Stabilizing assembly and method for stabilizing track

Abstract

The invention relates to a stabilizing assembly (11) for stabilizing a rail (2), comprising: a vibration generator (12) having rotational axes (20, 20a, 20b), which are oriented parallel to one another and have unbalanced masses (19, 19a-19g) for generating a direction-adjustable impact force (F, Fmax, Fred); and flanged wheels (15) and squeeze rollers (16) for transmitting impact forces (F, Fmax, Fred) to a track panel (4) of the track (2) to be stabilized, the track panel consisting of sleepers (6) and rails (6) fastened thereto, each flanged wheel (15) being rotatably supported about an axle (17) and having a running surface (23) with a wheel diameter (d). The rotating shaft (20, 20a, 20b) is arranged to generate an 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) is located up to 300 mm, in particular up to 260 mm, above the rolling plane (25) of the flanged wheel (15). The low level effective plane (14) prevents disturbing tilting torques during the stabilization process.

Description

Field of the Invention

[0001] The present invention relates to a stabilizing assembly for stabilizing a track, the stabilizing assembly having a vibration generator including a rotating shaft and unbalanced mass blocks aligned parallel to each other, the unbalanced mass blocks being for generating impact forces with adjustable directions, and the stabilizing assembly having flange wheels and squeeze rollers for transmitting the impact forces to a track panel of the track to be stabilized, the track panel being composed of sleepers and rails fixed to the sleepers, each flange wheel being supported to rotate about a wheel axis and having a running surface with a wheel diameter. The present invention also relates to a rail vehicle having such a stabilizing assembly and a method for operating such a rail vehicle. Background Art

[0002] Ballasted tracks are continuously stressed by rail traffic and environmental effects. For example, the position of the track panel in the ballast bed changes. Over time, the ballast bed itself becomes contaminated due to wear and the introduction of foreign matter. Maintenance measures such as compaction processes or cleaning processes eliminate these problems. However, this causes the ballast bed to become temporarily loose. Subsequent settlement may occur even after optimal compaction by a compaction unit. To control such settlement, a machine is used to stabilize the track.

[0003] The machine can move along the track and includes a stabilizing assembly that is clamped to the rails of the track by unit rollers. A vibration generator arranged on the stabilizing assembly generates vibrations that are transmitted to the track panel. The design and dimensions of the vibration generator determine the impact force and the vibration frequency acting on the track. The stabilizing assembly is supported against a frame to generate a static applied load. The transmitted vibrations make the ballast in the ballast structure of the ballast bed movable, so that the ballast is displaced and rearranges itself with a higher degree of compaction. This optimized ballast compaction improves the load-bearing capacity of the track and the lateral track resistance.

[0004] AT 16604 U1 discloses an exemplary stabilizing assembly with variable impact forces. In this document, the vibration generator includes a plurality of rotating unbalanced mass blocks arranged on axially aligned shafts. The unbalanced mass blocks are driven with a phase shift that can be variably adjusted relative to each other. Based on the arrangement of the unbalanced mass blocks, the changed phase shift changes both the direction and the intensity of the impact force. Summary of the Invention

[0005] It is an object of the present invention to improve a stabilizing assembly of the above kind in such a way that the impact force acts on the track in an optimized manner. Furthermore, it is an object of the present invention to provide a rail vehicle that takes advantage of the extended application possibilities of the improved stabilizing assembly. In addition, an advantageous method for operating such a rail vehicle is provided.

[0006] According to the present invention, these objects are achieved by the features of independent claims 1, 12, and 14. The dependent claims provide advantageous designs of the present invention.

[0007] With the aid of this novel stabilizing assembly, the rotational axis is arranged to generate the impact force in the horizontal plane of action in such a way that the horizontal plane of action of the impact force lies at most 300 mm, in particular at most 260 mm, above the rolling plane of the flange wheel. The lowering of this horizontal plane of action prevents a destructive tilting moment during the stabilization process. In use, the rolling plane of the flange wheel corresponds to the plane of the track to be stabilized spanned by the top of the rail. If the horizontal impact force is at most 260 mm above the top of this rolling plane or the rail plane, the case where the sleeper rests on the saddle-shaped surface can be safely excluded. This also applies to the maximum value of 300 mm, where there is a greater free space below this stabilizing assembly for arranging a chord measuring system or an optical measuring system.

