STABILIZATION UNIT AND METHOD FOR STABILIZING A TRACK

Description

Technical field

[0001] The invention relates to a stabilization unit for stabilizing a track, comprising a vibration generator with parallel rotating shafts and unbalanced masses for generating an impact force with adjustable direction, and flanged wheels and pressure rollers for transmitting the impact force to a track grid of the track to be stabilized, consisting of sleepers and rails attached thereto, wherein each flanged wheel is rotatably mounted about a wheel axle and has a running surface with a wheel diameter. The invention further relates to a rail vehicle with such a stabilization unit and a method for operating the rail vehicle. State of the art

[0002] A ballasted track is subject to continuous stress from rail traffic and environmental influences. For example, the position of the track section within the ballast bed changes. The ballast bed itself becomes contaminated over time through abrasion and the introduction of foreign matter. Maintenance measures such as tamping or cleaning address these issues. However, these processes result in a temporary loosening of the ballast bed. Even after optimal compaction using a tamping unit, subsequent settlement can occur. A track stabilization machine is used to prevent such settlement.

[0003] The machine is movable on the track and includes a stabilization unit that is clamped to the rails by means of rollers. A vibration generator mounted on the stabilization unit produces vibrations that are transmitted to the track bed. The design and dimensions of the vibration generator determine the impact force that acts on the track at the vibration frequency. To generate a static load, the stabilization unit is supported against a machine frame. The transmitted vibrations cause the particles in the ballast bed to become mobile, shift, and settle into a denser arrangement. This optimized ballast compaction results in an increase in the load-bearing capacity and lateral resistance of the track.

[0004] AT 16604 U1 discloses an exemplary stabilization unit with variable impact force. The vibration generator comprises several rotating unbalanced masses arranged on parallel shafts. The unbalanced masses are driven with a variable phase shift relative to each other. Depending on the arrangement of the unbalanced masses, a change in phase shift alters both the direction and the magnitude of the impact force. Description of the invention

[0005] The invention is based on the objective of improving a stabilization unit of the type mentioned above in such a way that the impact force acts on the track in an optimized manner. Furthermore, it is an objective of the invention to provide a rail vehicle that utilizes the extended application possibilities of the improved stabilization unit. In addition, an advantageous method for operating such a rail vehicle is to be provided.

[0006] According to the invention, these problems are solved by the features of independent claims 1, 12 and 14. Dependent claims specify advantageous embodiments of the invention.

[0007] In the new stabilization unit, the rotating shafts for generating the impact force are arranged in a horizontal plane of action such that this horizontal plane lies no more than 300 millimeters, and in particular no more than 260 millimeters, above the rolling plane of the flanged wheels. This low horizontal plane of action prevents disruptive tilting moments during stabilization. In operation, the rolling plane of the flanged wheels corresponds to a plane defined by the top of the rails on the track to be stabilized. If the horizontal impact force lies no more than 260 millimeters above this rolling plane or the top of the rails, saddle bearing of the sleepers can be reliably prevented. This also applies at a maximum value of 300 millimeters, although in this case, a larger clearance is available below the stabilization unit for the installation of a chord-measuring system or an optical measuring system.

[0008] Advantageously, the horizontal plane of action lies less than half a wheel diameter above a horizontal plane defined by the respective wheel axle. The vibration generator is positioned correspondingly low, with the wheel diameters being sufficiently large to prevent harmful pressure peaks from forming on the rail surfaces. The flanged wheels are spaced far enough apart to allow sufficient clearance for the vibration generator. This also applies to the elements of a spreading axle, which press the flanged wheels against the rails during operation. In conventional stabilization units, the vibration generator is always located above the flanged wheels, resulting in a high horizontal plane of action for the impact force. The resulting tilting moments, if severe, can cause the sleepers to saddle over on a ballast layer in the middle of the track.

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

[0010] In a further improvement, at least one unbalanced mass is mounted on each rotating shaft and is rotatable. This unbalanced mass can be driven with a different angular position, rotational speed, and direction of rotation relative to an unbalanced mass fixed on the rotating shaft. This allows the direction and magnitude of the resulting centrifugal force to be adjusted.

[0011] Preferably, at least one unbalanced mass is coupled to the associated rotating shaft by means of a rotation-direction-dependent coupling element such that, when the direction of rotation changes, the unbalanced mass rotates relative to the rotating shaft, in particular by 180°. Together with an unbalanced mass fixed on the rotating shaft, this results in two different centrifugal forces depending on the direction of rotation. This allows the stabilizing unit to be operated with different impact forces at the same vibration frequency.

