Stabilization unit and method for stabilizing an orbit

Abstract

The present invention relates to a stabilizing unit (11) for stabilizing a track (2), which applies an impact force (F, F) with an adjustable direction. max ,F red The vibration generator (12) has unbalanced masses (19, 19a to 19g) for forming a vibration force (F, F) on a track grid (4) consisting of flanged wheels (15), sleepers (5) of a track (2) to be stabilized, and rails (6) fixed on the sleepers (5). max ,F red and a pressure roller (16) for transmitting the impact force (F, F), and each flanged wheel (15) is supported rotatably about a wheel axis (17) and has a running surface (23) with a predetermined wheel diameter (d). In this case, the rotation axes (20, 20a, 20b) are max ,F red ) in the horizontal plane of action (14), and the impact force (F, F max ,F red The horizontal working plane (14) of the flanged wheels (15) is located at most 300 mm, in particular at most 260 mm, above the rolling plane (25) of the flanged wheels (15). The low horizontal working plane (14) avoids undesirable tipping moments during the stabilization process.

Description

[Technical Field]

[0001] The present invention relates to a stabilization unit for stabilizing a track, the stabilization unit comprising a vibration generator with mutually parallel oriented rotation axes and an unbalanced mass for generating an impact force with an adjustable direction, flanged wheels, and pressure rollers for transmitting the impact force to a track grid consisting of sleepers of the track to be stabilized and rails fixed on the sleepers, each flanged wheel being supported rotatably about its wheel axis and having a running surface with a predetermined wheel diameter. The invention further relates to a railway vehicle equipped with such a stabilization unit and a method for operating a railway vehicle. [Background technology]

[0002] Ballast track is continuously subjected to loads due to rail traffic and environmental influences. For example, the position of the track grid in the ballast bed changes. The ballast bed itself becomes contaminated over time by wear and by introduced foreign components. Maintenance measures such as tamping or cleaning processes overcome these drawbacks, although they involve temporary demolition of the ballast bed. Even after optimal compaction by a tamping unit, subsequent settlement can occur. To prevent such settlement, track stabilization machines are used.

[0003] The machine can travel on the track and has a stabilizing unit that is clamped to the track rails by unit rollers. A vibration generator located in the stabilizing unit generates vibrations that are transmitted to the track grid. The vibration generator's structural type and dimensional design determine the impact force acting on the track at a given vibration frequency. To generate a static load, the stabilizing unit is supported on the machine frame. The transmitted vibrations cause particles in the granular structure of the ballast bed to become mobile, i.e., to move, resulting in a denser deposit. This optimized ballast compaction improves the track's bearing capacity and lateral resistance to shear.

[0004] Austrian Utility Model No. 16604 discloses an exemplary stabilization unit with variable impact force. In this case, a vibration generator includes multiple rotating unbalanced masses arranged on parallel-oriented axes. These unbalanced masses are driven with a variably adjustable phase shift relative to one another. Depending on the arrangement of the unbalanced masses, the altered phase shift alters both the direction and the strength of the impact force. Summary of the Invention [Problem to be solved by the invention]

[0005] The present invention is based on the object of improving a stabilizing unit of the type mentioned at the beginning so that impact forces act optimally on the track. It is also an object of the present invention to provide a railway vehicle that utilizes the expanded usability of the improved stabilizing unit. It also provides an advantageous method for operating such a railway vehicle. [Means for solving the problem]

[0006] According to the present invention, the above object is achieved by the features of the independent claims 1, 12 and 14. The respective dependent claims indicate advantageous embodiments of the invention.

[0007] In the new stabilization unit, the rotation axis is positioned so that the impact force is generated in a horizontal plane of action that is located at a maximum of 300 mm, and in particular at a maximum of 260 mm, above the rolling plane of the flanged wheels. The low horizontal plane of action prevents undesirable tilting moments during the stabilization process. In use, the rolling plane of the flanged wheels corresponds to the plane developed by the upper rail edge of the track to be stabilized. If the horizontal impact force is located at a maximum of 260 mm above the rolling plane or the plane of the upper rail edge, the saddle support of the sleepers can be reliably eliminated. This also applies to the maximum value of 300 mm, which allows for more free space below the stabilization unit for arranging a visual or optical measuring system.

[0008] Advantageously, the horizontal plane of action is located less than half a wheel diameter above a horizontal plane passing through the wheel axes. In this case, the vibration generators are arranged at a correspondingly low position, with the wheel diameter being large enough to prevent damaging pressure peaks from occurring on the rail surface. The flanged wheels are spaced apart so as to leave structural space for the vibration generators between them. This also applies to the elements of the flared axes that press the flanged wheels against the rail during operation. In conventional stabilization units, the vibration generators are always located in the area above the flanged wheels, which results in a horizontal plane of action with high impact forces. The resulting tilting moment, under strong loads, can lead to saddle support of the sleepers on the ballast bed in the center of the track.

