Stabilising assembly and method for stabilising a track

Abstract

The invention relates to a stabilising assembly (11) for stabilising a track (2), comprising a vibration generator (12) having rotating shafts (20, 20a, 20b) that are orientated in parallel to one another and have unbalanced masses (19, 19a-19g) for generating an impact force (F, Fmax, Fred) with an adjustable direction, and comprising flange wheels (15) and pressure rollers (16) for transmitting the impact force (F, Fmax, Fred) onto a track panel (4), consisting of sleepers (6) and rails (6) secured thereon, of the track (2) to be stabilised, wherein each flange wheel (15) is rotatably mounted about a wheel axle (17) and has a running surface (23) with a wheel diameter (d). The rotating shafts (20, 20a, 20b) are arranged to generate the impact force (F, Fmax, Fred) in a horizontal effective plane (14) such that the horizontal effective plane (14) of the impact force (F, Fmax, Fred) lies max. 300 mm, in particular max. 260 mm, above a rolling plane (25) of the flange wheels (15). The sunken horizontal effective plane (14) prevents interfering tilting moments during a stabilisation process.

Description

Description STABILIZATION UNIT AND METHOD FOR STABILIZING A TRACK Technical area

[0001] The invention relates to a stabilization unit for stabilizing a track, comprising a vibration generator comprising parallel rotating shafts with imbalance masses for generating an impact force with an adjustable direction, and having flanged wheels and pressure rollers for transmitting the impact force to a track grid consisting of sleepers and rails attached thereto of the track to be stabilized. Each flanged wheel is rotatably mounted about a wheel axis and has a running surface with a wheel diameter. Furthermore, the invention 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 subjected to continuous wear and tear by rail traffic and environmental influences. For example, the position of a track grid within the ballast bed changes. The ballast bed itself becomes contaminated over time due to abrasion and the introduction of foreign matter. Maintenance measures such as tamping or cleaning can remedy these defects. However, this temporarily loosens the ballast bed. Even after optimal compaction using a tamping unit, subsequent settlement can occur. A track stabilizing machine is used to anticipate such settlement.

[0003] The machine is movable along the track and includes a stabilizing unit that is clamped to the rails by means of rollers. A vibration generator mounted on the stabilizing unit generates vibrations that are transmitted to the track grid. The design and dimensions of the vibration generator determine an impact force that acts on the track at the vibration frequency. To generate a static load, the stabilizing unit is supported against a machine frame. The transmitted vibrations cause the grains in the grain structure of the ballast bed to become mobile, shift, and become denser. This optimized ballast compaction results in an increase in the load-bearing capacity and the lateral displacement 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 variably adjustable phase shift relative to each other. Depending on the arrangement of the unbalanced masses, a changed phase shift changes both the direction and the strength of the impact force. Description of the invention

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

[0006] According to the invention, these objects are achieved 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 are arranged in a horizontal plane of action to generate the impact force, so that the horizontal plane of action of the impact force is at most 300 millimeters, in particular at most 260 millimeters above a rolling plane of the The low horizontal plane of action prevents disruptive tipping moments during a stabilization process. In operation, the rolling plane of the flanged wheels corresponds to a plane spanned by the top edges of the rails on the track to be stabilized. If the horizontal impact force is at most 260 millimeters above this rolling plane or the top edge of the rail, saddle bearing of the sleepers can be safely ruled out. This also applies at a maximum value of 300 millimeters, whereby more space is available below the stabilization unit for the installation of a chord measuring system or an optical measuring system.

[0008] Advantageously, the horizontal effective plane is less than half a wheel diameter above a horizontal plane passing through the respective wheel axle. The vibration generator is positioned correspondingly low, with the wheel diameters being sufficiently large to prevent harmful pressure peaks from occurring on the rail surfaces. The flanged wheels are spaced far enough apart to leave space between them for the vibration generator. This also applies to the elements of a spread 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 effective plane of the impact force. The resulting tipping moments, if severe, can lead to the sleepers saddled on a ballast layer in the center of the track.