[0012] In a further development of this variant, at least one unbalanced mass is coupled to the associated rotating shaft by means of a centrifugal locking mechanism. This centrifugal locking mechanism secures the unbalanced mass to the associated rotating shaft as soon as a predetermined rotational speed is exceeded. This ensures that no unwanted flipping of the unbalanced mass occurs during operation.

[0013] An advantageous design with a low center of gravity comprises a central rotating shaft parallel to one longitudinal axis of the assembly, with a lateral rotating shaft to the left and right of it. This results in a symmetrical design with different drive options, largely avoiding disruptive tilting moments during operation.

[0014] With an improved design, the unbalanced masses assigned to the central rotating shaft exhibit twice the imbalance of the unbalanced masses assigned to the respective lateral rotating shafts. This allows for stepless adjustment of the impact force from zero.

[0015] A further improvement involves coupling directly driven eccentric masses with a common drive, while indirectly driven eccentric masses are coupled to the directly driven eccentric masses via a planetary gear. The resulting combined centrifugal force of all eccentric masses can be adjusted via the planetary gear.

[0016] Advantageously, one cage of the planetary gear is rotatably mounted and coupled to a rotary drive. The rotary drive sets the cage in motion, thereby changing the relative angular velocity of the directly driven eccentric masses compared to the indirectly driven eccentric masses.

[0017] A further preferred embodiment of the stabilizing unit comprises an acceleration sensor for detecting acceleration caused by the vibration generator. Either the movements of the stabilizing unit or of the vibrating track grid are detected in order to infer the reaction force of the track grid.

[0018] The rail vehicle according to the invention comprises a machine frame, which is supported on rail bogies and is movable on a track, and on which the stabilization unit according to the invention is arranged, wherein a further stabilization unit is arranged on the machine frame, wherein the front stabilization unit is attached to the machine frame by first height adjustment drives and wherein the rear stabilization unit is attached to the machine frame by second height adjustment drives. In this way, the stabilization units can be operated independently of one another with different superimposed loads and different impact forces.

[0019] Advantageously, the vibration generators and the height adjustment drives are controlled by a common control unit, which is configured to separately control each vibration generator and each height adjustment drive. The two stabilization units can be operated in a coordinated manner using the common control unit. For example, the track grid is subjected to synchronized vibrations.

[0020] In the inventive method for operating the rail vehicle according to the invention with two stabilization units, a forward movement occurs along the track to be stabilized, wherein the front stabilization unit is operated with a vertical impact force and the rear stabilization unit is operated with a horizontal impact force. This operating mode simulates the passage of a rail vehicle in normal operation, because a sinusoidal movement of the rail vehicle's bogies usually follows a lifting shaft that precedes the rail vehicle. The rail vehicle according to the invention anticipates these processes and thus leaves behind a particularly sustainably stabilized track.

[0021] A further development of the method utilizes an acceleration sensor of the front stabilization unit to detect vertical accelerations and derive a reaction force profile of the track grid. Specifically, the corresponding reaction force is determined using the measured force-proportional acceleration and the known forces from the dynamic excitation. Brief description of the drawings

[0022] The invention is explained below by way of example with reference to the accompanying figures. These show, in schematic representation: Fig. 1 Rail vehicle with stabilization unit Fig. 2 Track cross-section with stabilization unit Fig. 3 Flanged wheel and pressure roller in rail contact Fig. 4 Top view and cross-section of a stabilization unit with three mechanically coupled rotating shafts Fig. 5 Change of direction of rotation of the rotating shafts of the stabilization unit according to Fig. 4 Fig. 6 Stabilization unit according to Fig. 4 with maximum horizontal force excitation Fig. 7 Stabilization unit according to Fig. 4 with maximum vertical force excitation Fig. 8 Stabilization unit according to Fig. 4 During rotation of the planetary gear Fig. 9 stabilization unit according to Fig. 4 with differently set impact force Fig. 10 Impact force curves at different drive states of the stabilization unit according to Fig. 4 Fig. 11 Reduction factor of a vibration amplitude of the stabilization unit according to Fig. 4 Fig. 12 Working diagram of a vibrating track grid area Fig. 13 Drive shaft with unbalanced masses Fig. 14 Unbalanced mass with centrifugal locking mechanism Description of the embodiments