[0009] Advantageously, at least two rotating shafts and / or unbalanced masses are coupled to multiple transmission elements and driven by a single common drive. In this way, a single common drive with optimized controls can be used to drive all rotating shafts or unbalanced masses. The type of coupling determines how impact forces result from the centrifugal forces caused by the unbalanced masses. Preferably, centrifugal forces in a desired plane of action are amplified, while centrifugal forces in other planes of action cancel each other out.

[0010] According to another refinement, at least one unbalanced mass is rotatably supported on each rotation shaft, which can be driven at different angular positions, rotational speeds and rotational directions relative to the unbalanced mass fixed on the rotation shaft, thereby making it possible to adjust the direction and magnitude of the centrifugal force produced.

[0011] Preferably, at least one unbalanced mass is coupled to the associated rotation axis by means of a coupling element that depends on the direction of rotation, such that the unbalanced mass rotates, in particular by 180°, relative to the rotation axis when the direction of rotation changes. With the unbalanced mass fixedly arranged on the rotation axis, two different centrifugal forces are generated depending on the direction of rotation, which allows the stabilization unit to operate at the same vibration frequency and under different impact forces.

[0012] In one development of the above variant, at least one unbalanced mass is coupled to the associated rotating shaft by means of a centrifugal locking device, which locks the unbalanced mass on the associated rotating shaft as soon as a set rotation speed is exceeded, thereby ensuring that undesired directional changes of the unbalanced mass do not occur during continuous operation.

[0013] The advantageous low centre of gravity design has a central rotation axis parallel to the longitudinal axis of the unit, with two lateral rotation axes on either side, resulting in a symmetrical design with different drive variations, which largely avoids damaging tipping moments during operation.

[0014] In one refinement of this design, the unbalanced mass associated with the central axis of rotation has an unbalance that is twice as large as the unbalanced masses associated with each of the side axes of rotation, allowing the impact force to be adjusted continuously away from zero.

[0015] In another refinement, the directly driven unbalanced masses are connected to a common drive unit, and the indirectly driven unbalanced masses are connected to the directly driven unbalanced masses via a recirculating transmission, by means of which the common centrifugal force action for all unbalanced masses can be adjusted.

[0016] In this case, the cage of the recirculating transmission is advantageously rotatably supported and connected to a rotary drive which can be used to rotate the cage, thereby changing the relative angular velocity of the directly driven unbalanced mass with respect to the indirectly driven unbalanced mass.

[0017] Another preferred development of the stabilization unit includes an acceleration sensor that detects the acceleration caused by the vibration generator, whereby the movement of the stabilization unit or the movement of the vibrated track grid is detected in order to estimate the reaction force of the track grid.

[0018] The railway vehicle according to the invention comprises a machine frame capable of running on a track on a rail running mechanism and at least two of the above-mentioned stabilizing units, where a front stabilizing unit with a first height adjustment drive is fixed to the machine frame and a rear stabilizing unit with a second height adjustment drive is fixed to the machine frame, so that the stabilizing units can be driven independently of each other with different loads and different impact forces.

[0019] Advantageously, the vibration generators and the height adjustment drives are driven by a common control device, which is configured to drive each vibration generator and each height adjustment drive separately. Using the common control device, the two stabilization units can be operated in a coordinated manner with each other. For example, synchronized vibrations are applied to the track grid.

[0020] In the method for operating a railway vehicle according to the invention, forward movement is performed along the track to be stabilized, with the front stabilization units being driven with vertical impulse forces and the rear stabilization units being driven with horizontal impulse forces. This operating method typically involves a sinusoidal movement of the rail running gear of the railway vehicle following the preceding lift-off wave of the railway vehicle, so that the passage of the railway vehicle is simulated in the control operation. The railway vehicle according to the invention anticipates this process, thus ensuring that a particularly stable track is maintained.

[0021] In one development of the method, an acceleration sensor of the front stabilization unit is used to detect vertical accelerations and derive the reaction force characteristics of the track grid from this acceleration, specifically, an acceleration proportional to the measured force and a force from a known dynamic excitation are used to determine the corresponding reaction force.