[0009] Advantageously, at least two rotating shafts and / or unbalanced masses are coupled to gear elements and driven by a common drive. This allows the common drive to be used with optimized control to drive all rotating shafts or unbalanced masses. The type of coupling determines how the resulting impact force is derived from the centrifugal forces generated by the unbalanced masses. The centrifugal forces preferably reinforce each other in a desired plane of action, whereas the centrifugal forces cancel each other out in other planes of action.

[0010] In a further improvement, at least one unbalanced mass is rotatably mounted on each rotating shaft. This unbalanced mass can be driven with a different angular position, rotational speed, and rotational direction 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, upon a change in rotation direction, the unbalanced mass rotates relative to the rotating shaft, in particular by 180°. Together with an unbalanced mass fixedly mounted on the rotating shaft, two different resulting centrifugal forces result depending on the direction of rotation. This allows the stabilization 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 locks the unbalanced mass to the associated rotating shaft as soon as a specified speed is exceeded. This ensures that no unwanted rotation of the unbalanced mass occurs during operation.

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

[0014] An improvement to this design allows the unbalanced masses assigned to the central rotating shaft to exhibit twice the imbalance of the unbalanced masses assigned to the respective lateral rotating shafts. This allows for a continuous adjustment of the impact force from zero.

[0015] A further improvement is that directly driven unbalance masses are coupled to a common drive, whereby Indirectly driven imbalance masses are coupled to the directly driven imbalance masses via an epicyclic gear. The resulting combined centrifugal force of all imbalance masses can be adjusted via the epicyclic gear.

[0016] In this case, a cage of the epicyclic gear is advantageously mounted on a rotatable support and coupled to a rotary drive. The rotary drive allows the cage to rotate, thereby changing the relative angular velocity of the directly driven imbalance masses compared to the indirectly driven imbalance masses.

[0017] A further preferred development of the stabilization unit comprises an acceleration sensor for detecting an acceleration induced by the vibration generator. In this case, either the movements of the stabilization unit or the vibrating track grid are detected in order to determine the reaction force of the track grid.

[0018] The rail vehicle according to the invention comprises a machine frame that can be moved on rail bogies on a track, and at least two of the above-described stabilization units, wherein a front stabilization unit is attached to the machine frame with first height adjustment drives, and wherein a rear stabilization unit is attached to the machine frame with second height adjustment drives. In this way, the stabilization units can be operated independently of one another with different loads and different impact forces.

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

[0020] In the method according to the invention for operating the rail vehicle, a forward movement takes place along the track to be stabilized, wherein the front stabilization unit is actuated 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 lift-off wave preceding the rail vehicle is usually followed by a sinusoidal waveform of the rail vehicle's bogies. 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 uses an acceleration sensor in the front stabilization unit, which measures vertical accelerations and derives a reaction force curve 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. Short description of the drawings

[0022] The invention is explained below by way of example with reference to the accompanying figures. They show schematically: 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 rotation shafts Fig. 5 Change of direction of rotation of the rotation 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 epicyclic gear Fig. 9 Stabilization unit according to Fig. 4 with different set impact force Fig. 10 Impact force curves for 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 grate area Fig. 13 Drive shaft with unbalanced masses Fig. 14 Unbalanced mass with centrifugal locking Description of the embodiments

[0023] A rail vehicle 1 shown in Fig. 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 fastened thereto, 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 targeted settlement of the track grid 4 is recorded 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 along 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 shows a cross-section of the track 2 with the stabilization unit 11 during a stabilization process. The stabilization unit 11 comprises two independent main components, namely a vibration generator 12 and a pair of height adjustment drives 13 (load hydraulic cylinders). The vibration generator 12 generates an impact force F in an effective plane 14, alternating in two opposite directions, which causes vibrations of the stabilization unit 11. Preferably, the impact force F acts in a horizontal plane. This horizontal effective plane 14 is of essential importance for the present invention. For extended operation of the However, the stabilization unit 11 also has an effect of the impact force F in the vertical direction. The plane of action 14 is then a vertical plane.

[0025] Flanged wheels 15 and pressure rollers 16 transmit the vibrations to the track grid 4. Each flanged wheel 15 is mounted for rotation 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 load A is applied by means of the height adjustment drives 13.