[0023] A in Fig. 1 The depicted rail vehicle 1 is a so-called dynamic track stabilizer for stabilizing a ballasted track 2 following a tamping process. The track 2 comprises a ballast bed 3 in which a track grid 4, consisting of sleepers 5 and rails 6 attached to them, is mounted. During continuous forward travel of the rail vehicle 1 in a longitudinal direction 7, the track grid 4 is set into vibration and pressed into the ballast bed 3. This controlled settling of the track grid 4 is detected by means of a chord-measuring system 8 or by means of optical measuring devices. The exemplary rail vehicle 1 comprises a machine frame 9, which is supported on rail bogies 10 and can be moved on the track 2 to be stabilized. Two stabilization units 11 are movably connected to the machine frame 9. In other machines, only a single stabilization unit 11 is arranged.

[0024] Fig. 2 Figure 1 shows a cross-section of track 2 with the stabilization unit 11 during a stabilization process. The stabilization unit 11 comprises two independent main components: a vibration generator 12 and a pair of height actuators 13 (surcharge hydraulic cylinders). The vibration generator 12 alternately generates an impact force F in two opposite directions within a plane of action 14, causing vibrations of the stabilization unit 11. Preferably, the impact force F acts in a horizontal plane. This horizontal plane of action 14 is essential for the present invention. However, for extended operation of the stabilization unit 11, the impact force F also acts in a vertical direction. In this case, the plane of action 14 is a vertical plane.

[0025] Flanged wheels 15 and pressure rollers 16 transmit the vibrations to the track grid 4. Each flanged wheel 15 is rotatably mounted about a wheel axle 17 and is guided along an inner rail edge. The wheel axles 17 lie in a common horizontal plane 18. The pressure rollers 16 are pressed against the rails 6 from the outside. A continuously adjustable surcharge A is applied by means of the height adjustment drives 13.

[0026] Advantageously, the stabilizing unit 11 comprises a self-supporting central section with the vibration generator 12. The vibration generator 12 includes unbalanced masses 19 mounted on rotating shafts 20. Viewed in the longitudinal direction of the track, a side frame is connected to the central section on each side. The connection of the central section to the respective side frame is achieved, for example, by means of bolted connections on a circumferential flange. The flanged wheels 15 and the pressure rollers 16 are mounted exclusively on the associated side frame. To implement a spreading axis 21, for example, the flanged wheels 15 assigned to each of the side frames are each coupled to a hydraulic drive to effect a displacement along the associated wheel axis 17. There is no common continuous shaft for the front or rear flanged wheels 15.The absence of a continuous shaft creates space for the low placement of the central section. This results in a low center of gravity 22 for the entire stabilization unit 11 and a low effective plane 14 for the vibration generator 12. Preferably, the center of gravity 22 lies in the horizontal effective plane 14.

[0027] Each flanged wheel 15 has a wheel diameter d, which is measured at a running surface 23. In operation, the running surfaces 23 of the flanged wheels 15 are in contact with the top edges 24 of the rails 6. The top edge 24 is the highest line on a rail head. A lower and an upper tangential plane are present on all running surfaces 23 of the flanged wheels 15. The lower tangential plane forms a rolling plane 25, in which, in operation, the edge contact points between the running surfaces 23 of the flanged wheels 15 and the top edges 24 of the rails lie. According to the invention, the vibration generator 12 is arranged so low that the vertical distance a between the horizontal effective plane 14 of the impact force F and the rolling plane 25 is at most 300 millimeters, in particular at most 260 millimeters. Very good results were achieved in tests with a vertical distance a = 250 mm. Even under unfavorable compaction conditions of the ballast, no saddle bearing of the sleepers occurred.

[0028] Advantageously, the horizontal effective plane 14 lies less than half a wheel diameter d / 2 of the respective flanged wheel 15 above the horizontal plane 18 defined by the respective wheel axle 17. The upper tangent plane of the running surfaces 23 forms a boundary plane below which the horizontal effective plane 14 lies. If the horizontal effective plane 14 lies above the wheel axles 17, a further vertical distance b between this effective plane 14 and the horizontal plane 18 is less than half a wheel diameter d / 2 of the respective flanged wheel 15. This characteristic is fulfilled whenever the horizontal effective plane 14 lies below the wheel axles 17. This results in two advantages. Firstly, the horizontal effective plane 14 lies sufficiently low, and secondly, the wheel diameters d of the flanged wheels 15 are large enough to prevent harmful pressure peaks from forming on the rail surfaces.

