strut 49 pointing downwards, a third strut 48 mounting the counterweight's rod seat onto the shaft 45 of the brake, to ensure the stability required for the mounting. Thus, the counterweight, in particular its counterweight head 50, can be mounted in an optimal space-saving manner and nevertheless mounted in an extended manner between the two arms 46' and 46″ of the swing arm, and can also provide an optimal compensation effect.

Another exemplary embodiment of a load cell is shown in fig. 10. Parts corresponding to those described above have the same reference numerals. In this exemplary embodiment, the swing arm is designed with a plurality of struts on both sides, in particular with additional vertical struts and horizontal struts. In principle, however, the attachment to the connecting rod 1 is similar to that described further above in fig. 7 and 9. The counterweight is also mounted in a similar manner; the counterweight head 50 is here constructed as a layered body, so that an optional in-situ adjustment of the mass of the counterweight head, i.e. the addition of more layers, is possible. Furthermore, the counterweight head 50 is configured as a fork, so to speak, the arms of which cover the undercarriage bearing 29 at least at the top and bottom portions. In this way, the counterweight can be arranged ideally close to the undercarriage bearing and in the region of the undercarriage bearing, so that vibration compensation takes place in an optimal manner. The counterweight is here mounted by means of a mounting body 47, on which a plurality of struts are likewise arranged, and is also coupled to the vibration unit by means of a counterweight connecting rod mount 37 a. This mounting body penetrates the struts of the swing arm in a manner so that it is mounted in an optimal, space-saving and compact manner.

Also seen in this exemplary embodiment are an actuator 52 and a designated V-belt 51 for adjusting the acme threaded nut and, correspondingly, for adjusting the eccentricity and associated vibration amplitude. Also seen is a motor 54 for driving the spindle 12 and a corresponding V-belt 53.

List of reference numerals

1 connecting rod

1' connecting rod for counterweight

1a connecting rod head

2. Screw bolt

3. Bearing ring

4. Ball bearing

5. Sliding block

6. Eccentric disc

6' eccentric disc for counterweight

7. Mounting screw

8. Clamping ring

9. Shell cover

10. Screw bolt

11. Bearing assembly

12. Main shaft

12a shoulder surface

12b sliding surface

13. Adjusting element

13a incision area

14. Mounting screw

15. Bearing pedestal

16. Bearing assembly

17. Trapezoidal thread nut

18. Trapezoidal mandrel

19. Left bearing shell

19a right bearing housing

20. Shaft clamping nut

21. Lock ring

22. Shaft clamping nut

23. Lock ring

24 V-shaped belt pulley

25. Clamping nut

26' connecting rod end bearing

27' connecting rod end bearing

28. Bottom plate

29. Crank bearing

30. Brake device

31. Force transmission element

32. Swing arm

33. Crank arm

34. Linear sliding part

35. Linear guide

36. Counterweight for vehicle

37. Connecting rod seat swing arm

Counterweight of 37a connecting rod seat

38 12, axial blind hole in the housing

39 radial through holes

40 12, fastening region

41 6 contact surface at 9

42 9 contact surface on 12b

43 6, grooves in

44 28 for 1

Shaft of 45 brake

46 32, strut

46', 46"32 arm

47 to the shaft of the brake

48 weight strut from the shaft of the brake to the weight linkage mount

49 weight strut from weight-bearing rod seat to weight-bearing head

50 counterweight head

51V-belt for starting trapezoidal thread nut

52 motor for actuating trapezoidal thread nut by 51

53V-belt for drive motor of spindle 12

54 are used to drive the motor of the spindle 12.

Claims

1. A load cell, in particular a bicycle load cell, has at least one pedal device for a user and has a vibration unit,

it is characterized in that the method comprises the steps of,

the vibration unit has at least one spindle (12) which is driven directly or indirectly by a motor (54) and which has an eccentric disc (6) fastened thereto, wherein the eccentric disc (6) is rotatably coupled to a connecting rod (1) and

the connecting rod (1) transmits vibrations to the bearing (29) of the pedal device via a connecting rod head (1 a) arranged opposite one of its eccentric discs (6), so that the vibrations act on the bearing (29) substantially only in the vertical direction.

2. Dynamometer according to claim 1, characterized in that the vibration unit is arranged below the bearing (29), the connecting rod head (1 a) being directly coupled with the bearing (29), preferably forming a bearing shell of the bearing (29), the connecting rod (9) alone supporting essentially all loads directed vertically downwards on the bearing (29) without any other guiding means, wherein the axis of the main shaft (12) preferably extends parallel to the axis of the bearing (29).

3. Dynamometer according to any of the preceding claims, characterized in that the bearing (29) of the pedal device is mounted in a vertical linear guide (35) with a linear slide (34), wherein the linear slide (34) is fixedly connected with the bearing (29) at the top and with the connecting rod eye (1 a) at the bottom, wherein the axis of the spindle (12) preferably extends parallel to the axis of the bearing (29).

4. A load cell according to any of the preceding claims 2 and 3, characterized in that a bottom plate (28) is provided, under which the spindle (12) and preferably the motor are arranged, and above which the pedal device is arranged, wherein a recess (44) is provided in the bottom plate (28), through which recess the connecting rod (1) passes and through which recess its connecting rod head (1 a) is directly coupled to the bearing (29).