[0022] The invention will now be described by way of example only and with reference to the accompanying drawings, in which: FIG. [Brief explanation of the drawings]

[0023] [Figure 1] FIG. 1 shows a railway vehicle equipped with a stabilization unit. [Figure 2] FIG. 1 is a cross-sectional view showing the track with the stabilization unit. [Figure 3] FIG. 10 shows a flanged wheel and a pressure roller in contact with a rail. [Figure 4] 1A and 1B show plan and cross-sectional views of a stabilization unit with three mechanically coupled rotation axes. [Figure 5] 5A and 5B are diagrams illustrating changes in the rotation direction of the rotation axis of the stabilization unit of FIG. 4. [Figure 6] FIG. 5 shows the stabilization unit of FIG. 4 with maximum horizontal force excitation. [Figure 7] FIG. 5 shows the stabilization unit of FIG. 4 with maximum normal force excitation. [Figure 8] 5 shows the stabilization unit of FIG. 4 during rotation of the circular transmission. [Figure 9] 5 is a schematic diagram showing the stabilization units of FIG. 4 each having differently adjusted impact forces. [Figure 10] 5 is a diagram showing impact force characteristics when the driving state of the stabilization unit in FIG. 4 is different. [Figure 11] FIG. 5 shows the vibration amplitude reduction factor of the stabilization unit of FIG. 4. [Figure 12] FIG. 10 is a diagram illustrating the work load showing the trajectory grid area that is vibrated. [Figure 13] FIG. 1 is a diagram showing a drive shaft having an unbalanced mass. [Figure 14] FIG. 10 is a diagram showing an unbalanced mass equipped with a centrifugal force locking portion. DETAILED DESCRIPTION OF THE INVENTION

[0024] The railway vehicle 1 shown in FIG. 1 is a so-called dynamic track stabilization system for stabilizing a ballasted track 2 following a tamping process. The track 2 includes a ballast bed 3, in which a track grid 4 consisting of sleepers 5 and rails 6 fixed to the sleepers 5 is supported. During continuous forward movement of the railway vehicle 1 in the longitudinal direction 7, the track grid 4 is vibrated and pressed into the track bed 3. The targeted subsidence of the track structure 4 is detected by a visual or optical measuring system 8. The exemplary railway vehicle 1 has a machine frame 9 that is supported on a rail running mechanism 10 and can run over the track 2 to be stabilized. Two stabilization units 11 are movably connected to the machine frame 9. Only one stabilization unit 11 is arranged on other machines.

[0025] 2 shows a cross section of the track 2 with the stabilizing unit 11 during the stabilization process. The stabilizing unit 11 comprises two main components independent of each other: a vibration generator 12 and a pair of height adjustment drives 13 (weighted hydraulic cylinders). The vibration generator 12 generates, in an action plane 14, impact forces F alternately in two mutually opposite directions, which cause vibrations of the stabilizing unit 11. Preferably, the impact forces F act in a horizontal plane. The horizontal action plane 14 is of primary importance for the present invention. However, for the extension movement of the stabilizing unit 11, the action of the impact forces F in the vertical direction is also important. In this case, the action plane 14 is a vertical plane.

[0026] The flanged wheels 15 and pressure rollers 16 transmit vibrations to the track grid 4. Each flanged wheel 15 is rotatably supported about a wheel axis 17 and guided along the inner edge of the rail 6, with the wheel axes 17 lying in a common horizontal plane 18. The pressure rollers 16 press against the rail 6 from the outside. A continuously controllable load A is applied by a height adjustment drive 13.

[0027] Advantageously, the stabilization unit 11 has a self-supporting central section with a vibration generator 12. The vibration generator 12 includes an unbalanced mass 19 supported on a rotation axis 20. Side frames adjoin the central section on each side in the longitudinal direction of the track. The central section is connected to the respective side frames, for example, by screw fasteners on annular flanges. In this case, the flanged wheels 15 and pressure rollers 16 are supported exclusively on the associated side frames. To achieve the expansion axis 21, for example, the flanged wheels 15 associated with the side frames are each connected to a hydraulic drive, which allows for movement along the associated wheel axis 17. There is no consistent, common axis for the front flanged wheels 15 or the rear flanged wheels 15. The lack of a continuous axis provides space for a lower positioning of the central section. This results in a low center of gravity 22 for the entire stabilization unit 11 and a low working plane 14 for the vibration generators 12. Preferably, the center of gravity 22 is located in the horizontal plane of action 14 .

[0028] Each flanged wheel 15 has a wheel diameter d measured at the running surface 23. In use, the running surface 23 of the flanged wheel 15 is in contact with the upper rail edge 24 of the rail 6, which is the highest line of the railhead. The running surface 23 of every flanged wheel 15 is abutted by a lower tangential plane and an upper tangential plane. The lower tangential plane forms the rolling plane 25, within which the point of contact between the running surface 23 of the flanged wheel 15 and the upper rail edge 24 is located in use. According to the invention, the vibration generator 12 is positioned at a depth such that the vertical distance a between the horizontal application plane 14 of the impact force F and the rolling plane 25 is at most 300 mm, in particular at most 260 mm. Tests with a vertical distance a of 250 mm have shown excellent results. Even with unfavorable ballast compaction, no saddle support of the sleepers occurred.