[0029] An advantageous embodiment of the vibration generator 12 with reduced height is described with reference to the Figuren 4-11 As explained, to achieve a deep-lying force excitation, cylindrical unbalanced masses 19 are arranged, rotating about axes 7 aligned in the longitudinal direction. The unbalanced masses 19 are divided into three groups in the axial direction to allow for a continuously adjustable amplitude of the effective impact force F. In the illustrated example, most of the unbalanced masses 19a, 19b, 19c, 19d, 19e are freely rotatable on a driven central rotating shaft (drive shaft) 20a or on coupled lateral rotating shafts (auxiliary shafts) 20b. Other unbalanced masses 19f, 19g are fixedly connected to the associated rotating shaft 20a, 20b.

[0030] The unbalanced masses 19a, 19b, 19c, which are mounted on the drive shaft 20a, have a direction-of-rotation-dependent coupling element 26 by which they are connected to a respective drive mechanism. Such a drive mechanism is, for example, a cylindrical wheel fixed to the drive shaft 20a with a corresponding recess for a reversing bolt. By changing the direction of rotation of the drive shaft 20a, the unbalanced masses 19d, 19e, 19f, 19g on the auxiliary shafts 20b are rotated by 180° relative to an initial position, while the unbalanced masses 19a, 19b, 19c on the drive shaft 20a retain their position. This principle is described in Fig. 5 The process is depicted in three successive phases 27, 28, and 29. In the first phase 27, the unbalanced masses 19a-19g are in a starting position for horizontal force application, with only one unbalanced mass 19b from the middle group being shown. All unbalanced masses 19a-19g are pointing to the right. The second phase 28 shows the reversal process, and the third phase 29 shows the unbalanced position for vertical force application.

[0031] The unbalanced masses 19b, 19d, 19e of the middle group have twice the unbalance (product of mass and eccentricity, U=m·e) compared to the unbalanced masses 19a, 19c, 19f, 19g of the two outermost groups. Additionally, an unbalanced mass 19a, 19b, 19c of the drive shaft 20a has twice the unbalance of an unbalanced mass 19a, 19c, 19f, 19g of the auxiliary shaft 20b of the corresponding group. For example, the unbalanced masses 19a, 19b, 19c have twice the unbalance of the unbalanced masses 19d, 19e, 19f, 19g, wherein 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 enables a vibration generator 12 with a continuously adjustable impact force amplitude of the overall system, whereby the centrifugal forces cancel each other out in one direction.

[0032] The unbalanced masses 19a-19g are driven by the drive shaft 20a and a planetary gear 30. In Fig. 4 The diagram shows a rotary motion 31 of the drive shaft 20a for horizontal excitation and a rotary motion 32 of the drive shaft 20a for vertical excitation. The rotary motion 31, 32 of the drive shaft 20a is transmitted directly to the unbalanced masses 19a, 19c on the drive shaft 20a. Subsequently, the rotary motion is transmitted from the unbalanced mass 19c via gear wheels to the adjacent unbalanced masses 19f, thereby also driving the unbalanced masses 19g, which are rigidly connected to the respective countershaft 20b. Additionally, the rotary motion is transmitted via a gear wheel on the unbalanced mass 19c of the drive shaft 20a to a gear drive shaft 33 of the planetary gear set 30. In normal operation, a differential cage 34 of the planetary gear 30 is at rest, whereby a counter-rotating rotary motion with the same rotational speed is transmitted via a gear output shaft 35 to the unbalance mass 19e of the countershaft 20b.The other unbalance masses 19b, 19d of the middle group are driven via a coupling with this unbalance mass 19e.

[0033] Fig. 4 The top image shows the differential cage 34 with differential bolts 36, compensating bevel gears 37 and axle bevel gears 38. The middle image shows the top view of the assembly 11, with a section line through the middle group of unbalance masses 19b, 19d, 19e giving the cross-section of the assembly 11 in the bottom image.

[0034] Fig. 6 Figure 19a-19g shows the system of unbalanced masses for a maximum possible horizontal force excitation. On the left, an unbalanced position S1 with the maximum horizontal impact force Fmax is shown. On the right, the system is in an unbalanced position S2, rotated by an angle α = 90°, where the resulting excitation force Ferr is zero.

[0035] In Fig. 7 The operating principle of the system of unbalanced masses 19a-19g for a maximum possible vertical force excitation is shown. A left unbalance position S3 shows the system with the resulting excitation force Ferr equal to zero. On the right, the system is in an unbalance position S4, rotated by the rotation angle α = 90°, resulting in the maximum vertical impact force Fmax.