5. Dynamometer according to any of the preceding claims, characterized in that a brake (30) is preferably arranged substantially in the same horizontal position as the pedal device, the brake (30) being coupled to the pedal device by a force transmission element (31), the force transmission element (31) preferably being in the form of a chain, a timing belt or a V-belt, wherein the bearing (29) of the pedal device is mounted pivotable about a horizontal rotation axis, which is preferably arranged in the horizontal position of the axle (45) of the brake (30), wherein the rotation axis (45) is preferably arranged such that a pivoting movement at the position of the bearing (29) is substantially only allowed in the vertical direction.

6. Dynamometer according to claim 5, characterized in that the rotational axis base of the bearing (29) is defined by a substantially fork-shaped structure, wherein the fork-shaped ends of the arms (46', 46 ") are mounted rotatable around the rotational axis, while opposite converging arms are connected to the bearing (29), preferably forming bearing seats for the bearing (29) of the pedal device in converging areas.

7. Dynamometer according to any of the preceding claims 5 and 6, characterized in that the vibration unit is arranged below the brake (30), preferably above a base plate (28), wherein the coupling of the connecting rod (1) to the bearing (29) is preferably achieved by at least one strut (46) which extends obliquely upwards and connects the connecting rod head (1 a) directly or indirectly to the bearing (29), and wherein the strut (46) is preferably also rigidly connectable to the rotating shaft base.

8. Ergometer according to any of the preceding claims, characterized in that the ergometer is mounted on a base plate which acts as a mechanical high-pass filter for the vibrations generated by the vibration unit and/or that a further eccentric disc (6 ') is provided on the spindle (12), by means of which further eccentric disc (6 ') a counterweight (36) is placed in a state of compensating vibrations, wherein the further eccentric disc (6 ') is preferably arranged on the spindle (12) with an eccentricity opposite to the eccentric disc (6) for driving the connecting rod.

9. Dynamometer according to claim 8, characterized in that the further eccentric disc (6 ') drives a further connecting rod (1'), which further connecting rod (1 ') is rotatably mounted on the further eccentric disc (6') and coupled with a counterweight (36) which is placed in vibration and whose vibration direction is substantially the same as the vibration direction of the vibration means on the bearing (29), but which functions to compensate the vibration on the bearing (29), preferably the vibration on the counterweight (36) is 180 ° offset from the vibration on the bearing (29).

10. Dynamometer according to any of the preceding claims 8 and 9, characterized in that a brake (30) is preferably arranged substantially in the same horizontal position as the pedal device, said brake (30) being coupled to the pedal device by a force transmitting element (31), preferably in the form of a chain, a timing belt or a V-belt, said counterweight (36) being mounted pivotable about a horizontal rotation axis base, preferably arranged in the horizontal position of the axle (45) of the brake (30), wherein the rotation axis (45) is preferably arranged such that the counterweight (36) in the area of the bearing (29) performs the pivoting movement substantially only in the vertical direction, wherein the counterweight (36) in the area of the bearing (29) preferably has a counterweight head (50), and the counterweight head (50) also preferably covers the bearing area at least partly in a fork shape at the top and bottom.

11. Ergometer according to any of the preceding claims, characterized in that the eccentric disc (6) and/or an optionally further eccentric disc (6') is mounted on the spindle (12) so as to be displaceable and adjustable in a direction perpendicular to the axis of rotation of the spindle (12), wherein the mounting is preferably effected by a door guide (5, 13 a) in which the eccentric disc (6) is caused to move in a direction perpendicular to the axis of rotation of the spindle when at least one adjustment element (13) is displaced along the axis of the spindle (12).

12. Dynamometer according to claim 11, characterized in that said at least one adjusting element (13) is mounted in a recess (38) or through hole of said spindle (12) for adjustable displacement by actuating means (18), a door (13 a) in or on said adjusting element adjusting the eccentricity of said eccentric disc (6) by interacting with a slider (5) on said eccentric disc (6).

13. Ergometer according to any of the preceding claims 11 and 12, characterized in that an eccentric disc (6) for generating the required vibrations and a further eccentric disc (6 ') for the counterweight are mounted on the spindle (12), or that an adjusting element (13) is provided, by means of which adjusting element (13) the eccentricity of the two eccentric discs is adjusted in a relevant manner so as to be offset by 180 °, or that two separate adjusting elements are provided for the individual eccentric discs (6, 6 '), by means of which the eccentricity of the eccentric discs (6, 6 ') is individually adjusted.

14. Ergometer according to any of the preceding claims, characterized in that the ergometer is designed for operation at an operating frequency of 1-50Hz and that the amplitude of vibration at the bearing (29) is in the range of 1-10mm, preferably in the range of 3-7mm, the load preferably being in the range of 50-500W, in particular in the range of 100-300W.

15. Use of a load cell according to any of the preceding claims in therapy and/or shaping therapy, wherein the frequency at the bearing (29) is preferably adjusted in the range of 5-50Hz, preferably in the range of 7-25Hz, and/or the amplitude is adjusted in the range of 1-10mm, preferably in the range of 3-7 mm.