[0029] Advantageously, the horizontal plane of action 14 is located less than half the wheel diameter of each flanged wheel 15, i.e., less than d / 2, above a horizontal plane 18 passing through each wheel axis 17. In this case, the upper tangential plane of the running surface 23 forms an interface with the horizontal plane of action 14 located below. If the horizontal plane of action 14 is located above the wheel axis 17, the further vertical distance b between the plane of action 14 and the horizontal plane 18 is less than half the wheel diameter of each flanged wheel 15, i.e., d / 2. This characteristic is met in any case if the horizontal plane of action 14 is located below the wheel axis 17. This results in two advantages: on the one hand, the horizontal plane of action 14 is located sufficiently low, and on the other hand, the wheel diameter d of the flanged wheels 15 is large enough so that no damaging pressure peaks occur on the rail surface.

[0030] Advantageous features of the vibration generator 12 with a reduced structural height are explained with reference to FIGS. 4 to 11. To achieve low-level force excitation, cylindrical unbalanced masses 19 are arranged, rotating about axes aligned in the longitudinal direction 7. In this case, to achieve a continuously adjustable amplitude of the effective impact force F, the unbalanced masses 19 are divided into three groups in the axial direction. In the illustrated example, most of the unbalanced masses 19a, 19b, 19c, 19d, and 19e are freely rotatably supported on the driven mean rotation shaft 20a or the associated lateral rotation shafts 20b. The remaining unbalanced masses 19f and 19g are fixedly connected to the associated rotation shafts 20a and 20b.

[0031] The unbalanced masses 19a, 19b, and 19c supported on the drive shaft 20a have coupling elements 26 that depend on the direction of rotation, through which the unbalanced masses are connected to the respective drive mechanisms. The drive mechanisms in this case are, for example, cylinder wheels fixedly mounted on the drive shaft 20a and equipped with cutouts corresponding to the cover bolts. By changing the direction of rotation of the drive shaft 20a, the unbalanced masses 19d, 19e, 19f, and 19g on the secondary shaft 20b are rotated 180° relative to their starting positions, while the unbalanced masses 19a, 19b, and 19c on the drive shaft 20a maintain their respective positions. This is illustrated in Figure 5 in three successive phases 27, 28, and 29. In the first phase 27, the unbalanced masses 19a-19g are in their starting positions for horizontal force excitation, although only one unbalanced mass 19b from the central group is shown. In this case, all of the unbalanced masses 19a-19g are on the right side. The second phase 28 shows the impact process and the third phase 29 shows the position of the unbalanced mass upon excitation of the normal force.

[0032] The unbalanced masses 19b, 19d, and 19e in the center group have twice the unbalance (mass multiplied by eccentricity, U=m·e) of the unbalanced masses 19a, 19c, 19f, and 19g in the two edge groups. Additionally, the unbalanced masses 19a, 19b, and 19c on one of the drive shafts 20a have twice the unbalance of the unbalanced masses 19a, 19c, 19f, and 19g on the associated secondary shaft 20b. For example, the unbalanced masses 19a, 19b, and 19c have twice the unbalance of the unbalanced masses 19d, 19e, 19f, and 19g, where the unbalanced mass 19d has the same unbalance as the unbalanced mass 19e, and the unbalanced mass 19f has the same unbalance as the unbalanced mass 19g. This arrangement results in a vibration generator 12 with an infinitely adjustable system-wide impact force amplitude, canceling centrifugal forces in one direction.

[0033] The unbalanced masses 19a-19g are driven by the drive shaft 20a and the recirculating transmission 30. FIG. 4 shows the rotational motion 31 of the drive shaft 20a for horizontal excitation and the rotational motion 32 of the drive shaft 20a for vertical excitation. The rotational motions 31 and 32 of the drive shaft 20a are directly transmitted to the unbalanced masses 19a and 19c on the drive shaft 20a. The rotational motion is then transmitted from the unbalanced mass 19c to the adjacent unbalanced mass 19f via a cylinder wheel, which also drives the unbalanced mass 19g fixedly connected to each secondary shaft 20b. In addition, the rotational motion is further transmitted to the transmission drive shaft 33 of the recirculating transmission 30 via the cylinder wheel of the unbalanced mass 19c of the drive shaft 20a. In normal operation, the differential cage 34 of the recirculating transmission 30 is stationary, thereby transmitting rotational motion at the same rotational speed but in the opposite direction to the unbalanced mass 19e of the secondary shaft 20b via the transmission driven shaft 35. By being coupled to this unbalanced mass 19e, further unbalanced masses 19b and 19d of the central group are driven.