[0036] Fig. 8 Figure 1 shows two unbalance positions S5 and S6 for the system of unbalance masses 19a-19g during rotation of the planetary gear 30. On the left, an unbalance position S5 is shown for a resulting angular difference of 90° between the unbalance masses 19a, 19g and 19c, 19f of the outer groups and the unbalance masses 19b, 19d, 19e of the middle group, and on the right, an unbalance position S6 is shown for a resulting angular difference of 180° between the unbalance masses 19a, 19g and 19c, 19f of the outer groups and the unbalance masses 19b, 19d, 19e of the middle group (unexcited operation, no-load operation).

[0037] To reduce the amplitude of the effective impact force F of the system, the unbalanced masses 19b, 19d, 19e of the middle group are rotated relative to the other unbalanced masses 19a, 19c, 19f, 19g, so that the excitation forces FM of the middle group are reduced or equalized when superimposed with the excitation forces FR of the outer groups, depending on the setting ( Fig. 9 and Fig. 10 For this purpose, the differential cage 34 of the planetary gear 30 is rotated during operation by a rotary drive 39 coupled to the differential cage 34 at an angular velocity ωD. Using the known angular velocity ω1 of the gear input shaft 33, the angular velocity ω2 of the gear output shaft 35 can be determined with the Willis equation when the planetary gear 30 is rotated in the opposite direction to the rotation of the gear input shaft 33. ω 2 = 1 i 0 ⋅ ω 1 − ω D ⋅ 1 − i 0

[0038] For differential gears, the standard gear ratio i 0 =-1 can be assumed: ω 2 = 2 ⋅ ω D − ω 1

[0039] In Fig. 9 The system of unbalance masses 19a-19g with the three unbalance positions S1, S5 and S6 from the Figuren 6-8 The diagram shows that in the upper image, the difference angle β between the outermost groups and the middle group is zero. This difference angle β is 90° in the middle image and 180° in the lower image. To the right of the groups of unbalanced masses 19a-19g, the force angles of the respective excitation forces FE and the respective resulting excitation force Ferr are shown.

[0040] Under normal operating conditions, the differential cage 34 is at rest and ωD = 0, meaning ω₂ = -ω₁. The gear output shaft 35 and the gear input shaft 33 then have the same angular velocity but opposite directions of rotation. When the planetary gear 30 rotates, the gear output shaft 35 moves faster than the gear input shaft 33, with the difference corresponding to twice the angular velocity ωD of the rotating differential cage 34. A twist between the unbalanced masses 19b, 19d, 19e of the middle group and the unbalanced masses 19a, 19c, 19f, 19g of the outer groups results from the gear ratio between the shafts 33, 35 of the planetary gear 30 and the corresponding unbalanced masses 19c, 19e.If the gear ratio between the shafts 33, 35 of the planetary gear 30 and the unbalanced masses 19c, 19e is, for example, i=-1 (opposite direction of rotation at the same angular velocity), a rotation (and subsequent fixing) of the differential cage 34 by an angle β / 2 results in an angular difference of β between the unbalanced masses 19b, 19d, 19e of the middle group and the unbalanced masses 19a, 19c, 19f, 19g of the outer groups. Due to this angular difference β (phase shift), the effective horizontal impact force F of the system in horizontal operation or the effective vertical impact force F in vertical operation is reduced.

[0041] The excitation force FE (centrifugal force) of a single unbalanced mass 19 results from the product of mass m, eccentricity e and the square of the angular velocity ω U at the center of rotation: F E = m ⋅ e ⋅ ω U 2

[0042] In horizontal operation, any vertical components of the respective unbalance mass groups cancel each other out (e.g., the vertical components of unbalance masses 19a and 19g cancel each other out), whereby the maximum excitation force FRmax, FMmax of an unbalance mass group is reached precisely when all unbalance masses 19a-19g of the respective group are either horizontal or vertical. This occurs without rotation of the unbalance mass groups relative to each other (angle difference β=0, unbalance mass position S1 in). Fig. 6 left and S4 in Fig. 7 (right), the excitation forces FE of the unbalanced masses 19a, 19g and 19c, 19f of the outer groups and the unbalanced masses 19b, 19d, 19e of the middle group add up completely, with the excitation force FM of the middle group having the same magnitude as the excitation force FR of the two synchronously running outer groups. In this state, the system operates with the maximum resulting impact force F max (maximum possible impact force amplitude).