[0026] The stabilization unit 11 advantageously comprises a self-supporting central section with the vibration generator 12. The vibration generator 12 comprises 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 central section is connected to the respective side frame, for example, by means of screw connections on a circumferential flange. The flanged wheels 15 and the pressure rollers 16 are mounted exclusively on the associated side frame. To create a spread axle 21, for example, the flanged wheels 15 associated with one of the side frames are each coupled to a hydraulic drive to effect displacement along the associated wheel axle 17. There is no common through shaft for the front or rear flanged wheels 15.The absence of a continuous shaft creates space for the low arrangement of the center section. This results in a low center of gravity 22 of the entire stabilization unit 11 and the low operating plane 14 of the vibration generator 12. Preferably, the center of gravity 22 is located in the horizontal operating plane 14.

[0027] Each flanged wheel 15 has a wheel diameter d, which is measured at a running surface 23. In use, 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 lie adjacent to all running surfaces 23 of the flanged wheels 15. The lower tangential plane forms a rolling plane. 25, in which, during use, the contact points are located between the running surfaces 23 of the flanged rollers 15 and the rail top edges 24. 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 ballast compaction conditions, 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 passing through the respective wheel axle 17. The upper tangential 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 feature is met in any case if the horizontal effective plane 14 lies below the wheel axles 17. This results in two advantages. On the one hand, the horizontal effective plane 14 lies deep enough and, on the other hand, the wheel diameters d of the flanged wheels 15 are large enough to prevent harmful pressure peaks from occurring on the rail surfaces.

[0029] An advantageous embodiment of the vibration generator 12 with reduced overall height is explained with reference to Figures 4-11. To achieve a low-lying force excitation, cylindrical unbalanced masses 19 are arranged, which rotate about axes aligned in the longitudinal direction 7. The unbalanced masses 19 are divided into three groups in the axial direction to enable a continuously adjustable amplitude of the effective impact force F. In the example shown, most of the unbalanced masses 19a, 19b, 19c, 19d, 19e are freely rotatably mounted on a driven central rotation shaft (drive shaft) 20a or on coupled lateral rotation shafts (auxiliary shafts) 20b. Other Unbalanced masses 19f, 19g are firmly 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-dependent coupling element 26, through which they are connected to a respective drive mechanism. Such a drive mechanism is, for example, a cylindrical gear permanently mounted on the drive shaft 20a with a corresponding recess for a reversing pin. 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 their initial position, while the unbalanced masses 19a, 19b, 19c on the drive shaft 20a retain their position. This principle is illustrated in Fig. 5 in three sequential phases 27, 28, 29. In the first phase 27, the unbalanced masses 19a-19g are in a starting position for a horizontal force excitation, with only one unbalanced mass 19b of the middle group shown. All unbalanced masses 19a-19g point to the right.The second phase 28 shows the turnover process and the third phase 29 shows the unbalance position for a vertical force excitation.

[0031] The unbalanced masses 19b, 19d, 19e of the middle group have twice the unbalance (product of mass and eccentricity, U=me) compared to the unbalanced masses 19a, 19c, 19f, 19g of the two outer groups. In addition, one unbalanced mass 19a, 19b, 19c of the drive shaft 20a has twice the unbalance of one unbalanced mass 19a, 19c, 19f, 19g of the secondary shaft 20b of the associated group. For example, the unbalanced masses 19a, 19b, 19c have twice the imbalance of the unbalanced masses 19d, 19e, 19f, 19g, with the unbalanced mass 19d having the same imbalance as the unbalanced mass 19e, and the unbalanced mass 19f having the same imbalance as the unbalanced mass 19g. This arrangement enables a vibration generator 12 with continuously adjustable impact force amplitude for the entire system, with the centrifugal forces canceling each other out in one direction.

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

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

[0034] Fig. 6 shows the system of unbalanced masses 19a-19g for 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 further rotated through an orbital angle a=90° in an unbalanced position S2, with the resulting excitation force Fen- being zero.