[0034] 4 shows in the top view the differential cage 34 with differential bolts 36, compensating cone wheels 37 and axle cone wheels 38. The central view shows a plan view of unit 11, where a cut line through the central group of unbalanced masses 19b, 19d, 19e gives a cross section of unit 11 in the bottom view.

[0035] In Figure 6, a system of unbalanced masses 19a-19g for exciting the maximum possible horizontal force is shown. On the left, the maximum horizontal impact force F max On the right, the system is rotated by a rotation angle α=90° in an unbalanced position S2, in which case the resulting excitation force F err is zero.

[0036] In Figure 7, the functioning scheme of the system of unbalanced masses 19a-19g for exciting the maximum possible vertical force is shown. The left unbalanced position S3 is the maximum possible vertical force F err On the right, the system is rotated through a rotation angle α=90° in the unbalanced position S4, from which the maximum vertical impact force F max occurs.

[0037] 8 shows two unbalanced positions S5, S6 for the system of unbalanced masses 19a to 19g when the cyclic transmission device 30 is rotated. On the left side, the unbalanced position S5 is shown for the angular difference occurring between the unbalanced masses 19a, 19g and 19c, 19f of the edge group and the unbalanced masses 19b, 19d, 19e of the 90° center group, and on the right side, the unbalanced position S6 (operation without excitation, idling) is shown for the angular difference occurring between the unbalanced masses 19a, 19g and 19c, 19f of the edge group and the unbalanced masses 19b, 19d, 19e of the 180° center group.

[0038] To reduce the amplitude of the effective impact force F of the system, the unbalanced masses 19b, 19d, 19e of the central group are rotated relative to the other unbalanced masses 19a, 19c, 19f, 19g, thereby reducing the amplitude of the excitation force F of the central group. M is the excitation force F of the edge group R , the rotational speed ω is reduced or leveled according to the adjustment (FIGS. 9 and 10). For this purpose, the differential cage 34 of the recirculating transmission 30 is rotated during operation by a rotary drive 39 coupled to the differential cage 34 at an angular speed ω D With the known angular velocity ω of the transmission drive shaft 33, the Willis equation can be used to determine the angular velocity ω of the transmission driven shaft 35 when the transmission 30 rotates in the direction opposite to the direction of rotation of the transmission drive shaft 33, i.e. ω2=1 / i0·(ω1-ω D (1-i0)) In the case of a differential gear, based on the standard transmission ratio i0=-1, ω2=2 ω D -ω1 It can be said that:

[0039] FIG. 9 shows a system of unbalanced masses 19a-19g with the three unbalanced positions S1, S5, and S6 of FIGS. 6-8. In the upper image, the angular difference β between the edge group and the center group is zero. Here, the angular difference β is 90° in the center image and 180° in the lower image. To the right of the group of unbalanced masses 19a-19g, the individual excitation forces F E and the resulting excitation force F err The extent of the power is shown.

[0040] In normal operation, the differential cage 34 is stationary, i.e., ω D = 0, so that ω2 = -ω1. In this case, the transmission driven shaft 35 and the transmission drive shaft 33 have the same angular velocity, but opposite directions of rotation. When the recirculating transmission 30 rotates, the transmission driven shaft 35 moves faster than the transmission drive shaft 33, and in this case the difference is the angular velocity ω of the rotating differential cage 34. D , which corresponds to twice the angular difference β. The torsion between the unbalanced masses 19b, 19d, 19e of the central group and the unbalanced masses 19a, 19c, 19f, 19g of the edge group is caused by the transmission between the shafts 33, 35 of the recursive transmission 30 and the corresponding unbalanced masses 19c, 19e. If the transmission ratio between the shafts 33, 35 of the recursive transmission 30 and the unbalanced masses 19c, 19e is, for example, i=-1 (same angular velocity but opposite direction of rotation), then rotating the differential cage 34 by a given angle β / 2 (still fixed) will result in an angular difference β between the unbalanced masses 19b, 19d, 19e of the central group and the unbalanced masses 19a, 19c, 19f, 19g of the edge group. Due to this angular difference β (phase shift), a reduced amplitude of the effective horizontal impact force F of the system in horizontal operation or the effective vertical impact force F in vertical operation will be obtained.