[0043] If the unbalance masses 19a-19g are rotated relative to each other, the maximum excitation forces F Rmax , F Mmax can never be completely superimposed ( Fig. 9 and Fig. 10 ) and a reduced impact force Fred of the system occurs. Due to the rotation of the unbalanced masses 19a-19g relative to each other, for example, the vibration generated by the middle group leads the vibration generated by the outer groups (phase shift 40). This effect is in Fig. 10 shown, wherein the vibrations of the middle group and the two synchronously running outer groups are depicted for two complete revolutions of the unbalance masses 19a-19g, starting from the horizontal position of the outer group unbalance masses 19a, 19c, 19f, 19g.

[0044] All three diagrams branch off into Fig. 10 Curves of the summed excitation force FR of the marginal groups and the excitation force FM of the middle group as well as the resulting excitation force F err over a rotation angle α, where the initial position corresponds to the three unbalance positions S1, S5 and S6 in Fig. 9 The diagram shows the course of the summed excitation force FR of the marginal groups with dotted lines, the course of the excitation force FM of the middle group with a dashed line, and the course of the resulting excitation force Ferr with a solid line. The phase shift 40 of the amplitude of the overall system relative to the amplitude of the marginal groups or to the amplitude of the middle group corresponds to half the difference angle β. This difference angle β is 0° in the upper diagram, 90° in the middle diagram, and 180° in the lower diagram.

[0045] The vibration of the overall system results from the superposition of the vibrations of the boundary groups and the middle group. Since the maximum amplitude of the horizontal or vertical excitation force FMmax of the middle group is the same as the sum of the excitation forces FRmax of the two boundary groups (FRmax = FMmax = Fmax / 2), the reduced (horizontal) excitation force Ferr can be described as a function of the rotation angle α and the difference angle β as follows: F err = F Rmax ⋅ cos α + F Mmax ⋅ cos α + β = F max 2 ⋅ cos α + F max 2 ⋅ cos α + β

[0046] A reduced impact force F red as a function of the maximum impact force F max (maximum excitation force) is derived from this equation by means of an extremum analysis and is obtained as a function of the difference angle β between the boundary group and the middle group: F red = F max ⋅ cos β 2

[0047] The reduction factor cos(β / 2) is in Fig. 11 The diagram shows that the maximum impact force Fmax (maximum excitation amplitude) occurs at a difference angle β of 0°. Specifically, it shows Fig. 11 the course of the reduction factor cos(β / 2) over the difference angle β between 0° and 180°.

[0048] A stabilization unit 11 according to the invention is preferably operated in pairs, as shown in Fig. 1 As shown. With two stabilization units 11 used in series, several combination possibilities for gravel compaction arise due to the variable excitation direction: both units 11 in horizontal operation, both units 11 in vertical operation, or one unit 11 in vertical operation and the other unit 11 in horizontal operation.

[0049] An advantage of the present invention with regard to the compaction effect lies in the low-lying center of gravity 22 or in the low-lying plane of action 14, in which the point of application of the horizontal force excitation is located. This makes a predominantly translational excitation of the track grid 4 possible.

[0050] Previously, vertical force excitation for compacting the track ballast was only possible in the spaces between the sleepers and on the flanks of the ballast superstructure using spacer compactors and head compactors. The present invention additionally enables vertical excitation of the ballast below each sleeper. Care must be taken to ensure that the stabilizing unit 11 does not lift off the railheads to avoid damage (head checks, corrugation). For safe operation, the vertical load A is set high enough by means of the height adjustment drives 13 to limit the relieving effect of the centrifugal forces of the vibration generator 12.

[0051] Since the vertical stiffness below the sleeper 5 is greater than in the horizontal direction, vertical operation results in a stronger interaction between the rail vehicle 1, the track grid 4 and the ballast bed 3. The machine parameters must therefore be particularly carefully adapted to the local conditions during purely vertical excitation, especially to the condition of the track ballast, the geometry of the ballast bed 3 and the underlying ground.

[0052] An operating mode in which, in one direction of travel, the front of the two stabilization units 11 is excited vertically and the rear of the two stabilization units 11 is excited horizontally, simulates a rail vehicle passing over the track during normal operation. During such a crossing in normal operation, a leading lift wave typically occurs in front of the rail vehicle (vertical excitation) followed by a sinusoidal wave (horizontal excitation). Because this load from the compaction process thus approximates the subsequent load from rail traffic, it has a beneficial effect on the durability of the preceding track geometry corrections.