[0008] Advantageously, the horizontal plane of action lies less than half the wheel diameter above the horizontal plane passing through the respective wheel axis. The vibration generator is arranged there at a correspondingly low level, where the wheel diameter is large enough such that no destructive pressure peaks occur on the rail surface. The flange wheels are spaced relatively far apart from one another such that there is construction space for the vibration generator therebetween. This also applies to the elements that extend the wheel axle and that press the flange wheels against the rail during operation. In conventional stabilizing assemblies, the vibration generator is always arranged in the region above the flange wheel, from which a high horizontal plane of action of the impact force is generated. The resulting tilting moment can, in severe cases, lead to a saddle-shaped support of the sleeper in the intermediate ballast layer of the track.

[0009] Advantageously, at least two rotational axes and / or unbalanced mass blocks are coupled to a gear drive element and driven by a common drive. In this way, all rotational axes or unbalanced mass blocks can be driven using an optimally controlled common drive. The type of coupling determines the manner in which the centrifugal force caused by the unbalanced mass block generates the impact force. Preferably, the centrifugal force increases in the desired plane of action while the centrifugal forces in other planes of action cancel each other out.

[0010] In a further refinement, at least one unbalanced mass block is rotatably supported on each rotational axis. Compared to an unbalanced mass block fixed to the rotational axis, this unbalanced mass block can be driven using changes in the angular position, rotational speed, and direction of rotation. This allows for adjustment of the direction and magnitude of the resulting centrifugal force.

[0011] Preferably, at least one unbalanced mass is coupled to a specified axis of rotation by a coupling element that depends on the direction of rotation in such a way that when the direction of rotation changes, the unbalanced mass rotates relative to the axis of rotation, in particular by 180°. Together with the unbalanced mass permanently arranged on the axis of rotation, the at least one unbalanced mass generates two different centrifugal forces depending on the direction of rotation. This means that the stabilizing assembly can operate with different impact forces at the same vibration frequency.

[0012] In a further refinement of this variant, the at least one unbalanced mass is coupled to the specified axis of rotation by a centrifugal locking mechanism. Once a specified speed is exceeded, the centrifugal locking mechanism immediately locks the unbalanced mass to the specified axis of rotation. This ensures that during continuous operation, the unbalanced mass does not undergo an unwanted reversal.

[0013] An advantageous design with a low center of gravity includes a central axis of rotation parallel to the longitudinal direction of the assembly and lateral axes of rotation located on the left and right sides of the central axis of rotation. This results in a symmetric design with different drive variants, which largely prevents destructive tilting moments during operation.

[0014] If the design is refined, the degree of imbalance of the unbalanced mass assigned to the central axis of rotation is twice as large as that of the unbalanced mass assigned to the corresponding lateral axis of rotation. In this way, the impact force can be adjusted without a separation step from scratch.

[0015] A further refinement provides that the directly driven unbalanced masses are coupled to a common drive, while the indirectly driven unbalanced masses are coupled to the directly driven unbalanced masses via a planetary gear train. The combined centrifugal force effect generated by all the unbalanced masses can be adjusted via the planetary gear train.

[0016] Here, the cage of the planetary gear train is advantageously rotatably supported and coupled to the rotary drive. By means of the rotary drive, the cage can be set to rotate, thereby changing the relative angular velocity between the directly driven unbalanced masses and the indirectly driven unbalanced masses.

[0017] A further preferred refinement of the stabilizing assembly includes an acceleration sensor for recording the acceleration caused by the vibration generator. The movement of the stabilizing assembly or the movement of the track arrangement set to vibrate is recorded in order to draw conclusions about the reaction forces of the track arrangement.

[0018] A rail vehicle according to the present invention includes: a frame that can move on a track when located on a running gear on the track; and at least two of the above-mentioned stabilizing components, wherein a front stabilizing component having a first height actuator is fastened to the frame, and a rear stabilizing component having a second height actuator is fastened to the frame. In this way, the stabilizing components can operate independently of each other under different applied loads and different impact forces.

[0019] Advantageously, the vibration generator and the height actuator are actuated by a common control device, wherein the control device is arranged to individually actuate the corresponding vibration generator and the corresponding height actuator. The two stabilizing components can be coordinated in operation by a common control unit. For example, synchronous vibrations are applied to the track panel.