[0041] Excitation force F of each unbalanced mass 19 E (Centrifugal force) is the mass m, eccentricity e, and angular velocity ω at the center of rotation U is obtained by multiplying with the square of

number

[0042] In horizontal operation, the vertical components that may occur in each unbalanced mass group cancel each other out (for example, the vertical component of unbalanced mass 19a and the vertical component of unbalanced mass 19g cancel each other out), so that the maximum excitation force F of one unbalanced mass group is Rmax ,F Mmax is exactly achieved when all the unbalanced masses 19a to 19g of each group are upright horizontally or vertically. When there is no mutual rotation of the unbalanced mass groups (angle difference β=0, unbalanced position S1 on the left in FIG. 6 and unbalanced position S4 on the right in FIG. 7), the excitation force F of the unbalanced masses 19a, 19g and 19c, 19f of the edge groups is E and the unbalanced masses 19b, 19d, and 19e of the central group are summed, where the excitation force F of the central group is M is the excitation force F of the two edge groups that operate synchronously. R In this state, the system has the maximum impact force F max (maximum possible impact force amplitude).

[0043] When the unbalanced masses 19a to 19g are rotated relative to each other, the maximum excitation force F Rmax ,F Mmax completely overlap at any time (Fig. 9 and Fig. 10), and the system's reduced impact force F redThe unbalanced masses 19a-19g are rotated relative to one another so that, for example, the vibrations generated by the center group precede those generated by the edge groups (phase shift 40). This effect is shown in Figure 10, where the vibrations of two edge groups moving synchronously with the center group are mapped onto the unbalanced masses 19a-19g over two complete revolutions, starting from the horizontal position of the unbalanced masses 19a, 19c, 19f, 19g of the edge groups.

[0044] In this case, all three graphs shown in FIG. 10 show the total excitation force F of the edge group with respect to the rotation angle α. R The properties of the central group excitation force F M and the resulting excitation power F err 9. The characteristics of the total excitation force F of the edge group, represented by the dotted line, are shown in the figure, where each starting position corresponds to one of the three unbalanced positions S1, S5, and S6 shown in Figure 9. R The characteristics of the central group excitation force F M The characteristic of the resulting excitation force F is represented by the dashed line. err The phase shift 40 of the amplitude of the whole system relative to the amplitude of the edge group or the amplitude of the center group corresponds to half the angle difference β, which is 0° in the upper graph, 90° in the middle graph, and 180° in the lower graph.

[0045] The vibration of the whole system results from the superposition of the vibration of the edge group and the vibration of the center group. The horizontal or vertical excitation force F of the center group Mmax The maximum amplitude of the total excitation force F of the two edge groups Rmax (F Rmax =F Mmax =F max / 2), the reduced (horizontal) excitation force F depends on the rotation angle α and the angle difference β. err of,

number

[0046] Maximum impact force F max Reduced impact force F, which depends on (maximum excitation force) red is derived from the above equation using extreme value considerations and depends on the angle difference β between the edge group and the central group:

number

[0047] The stabilizing units 11 according to the invention are preferably operated in pairs, as shown in Figure 1. With two stabilizing units 11 used in succession, multiple combination possibilities for ballast compaction are possible due to the variable excitation direction, i.e. both units 11 operating horizontally, both units 11 operating vertically, or one unit 11 operating vertically and the other unit 11 operating horizontally.

[0048] One of the advantages of the present invention with regard to compaction action is the low center of gravity 22, or low plane of action 14, at which the point of application of the horizontal force excitation is located, which allows for a primarily translational excitation of the track structure 4.

[0049] Up until now, the vertical force for compacting the track ballast has only been excited in the intermediate section and at the edge of the ballast superstructure by the intermediate and forward section compactors. The present invention additionally allows for vertical excitation of the ballast below each sleeper. In this case, to avoid damage (head checking, dent formation), care must simply be taken not to lift the stabilizing unit 11 from the rail head. For reliable operation, the vertical load A is set by the height adjustment drive 13 to a height at which the centrifugal force relief effect of the vibration generator 12 remains limited.

[0050] Since the vertical stiffness below the sleepers 5 is greater than the horizontal stiffness, during vertical motion there is a stronger interaction between the railway vehicle 1, the track grid 4 and the ballast 3. Therefore, in the case of purely vertical excitations the machine parameters must be particularly carefully adjusted to the local conditions, in particular the state of the track ballast, the geometry of the ballast 3 and the existing base.

[0051] For an operating mode in which the front stabilization unit 11 of the two stabilization units 11 is excited vertically in the direction of travel and the rear stabilization unit is excited horizontally, a controlled passing run of a railway vehicle is simulated. During such a controlled passing run, a leading lift-off wave (vertical excitation) and a trailing sinusoidal movement (horizontal excitation) usually occur in front of the railway vehicle. The loads from the compaction process are therefore similar to the subsequent loads from rail traffic, which has a favorable effect on the durability of the leading track condition correction.