[0053] The ability to compact the track ballast below each sleeper (5) using vertical force excitation, combined with subsequent horizontal excitation, leads to improved compaction results. A significant advantage of this operating mode lies in the compaction control. For this purpose, the vertical force excitation is selected to be so low that no compaction effect occurs. In this way, statements about the vertical stiffness can be made without disturbing the ballast structure and thus the track geometry.

[0054] To determine the ballast compaction, acceleration signals are measured at the stabilization unit 11, as described in AT 521481 A4. Since the measured accelerations are proportional to the force and the forces from the dynamic excitation are known, the reaction force profile from the track grid 4 can be determined from the difference. To assess the compaction success, a characteristic value is subsequently derived from a corresponding operating diagram (at constant excitation frequency) or via an impedance function (at variable excitation frequency, dynamic stiffness). An example of an operating diagram is shown in Fig. 12 The diagram shows the displacement 41 of the activated track section on the abscissa. The ordinate indicates a contact force 42 below the activated sleepers 5. From this diagram, the stiffness (relationship between defined force difference 43 and measured displacement 44 when the track section 4 is loaded), the damping of the system (curvature of the curve), and the energy input 45 (circumscribed area) are determined. Dashed horizontal lines indicate a static load 46, a minimum vertical load 47, and a maximum vertical load 48.

[0055] For compaction control, a mechanical model of the track grid 4 is used. Through an optimization process, deterministic parameters of the track grid 4 are subsequently derived, which, given an excitation with a known power spectral density, lead to the measured response. A measured value determined in this way has the advantage that it is directly interpretable physically and serves as a basis for planning track maintenance.

[0056] Fig. 13 Figure 1 shows an improved version of the drive shaft 20a with the unbalanced masses 19a, 19b, and 19c. The two outer unbalanced masses 19a and 19c are fixedly connected to the drive shaft 20a. The middle unbalanced mass 19a is rotatably mounted on the drive shaft 20a and coupled to a gear 49 via the direction-dependent coupling element 26. The position of the middle unbalanced mass 19b relative to the gear 49 depends on the direction of rotation. In the position shown, the coupling element 26, designed as a reversing bolt, rests in an upper drive 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 reversing bolt rests in a lower drive recess 51 of the unbalanced mass 19b.

[0057] To prevent an unwanted flipping of the unbalanced mass 19b, a centrifugal locking device 52 is arranged. Fig. 14Figure 51 shows this detail in a side view. A lever 53 is assigned to both the upper drive recess 50 and the lower drive recess 51. Each lever 53 is rotatably mounted at one end on the unbalanced mass 19b. When the unbalanced mass 19b is stationary or rotating at low speed, the respective lever 53 is pressed inwards by an associated spring 54. In this state, the drive recesses 50 and 51 are free to receive the coupling element 26. As the rotational speed increases, centrifugal force pushes both levers 53 outwards. One of the levers 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. A sensor is conveniently located to monitor the position of the unbalanced mass 19b.

[0058] In the illustrated embodiment, the unbalanced mass 19b is coupled via the gear 49 to the two unbalanced masses 19d, 19e of the middle group and to the gearbox output shaft 35 of the planetary gear 30. A further gear 56 is arranged on the drive shaft 20a with a further direction-dependent coupling element 26. This further gear 56 rotates 180° relative to the drive shaft 20a when the direction of rotation changes and couples the drive shaft 20a to the two auxiliary shafts 20b and to the gearbox input shaft 33 of the planetary gear 30. These gear elements 30, 49, 56 thus couple all rotating shafts 20a, 20b and unbalanced masses 19a-19g, with the drive shaft 20a being connected to a common drive 57.

[0059] To carry out the method according to the invention, it is advantageous if the stabilization units 11 arranged on the rail vehicle 1 are controlled by means of a common control device 58. The control device 58 is configured to separately control the vibration generator 12 and the height adjustment drives 13 of the respective stabilization unit 11. Preferably, an acceleration sensor 59 arranged on the front stabilization unit 11 transmits an acceleration signal to the control device 58 in order to subsequently evaluate the reaction force profile of the track grid 4.