[0020] In a method for operating a rail vehicle according to the present invention, it moves forward along the track to be stabilized, wherein the front stabilizing component operates with a vertical impact force, and the rear stabilizing component operates with a horizontal impact force. This operating mode is used to simulate the driving conditions of a normally operating rail vehicle, because the undulation wave (Abhebewelle) generated in front of the rail vehicle usually causes a sinusoidal movement of the running gear of the rail vehicle on the track. The rail vehicle according to the present invention controls these processes and, in this way, leaves a particularly sustainably stabilized track.

[0021] A further improvement of the method is to use an acceleration sensor of the front stabilizing component to record the vertical acceleration through which the reaction force trend of the track panel is obtained. Specifically, the measured acceleration proportional to the force and the known force from dynamic excitation are used to determine the corresponding reaction force. Description of the Drawings

[0022] The present invention is explained below in an exemplary manner with reference to the accompanying drawings. It is schematically shown as follows:

[0023] Figure 1 is a rail vehicle with stabilizing components;

[0024] Figure 2 is a cross-sectional view of a track with stabilizing components;

[0025] Figure 3 is a flange wheel and a squeezing roller in contact with the rail;

[0026] Figure 4 is a top view and a cross-sectional view of a stabilizing component having three mechanically coupled rotating shafts;

[0027] Figure 5 is according to Figure 4 a change in the rotation direction of the rotating shaft of the stabilizing component;

[0028] Figure 6 is the stabilizing component with the maximum horizontal force excitation according to Figure 4 ;

[0029] Figure 7 is the stabilizing component with the maximum vertical force excitation according to Figure 4 ;

[0030] Figure 8 is the stabilizing component during the rotation of the planetary gear transmission according to Figure 4 ;

[0031] Figure 9 is the stabilizing component with different impact force settings according to Figure 4 ;

[0032] Figure 10 is Figure 4 the impact force curve of the stabilizing component under different driving conditions;

[0033] Figure 11 is Figure 4 the reduction coefficient of the vibration amplitude of the stabilizing component;

[0034] Figure 12 is the working drawing of the track panel area set to be in vibration;

[0035] Figure 13 is the drive shaft with an unbalanced mass block;

[0036] Figure 14 is the unbalanced mass block with a centrifugal force locking mechanism. Detailed implementation mode

[0037] Figure 1 The rail vehicle 1 shown in

[0038] Figure 2A cross-sectional view of the track 2 with the stabilizing assembly 11 during the stabilization process is shown. The stabilizing assembly 11 includes two independent main components, namely a vibration generator 12 and a pair of height actuators 13 (load-applying hydraulic cylinders). The vibration generator 12 alternately generates impact forces F in two opposite directions in the acting plane 14, which causes the stabilizing assembly 11 to vibrate. Preferably, the impact force F acts in the horizontal plane. This horizontal acting plane 14 is crucial for the present invention. However, for the extended operation of the stabilizing assembly 11, the influence of the impact force F in the vertical direction also plays a role. Therefore, the acting plane 14 is a vertical plane.

[0039] The flange wheels 15 and the squeeze rollers 16 transfer the vibration to the track panel 4. Each flange wheel 15 is supported to rotate about the wheel axis 17 and is guided along the inner edge of the rail. The wheel axis 17 lies in a common horizontal plane 18. The squeeze rollers 16 press against the rail 6 from the outside. The applied load A is applied by the height actuators 13, and the applied load A can be adjusted without a separation step.

[0040] Advantageously, the stabilizing assembly 11 includes a self-supporting central part with the vibration generator 12. The vibration generator 12 includes an unbalanced mass 19 supported on a rotating shaft 20. Viewed in the longitudinal direction of the track, the lateral frames are connected to the central part on each side. The central part is connected to the respective lateral frames, for example, by threaded connections on the circumferential flange. The flange wheels 15 and the squeeze rollers 16 are only supported on the designated lateral frames. To achieve the extended wheel axle 21, the flange wheels 15 assigned to one of the lateral frames are each connected to a hydraulic drive to enable displacement, for example, along the designated wheel axis 17. The front flange wheels 15 or the rear flange wheels 15 do not have a common through shaft. The lack of a through shaft creates space for a low setting of the central part. Therefore, the center of gravity 22 of the entire stabilizing assembly 11 is low and the acting plane 14 of the vibration generator 12 is low. Preferably, the center of gravity 22 lies in the horizontal acting plane 14.