[0052] Compacting the track ballast beneath each sleeper 5 with a vertical force excitation, in combination with a subsequent horizontal excitation, results in better compaction results. A further important advantage of this operating method is the ability to monitor compaction. For this purpose, the vertical force excitation is selected to be so small that no compaction action occurs. In this way, a statement can be made about the vertical stiffness without disturbing the ballast structure and thus the track geometry.

[0053] To calculate ballast compaction, acceleration signals are measured in the stabilization unit 11, as described in AU 521481. Since the measured acceleration is proportional to the force and the force is known from the dynamic excitation, the characteristic reaction force from the track structure 4 can be calculated from the difference. To evaluate the compaction results, the characteristic quantity is then derived from the corresponding work graph (if the excitation frequency is constant) or via an impedance function (relative to the dynamic stiffness if the excitation frequency is variable). An example of a work graph is shown in FIG. 12. The horizontal axis shows the vibration distance 41 of the activated track grid area. The vertical axis shows the contact force 42 below the activated sleeper 5. From this work graph, the stiffness (relationship between a given force difference 43 and the measured travel 44 when the track structure 4 is loaded), the damping of the system (curvature of the curve), and the introduced energy 45 (enclosed area) can be estimated. The horizontal dashed curves show the static load 46, the minimum vertical load 47 and the maximum vertical load 48.

[0054] The compaction monitoring is based on a mechanical modeling of the track structure 4. Subsequently, through an optimization process, deterministic parameters of the track structure 4 are derived, which give rise to a measured response at a given excitation with a known power spectral density. The measured quantities calculated in this way have the advantage that, on the one hand, they can be directly interpreted physically, and, on the other hand, they can be used as a basis for planning track maintenance.

[0055] FIG. 13 shows an improved version of the drive shaft 20a with unbalanced masses 19a, 19b, and 19c. The two outer unbalanced masses 19a and 19c are fixedly connected to the drive shaft 20a. The central unbalanced mass 19a is rotatably supported on the drive shaft 20a and is connected to the gear 49 via a coupling element 26 depending on the direction of rotation. In this case, the position of the central unbalanced mass 19b relative to the gear 49 depends on the direction of rotation. In the illustrated position, the coupling element 26, which is formed as a cover bolt, is located in the upper entrainment recess 50 of the unbalanced mass 19b. As soon as the direction of rotation of the gear 49 changes, the gear 49 rotates 180° relative to the unbalanced mass 19b until the cover bolt abuts in the lower entrainment recess 51 of the unbalanced mass 19b.

[0056] Centrifugal locks 52 are provided to prevent undesired directional movement of the unbalanced mass 19b. FIG. 14 shows a side view of the centrifugal locks 52 in detail. A lever 53 is associated with each of the upper and lower entraining recesses 50 and 51. Each lever 53 is rotatably supported at one end by the unbalanced mass 19b. When the unbalanced mass 19b is stationary or at a low rotational speed, each lever 53 is pushed inward by an associated spring 54. In this state, the entraining recesses 50 and 51 are free to accommodate the coupling element 26. As the rotational speed increases, centrifugal force pushes the two levers 53 outward. In this case, one lever 53 engages in a groove 55 of the coupling element 26, thereby locking the position of the unbalanced mass 19b relative to the gear 49. Sensors are provided to monitor the position of each of the unbalanced masses 19b.

[0057] In the illustrated embodiment, the unbalanced mass 19b is coupled via a gear 49 to the two unbalanced masses 19d, 19e of the central group and to the transmission driven shaft 35 of the recursive transmission 30. A further gear 56 is arranged on the drive shaft 20a by a further coupling element 26 that depends on the direction of rotation. When the direction of rotation changes, the further gear 56 rotates 180° relative to the drive shaft 20a and couples the drive shaft 20a to the two secondary shafts 20b and the transmission drive shaft 33 of the recursive transmission 30. These transmission elements 30, 49, 56 thus couple all rotating shafts 20a, 20b and the unbalanced masses 19a-19g, with the drive shaft 20a being connected to a common drive 57.

[0058] For carrying out the method according to the invention, it is advantageous if the stabilization units 11 arranged on the railway vehicle 1 are driven by one common control device 58. In this case, the control device 58 is configured to separately drive the vibration generators 12 and the height adjustment drives 13 of each stabilization unit 11. Preferably, an acceleration sensor 59 arranged on the front stabilization unit 11 transmits acceleration signals to the control device 58, which can then be used to evaluate the reaction force characteristics of the track structure 4.