Claims

1. A stabilizing unit (11) for stabilizing a track (2), with a vibration generator (12) comprising rotating shafts (20, 20a, 20b) aligned parallel to one another with unbalance masses (19, 19a-19g) for generating an impact force (F, Fmax, Fred) with adjustable direction, and with flanged wheels (15) and pressing rollers (16) for transmitting the impact force (F, Fmax, Fred) to a track panel (4) of the track (2) to be stabilized, consisting of sleepers (5) and rails (6) fixed thereon, with each flanged wheel (15) being mounted to rotate about a wheel axis (17) and having a running tread (23) with a wheel diameter (d), characterized in that the rotating shafts (20, 20a, 20b) are arranged to generate the impact force (F, Fmax, Fred) in a horizontal plane of action (14) in such a way that the horizontal plane of action (14) of the impact force (F, Fmax, Fred) is no more than 300 millimetres, in particular no more than 260 millimetres, above a rolling plane (25) of the flanged wheels (15).

2. A stabilizing unit (11) according to claim 1, characterized in that the horizontal plane of action (14) lies less than half a wheel diameter (d / 2) above a horizontal plane (18) passing through the respective wheel axis (17).

3. A stabilizing unit (11) according to claim 1 or 2, characterized in that at least two rotating shafts (20, 20a, 20b) and / or unbalance masses (19, 19a-19g) are coupled to gearbox elements (30, 49, 56) and driven by a common drive (57).

4. A stabilizing unit (11) according to one of the claims 1 to 3, characterized in that at least one unbalance mass (19, 19a-19g) is mounted to rotate on each rotating shaft (20, 20a, 20b).

5. A stabilizing unit (11) according to claim 4, characterized in that at least one unbalance mass (19b) is coupled to the assigned rotating shaft (20a) by means of a coupling element (26) that is dependent on the direction of rotation in such a way that when the direction of rotation changes, the unbalance mass (19a) is rotated relative to the rotating shaft (20a), in particular by 180°.

6. A stabilizing unit (11) according to claim 5, characterized in that the at least one unbalance mass (19b) is coupled to the assigned rotating shaft (20a) by means of a centrifugal-force locking mechanism (52).

7. A stabilizing unit (11) according to one of the claims 1 to 6, characterized in that a central rotating shaft (20a) parallel to a longitudinal direction (7) and a lateral rotating shaft (20b) are arranged to the left and right thereof.

8. A stabilizing unit (11) according to claim 7, characterized in that the unbalance masses (19a, 19b, 19c) assigned to the central rotating shaft (20a) have an unbalance twice as large as the unbalance masses (19d, 19e, 19f, 19g) assigned to the respective lateral rotating shaft (20b).

9. A stabilizing unit unit (11) according to claim 7 or 8, characterized in that directly driven unbalance masses (19a, 19c, 19g, 19f) are coupled to a common drive (57), and that indirectly driven unbalance masses (19b, 19d, 19e) are coupled to the directly driven unbalance masses (19a, 19c, 19g, 19f) via a planetary gearbox (30).

10. A stabilizing unit (11) according to claim 9, characterized in that a cage (34) of the planetary gearbox (30) is mounted to rotate and coupled to a rotation drive (39).

11. A stabilizing unit (11) according to one of the claims 1 to 10, characterized in that an acceleration sensor (59) is arranged for recording an acceleration caused by means of the vibration generator (12).

12. A rail vehicle (1) with a machine frame (9) which is movable supported on rail running gears (10) on a track (2), a stabilizing unit (11) according to one of the claims 1 to 11 is arranged on the machine frame (9) wherein another stabilizing unit (11) is arranged on the machine frame (9) wherein the front stabilizing unit (11) with first height setting drives (13) is fastened to the machine frame (9), and wherein the rear stabilizing unit (11) with second height setting drives (13) is fastened to the machine frame (9).

13. A rail vehicle (1) according to claim 12, characterized in that the vibration generators (12) and the height setting drives (13) are actuated by means of a common control device (58), and that the control device (58) is set up for separate actuation of the respective vibration generator (12) and the respective height setting drive (13).

14. A method for operating a rail vehicle (1) according to claim 12 or 13, with two stabilizing units (11) wherein the rail vehicle (1) is moved forwards, wherein the front stabilizing unit (11) is operated with a vertical impact force (F, Fmax, Fred), and wherein the rear stabilizing unit (11) is operated with a horizontal impact force (F, Fmax, Fred).

15. A method according to claim 14, characterized in that vertical accelerations are recorded at the front stabilizing unit (11) by means of an acceleration sensor (59) in order to derive a reaction force progression of the track panel (4) therefrom.