[0035] Fig. 7 illustrates the operation of the system of unbalanced masses 19a-19g for a maximum possible vertical force excitation. A left-hand unbalance position S3 shows the system with the resulting excitation force F equal to zero. The right-hand one shows the system in an unbalance position S4, rotated by an angle of rotation a=90°, resulting in the maximum vertical impact force F max.

[0036] Fig. 8 shows two unbalance positions S5, S6 for the system of unbalanced masses 19a-19g during rotation of the epicyclic gear 30. On the left, an unbalance position S5 is shown for a resulting angular difference of 90° between the unbalanced masses 19a, 19g and 19c, 19f of the peripheral groups and the unbalanced 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 unbalanced masses 19a, 19g and 19c, 19f of the peripheral groups and the unbalanced masses 19b, 19d, 19e of the middle group (excitation-free operation, idling).

[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, when superimposed with the excitation forces FR of the peripheral groups, are reduced or equalized depending on the setting (Fig. 9 and Fig. 10). For this purpose, the differential cage 34 of the epicyclic gear 30 is rotated during operation at the angular velocity OD via a rotary drive 39 coupled to the differential cage 34. With the known angular velocity at the gear input shaft 33, the angular velocity θ2 of the gear output shaft 35 can be determined using the Willis equation when the epicyclic gear 30 rotates in the opposite direction to the direction of rotation of the gear input shaft 33: 1 "2 = - ' ("1 - "D ' (1 - io)) lo For differential gears, the standard ratio io=-1 can be assumed: 602= 2 ■ ) D — 60 x

[0038] Fig. 9 shows the system of unbalanced masses 19a-19g with the three unbalance positions S1, S5, and S6 from Figures 6-8. In the upper image, the differential angle ß between the edge groups and the center group is zero. This differential angle ß is 90° in the center 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 Fen- are shown.

[0039] During normal operation, the differential cage 34 is at rest and OD=0, which means o2=-oi. The transmission output shaft 35 and the transmission input shaft 33 then have the same angular velocity but opposite directions of rotation. When the epicyclic gear 30 rotates, the transmission output shaft 35 moves faster than the transmission input shaft 33, with the difference corresponding to twice the angular velocity OD 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 epicyclic gear 30 and the corresponding unbalanced masses 19c, 19e.For example, if the gear ratio between the shafts 33, 35 of the epicyclic gear 30 and the unbalanced masses 19c, 19e is i=-1 (opposite direction of rotation at the same angular velocity), a rotation (and subsequent fixation) of the differential cage 34 by an angle ß / 2 results in an angular difference of ß 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 ß (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 results.

[0040] 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 ou in the center of rotation: F E = m - e ■ a>u

[0041] 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 Fpmax, FMmax of an unbalance mass group is reached precisely when all unbalance masses 19a-19g of the respective group are positioned either horizontally or vertically. Without rotation of the unbalance mass groups relative to each other (angular difference ß=0, unbalance mass position S1 in Fig. 6 left and S4 in Fig. 7 right), the excitation forces FE of the unbalance masses 19a, 19g and 19c, 19f of the outer groups and the unbalanced masses 19b, 19d, 19e of the middle group are completely synchronized, with the excitation force FM of the middle group being 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 Fmax (maximum possible impact force amplitude).

[0042] If the unbalanced masses 19a-19g are rotated relative to one another, the maximum excitation forces Fpmax, FMmax cannot be completely superimposed at any time (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 one another, for example, the oscillation generated by the middle group leads the oscillation generated by the edge groups (phase shift 40). This effect is illustrated in Fig. 10, where the oscillations of the middle group and the two synchronously running edge groups are depicted for two complete revolutions of the unbalanced masses 19a-19g, starting from the horizontal position of the edge group unbalanced masses 19a, 19c, 19f, 19g.

[0043] All three diagrams in Fig. 10 show the curves of the summed excitation force FR of the edge groups and the excitation force FM of the middle group as well as the resulting excitation force Fen- over a rotation angle a, whereby the initial position corresponds to the three unbalance positions S1, S5 and S6 in Fig. 9. The curve of the summed excitation force FR of the edge groups is shown with dotted lines, the curve of the excitation force FM of the middle group with dashed lines and the curve of the resulting excitation force F er r with a solid line. The phase shift 40 of the amplitude of the entire system relative to the amplitude of the peripheral groups or the amplitude of the central group corresponds to half the difference angle ß. This difference angle ß is 0° in the upper diagram, 90° in the middle diagram, and 180° in the lower diagram.