[0041] Each flange wheel 15 has a wheel diameter d which is measured on the running surface 23. In use, the running surface 23 of the flange wheel 15 contacts the rail top 24 of the rail 6. Here, the rail top 24 is the highest line on the rail head. The lower tangential plane and the upper tangential plane contact all the running surfaces 23 of the flange wheel 15. The lower tangential plane forms a rolling plane 25, and the contact point between the running surface 23 of the flange wheel 15 and the rail top 24 lies in this rolling plane 25 during use. According to the invention, the vibration generator 12 is arranged so low that the vertical distance a between the horizontal action plane 14 of the impact force F and the rolling plane 25 is at most 300 mm, in particular at most 260 mm. Very good results were achieved in tests with the vertical distance a = 250 mm. Even in the case of poor ballast compaction conditions, the sleeper does not rest on the saddle surface.

[0042] Advantageously, the horizontal action plane 14 lies less than half of the wheel diameter d / 2 of the corresponding flange wheel 15 above the horizontal plane 18 passing through the corresponding wheel axis 17. The upper tangential plane of the running surface 23 forms a boundary plane, and the horizontal action plane 14 lies below this boundary plane. If the horizontal action plane 14 lies above the wheel axis 17, the further vertical distance b between this action plane 14 and the horizontal plane 18 is less than half of the wheel diameter d / 2 of the corresponding flange wheel 15. In any case, this property is satisfied if the horizontal action plane 14 is below the wheel axis 17. This brings two advantages. On the one hand, the horizontal action plane 14 is low enough, and on the other hand, the wheel diameter d of the flange wheel 15 is large enough so that no destructive pressure peaks occur on the rail surface.

[0043] Referring to Figures 4 to 11 , a favorable form of the vibration generator 12 with reduced construction height is described. To achieve a low-lying force excitation, cylindrical unbalanced masses 19 are arranged, which rotate around an axis aligned in the longitudinal direction 7. The unbalanced masses 19 are divided into three groups in the axial direction so that the amplitude of the effective impact force F can be adjusted without a separation step. In the example shown, most of the unbalanced masses 19a, 19b, 19c, 19d, 19e are supported to rotate freely on the driven central rotation axis (drive shaft) 20a or the coupled lateral rotation axis (auxiliary shaft) 20b. The other unbalanced masses 19f, 19g are permanently connected to the specified rotation axes 20a, 20b.

[0044] The unbalanced mass blocks 19a, 19b, 19c supported on the drive shaft 20a have coupling elements 26, which are related to the direction of rotation, and through which the unbalanced mass blocks are connected to the corresponding drive mechanisms. One such drive mechanism is, for example, a cylindrical gear permanently mounted on the drive shaft 20a, which has corresponding recesses for the reversing pins. By changing the direction of rotation of the drive shaft 20a, the unbalanced mass blocks 19d, 19e, 19f, 19g on the auxiliary shaft 20b rotate 180° relative to their initial positions, while the unbalanced mass blocks 19a, 19b, 19c on the drive shaft 20a remain in their positions unchanged. This principle is shown in Figure 5 in three chronologically consecutive stages 27, 28, 29. In the first stage 27, the unbalanced mass blocks 19a - 19g are in their initial positions for horizontal force excitation, where only one unbalanced mass block 19b of the central group is shown. All the unbalanced mass blocks 19a - 19g point to the right. The second stage 28 shows the reversal process, and the third stage 29 shows the unbalanced positions for vertical force excitation.

[0045] Compared with the unbalanced mass blocks 19a, 19c, 19f, 19g of the two edge groups, the degree of unbalance (product of mass and eccentricity, U = m·e) of the unbalanced mass blocks 19b, 19d, 19e of the central group is twice that. Additionally, the degree of unbalance of the unbalanced mass blocks 19a, 19b, 19c on the drive shaft 20a is twice that of the unbalanced mass blocks 19a, 19c, 19f, 19g on the associated auxiliary shaft 20b. For example, the degree of unbalance of the unbalanced mass blocks 19a, 19b, 19c is twice that of the unbalanced mass blocks 19d, 19e, 19f, 19g, where the unbalanced mass block 19d has the same degree of unbalance as the unbalanced mass block 19e, and the unbalanced mass block 19f has the same degree of unbalance as the unbalanced mass block 19g. This arrangement enables the vibration generator 12 of the entire system to achieve an adjustable impact force amplitude without a separation step, while the centrifugal forces cancel each other out in one direction.