Claims

1. A stabilization unit (11) for stabilizing an orbit (2), comprising: Impact force (F, F) with adjustable direction ｍａｘ , F ｒｅｄ a vibration generator (12) having unbalanced masses (19, 19a to 19g) for forming a vibration generator (12) including rotation axes (20, 20a, 20b) oriented parallel to one another; a flanged wheel (15); The impact force (F, F) is applied to the track grid (4) consisting of the sleepers (5) of the track (2) to be stabilized and the rails (6) fixed on the sleepers (5). ｍａｘ , F ｒｅｄ a pressure roller (16) for transmitting the Equipped with In a stabilization unit (11), each flanged wheel (15) is rotatably supported about a wheel axis (17) and has a running surface (23) with a predetermined wheel diameter (d), The rotating shafts (20, 20a, 20b) are ｍａｘ , F ｒｅｄ ) in a horizontal plane of action (14), and the impact forces (F, F ｍａｘ , F ｒｅｄ The stabilizing unit (11) is characterized in that the horizontal working plane (14) of the flanged wheels (15) is located at most 300 mm, in particular at most 260 mm, above the rolling plane (25) of the flanged wheels (15).

2. 2. The stabilization unit (11) according to claim 1, wherein the horizontal plane of action (14) is located at a point less than half the wheel diameter (d / 2) above a horizontal plane (18) located through each wheel axis (17).

3. 3. The stabilization unit (11) according to claim 1 or 2, wherein at least two rotational shafts (20, 20a, 20b) and / or unbalanced masses (19, 19a-19g) are connected to transmission elements (30, 49, 56) and are driven by a common drive (57).

4. 4. The stabilizing unit (11) according to claim 1, wherein at least one unbalanced mass (19, 19a to 19g) is rotatably supported on each of the rotation shafts (20, 20a, 20b).

5. 5. The stabilization unit (11) according to claim 4, wherein at least one unbalanced mass (19b) is coupled to the associated rotation axis (20a) by means of a coupling element (26) that depends on the direction of rotation, such that the unbalanced mass (19a) rotates, in particular by 180°, relative to the rotation axis (20a) when the direction of rotation changes.

6. 6. The stabilization unit (11) according to claim 5, wherein the at least one unbalanced mass (19b) is coupled to the associated rotation shaft (20a) by means of a centrifugal lock (52).

7. 7. The stabilizing unit (11) according to claim 1, wherein a central rotation axis (20a) is arranged parallel to the longitudinal direction (7), and two lateral rotation axes (20b) are arranged to the left and right of the central rotation axis.

8. 8. The stabilization unit (11) according to claim 7, wherein the unbalanced mass (19a, 19b, 19c) associated with the central rotation axis (20a) has an unbalance amount twice as large as the unbalanced mass (19d, 19e, 19f, 19g) associated with each of the lateral rotation axes (20b).

9. 9. The stabilization unit (11) according to claim 7 or 8, wherein the directly driven unbalanced masses (19a, 19c, 19g, 19f) are connected to a common drive (57) and the indirectly driven unbalanced masses (19b, 19d, 19e) are connected to the directly driven unbalanced masses (19a, 19c, 19g, 19f) via a recirculating transmission (30).

10. 10. The stabilizing unit (11) according to claim 9, wherein the cage (34) of the recirculating transmission (30) is rotatably supported and connected to a rotary drive (39).

11. 11. The stabilization unit (11) according to claim 1, further comprising an acceleration sensor (59) for detecting the acceleration caused by the vibration generator (12).

12. A railway vehicle (1) having a machine frame (9) that can travel on a track (2) while being supported on a rail running mechanism (10), 12. A railway vehicle (1), characterized in that at least two stabilizing units (11) according to any one of claims 1 to 11 are arranged so that a front stabilizing unit (11) with a first height adjustment drive (13) is fixed to the machine frame (9) and a rear stabilizing unit (11) with a second height adjustment drive (13) is fixed to the machine frame (9).

13. 13. The railway vehicle (1) according to claim 12, wherein the vibration generators (12) and the height adjustment drives (13) are driven by a common control device (58), the control device (58) being configured to drive each vibration generator (12) and each height adjustment drive device (13) separately.

14. A method for operating a railway vehicle (1) according to claim 12 or 13, comprising: The railway vehicle (1) is moved forward, and the front stabilization unit (11) applies a vertical impact force (F, F ｍａｘ , F ｒｅｄ ) and the rear stabilization unit (11) is driven by a horizontal impact force (F, F ｍａｘ , F ｒｅｄ ) is driven by.

15. 15. The method according to claim 14, wherein vertical accelerations are detected in the front stabilization unit (11) using an acceleration sensor (59), and the reaction force characteristics of the track structure (4) are derived therefrom.

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