[0044] The oscillation of the entire system results from the superposition of the oscillations of the outer 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 summed excitation force FRmax of the two Marginal groups (FRmax=FMmax=F m ax / 2), the reduced (horizontal) excitation force F er r as a function of the orbit angle a and the difference angle ß as follows:

[0045] A reduced impact force Fred depending on the maximum impact force Fmax (maximum excitation force) is derived from this equation by means of an extreme value analysis and results from the difference angle ß between the edge group and the middle group: ß Fred F max ' COS(-) The reduction factor cos(ß / 2) is shown in Fig. 11, where the maximum impact force Fmax (maximum excitation amplitude) occurs at a difference angle ß of 0°. Specifically, Fig. 11 shows the curve of the reduction factor cos(ß / 2) over the difference angle ß between 0° and 180°.

[0046] A stabilization unit 11 according to the invention is preferably operated in pairs, as shown in Fig. 1. With two stabilization units 11 used one after the other, several combination options for ballast 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.

[0047] An advantage of the present invention with regard to the compaction effect lies in the low center of gravity 22 or the low effective plane 14, where the point of application of the horizontal force excitation is located. This enables a predominantly translational excitation of the track grid 4.

[0048] Until now, vertical force excitation for compacting the track ballast was only provided in the intermediate compartments and on the flanks of the ballast superstructure by means of intermediate compartment compactors and front-end compactors. The present invention additionally enables vertical excitation of the ballast beneath each sleeper. In this case, it is only necessary to ensure that the stabilization unit 11 does not lift off the rail heads in order to avoid damage (head checks, corrugation). For safe operation, the vertical load A is applied by means of the Height actuators 13 are set so high that the relieving effect of the centrifugal forces of the vibration generator 12 remains limited.

[0049] Since the vertical stiffness below the sleeper 5 is greater than in the horizontal direction, there is a stronger interaction between the rail vehicle 1, the track grid 4 and the ballast bed 3 during vertical operation. The machine parameters must therefore be carefully adapted to the local conditions in the case of purely vertical excitation, in particular to the condition of the track ballast, the geometry of the ballast bed 3 and the underlying subsoil.

[0050] An operating mode in which the front of the two stabilization units 11 is excited vertically and the rear of the two stabilization units 11 is excited horizontally in one direction of travel simulates a rail vehicle passing over the track during normal operation. During such a crossing in normal operation, a leading lift-off 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 positive effect on the durability of the preceding track geometry corrections.

[0051] The ability to compact the track ballast below the respective sleeper 5 using vertical force excitation, combined with a subsequent horizontal excitation, leads to improved compaction results. A key advantage of this operating mode is also compaction control. For this purpose, the vertical force excitation is selected to be so low that no compaction effect occurs. This allows statements to be made about the vertical stiffness without disturbing the structure of the ballast and thus the track geometry.

[0052] 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 curve from the track grid 4 can be determined from the difference. To assess the compaction success, a parameter from a corresponding work diagram (at constant excitation frequency) or via an impedance function (at variable excitation frequency, dynamic stiffness). An example of a work diagram is shown in Fig. 12. The abscissa indicates a vibration displacement 41 of the activated track grid area. The ordinate indicates a contact force 42 below the activated sleepers 5. From this work diagram, conclusions are drawn about the stiffness (relationship between defined force difference 43 and measured displacement 44 when the track grid 4 is loaded), the damping of the system (curvature of the curve) and the introduced energy 45 (circumscribed area). Dashed horizontal lines indicate a static load 46, a minimum vertical load 47 and a maximum vertical load 48.

[0053] A mechanical model of track deck 4 is used for compaction control. An optimization process is then used to derive deterministic parameters of track deck 4, which lead to the measured response for a given excitation with a known power spectral density. A measured value determined in this way has the advantage of being directly physically interpretable and serving as a basis for track maintenance planning.

[0054] Fig. 13 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 rotation-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 illustrated position, the coupling element 26, designed as a reversing pin, lies in an upper driver recess 50 of the unbalanced mass 19b. As soon as the direction of rotation of the gear 49 changes, the gear 49 rotates by 180° relative to the unbalanced mass 19b until the reversing bolt rests in a lower driving recess 51 of the unbalanced mass 19b.

[0055] To prevent unwanted turning of the unbalanced mass 19b, a centrifugal lock 52 is arranged. Fig. 14 shows this detail in a side view. A lever 53 is assigned to both the upper driver recess 50 and the lower driver recess 51. Each lever 53 is rotatably mounted at one end on the unbalanced mass 19b. When the unbalanced mass 19b is stationary or at low speed, the respective lever 53 is pressed inward by an associated spring 54. In this state, the driver recesses 50, 51 are free to receive the coupling element 26. As the speed increases, centrifugal force pushes the two levers 53 outward. 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 expediently arranged to monitor the current position of the unbalanced mass 19b.

[0056] 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 transmission output shaft 35 of the epicyclic gear 30. A further gear 56 is arranged on the drive shaft 20a with a further coupling element 26 dependent on the direction of rotation. 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 transmission input shaft 33 of the epicyclic gear 30. These transmission 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.

[0057] To implement the method according to the invention, it is expedient if the stabilization units 11 arranged on the rail vehicle 1 are controlled by 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 curve of the track grid 4.

Claims

Patent claims 1 . Stabilization unit (11) for stabilizing a track (2), with a vibration generator (12) comprising mutually parallel rotating shafts (20, 20a, 20b) with unbalanced masses (19, 19a-19g) for generating an impact force (F, F max, Fred) with an adjustable direction, and with flanged wheels (15) and pressure rollers (16) for transmitting the impact force (F, F max, Fred) to a track grid (4) of the track to be stabilized, consisting of sleepers (5) and rails (6) fastened thereto, wherein each flanged wheel (15) is rotatably mounted about a wheel axis (17) and has a running surface (23) with a wheel diameter (d), characterized in that the rotating shafts (20, 20a, 20b) are arranged in a horizontal effective plane (14) in such a way as to generate the impact force (F, F max, Fred), that the horizontal plane of action (14) of the impact force (F, Fmax, Fred) is not more than 300 millimetres,in particular, no more than 260 millimetres above a rolling plane (25) of the flanged wheels (15).

2. Stabilization unit (11) according to claim 1, characterized in that the horizontal effective plane (14) lies less than half a wheel diameter (d / 2) above a horizontal plane (18) laid through the respective wheel axle (17).

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

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

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

6. Stabilization unit (11) according to claim 5, characterized in that the at least one unbalanced mass (19b) is coupled to the associated rotary shaft (20a) by means of a centrifugal lock (52).

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

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

9. Stabilization 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 an epicyclic gear (30).

10. Stabilization unit (11) according to claim 9, characterized in that a cage (34) of the epicyclic gear (30) is rotatably mounted and coupled to a rotary drive (39).

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

12. Rail vehicle (1) with a machine frame (9) which is supported on rail bogies (10) and can be moved on a track (2), characterized in that at least two stabilization units (11) according to one of claims 1 to 11 are arranged in such a way that a front stabilization unit (11) is fastened to the machine frame (9) with first height adjustment drives (13) and that a rear stabilization unit (11) is fastened to the machine frame (9) with second height adjustment drives (13).

13. Rail vehicle (1) according to claim 12, characterized in that the vibration generators (12) and the height adjustment drives (13) are controlled by means of a common control device (58) and that the control device (58) is designed to separately control the respective vibration generator (12) and the respective height adjustment drive (13).

14. Method for operating a rail vehicle (1) according to claim 12 or 13, characterized in that the rail vehicle (1) is moved forward, that the front stabilization unit (11) is operated with a vertical impact force (F, F max, Fred) and that the rear stabilization unit (11) is operated with a horizontal impact force (F, Fmax, Fred).

15. Method according to claim 14, characterized in that vertical accelerations are detected at the front stabilization unit (11) by means of an acceleration sensor (59) in order to derive therefrom a reaction force curve of the track grid (4).

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