[0046] The unbalanced mass blocks 19a - 19g are driven by the drive shaft 20a and the planetary gear transmission 30. Figure 4Shows the rotational movement 31 for horizontal excitation and the rotational movement 32 for vertical excitation of the drive shaft 20a. The rotational movements 31, 32 of the drive shaft 20a are directly transmitted to the unbalanced mass blocks 19a, 19c on the drive shaft 20a. Subsequently, the rotational movement of the unbalanced mass block 19c is transmitted via a cylindrical gear to the adjacent unbalanced mass block 19f, which also drives the unbalanced mass block 19g rigidly connected to the corresponding auxiliary shaft 20b. Additionally, the rotational movement is further transmitted via a cylindrical gear on the unbalanced mass block 19c of the drive shaft 20a to the gear drive input shaft 33 of the planetary gear transmission 30. During normal operation, the differential cage 34 of the planetary gear transmission 30 remains stationary, which transmits the rotational movement of reverse rotation at the same rotational speed to the unbalanced mass block 19e of the auxiliary shaft 20b via the gear drive output shaft 35. The other unbalanced mass blocks 19b, 19d of the central group are driven via connection with this unbalanced mass block 19e.

[0047] Figure 4 The uppermost image shows the differential cage 34 with differential bolts 36, constant velocity bevel gears 37, and shaft bevel gears 38. The central image shows a top view of the assembly 11, where a section line passes through the unbalanced mass blocks 19b, 19d, 19e of the central group, resulting in a cross-sectional view of the assembly 11 in the lowermost image.

[0048] Figure 6 Shows the system of unbalanced mass blocks 19a - 19g for maximum possible horizontal force excitation. On the left, the unbalanced position S1 in the case of the maximum horizontal impact force F max is shown. On the right, the system is further rotated by a rotation angle α = 90° to the unbalanced position S2, where the resultant excitation force F err is zero.

[0049] Figure 7 Shows the operating mode of the system of unbalanced mass blocks 19a - 19g for maximum possible vertical force excitation. The left unbalanced position S3 shows the system where the resultant excitation force F err is zero. On the right, the system is further rotated by a rotation angle α = 90° to the unbalanced position S4, where the resultant excitation force F err is zero.

[0050] Figure 8Shows two unbalance positions S5, S6 of the system of unbalance mass blocks 19a - 19g when the planetary gear transmission 30 rotates. On the left side, the unbalance position S5 is shown for the 90° angular difference generated between the unbalance mass blocks 19a, 19g and 19c, 19f of the edge group and the unbalance mass blocks 19b, 19d, 19e of the central group, and on the right side, the unbalance position S6 is shown for the 180° angular difference generated between the unbalance mass blocks 19a, 19g and 19c, 19f of the edge group and the unbalance mass blocks 19b, 19d, 19e of the central group (non-excited operation, idling).

[0051] To reduce the amplitude of the effective impact force F of the system, the unbalance mass blocks 19b, 19d, 19e of the central group rotate relative to the other unbalance mass blocks 19a, 19c, 19f, 19g such that the excitation force F M of the central group R is reduced or becomes equal when superimposed with the excitation force F Figure 9 and Figure 10 ) of the edge group, depending on the setting( D ). For this purpose, the differential cage 34 of the planetary gear transmission 30 rotates at an angular velocity ω

[0052]

[0053] during continuous operation via a rotary drive 39 coupled to the differential cage 34. Given the angular velocity ω1 of the input shaft 33 of the known gear transmission, when the planetary gear transmission 30 rotates in a direction opposite to the rotation direction of the gear transmission input shaft 33, the angular velocity ω2 of the gear transmission output shaft 35 can be determined using the Willis equation:

[0054] ω2 = 2·ω D - ω1

[0055] Figure 9 Shows the system of unbalance mass blocks 19a - 19g in three unbalance positions S1, S5 and S6 Figures 6 to 8 . In the upper image, the angular difference β between the edge group and the central group is zero. This angular difference β is 90° in the central image and 180° in the lower image. On the right side of each group of unbalance mass blocks 19a - 19g, the force polygon of the corresponding excitation force F E and the corresponding resultant excitation force F err are shown.

[0056] During normal operation, the differential cage 34 is stationary and ω D= 0, which means ω2 = -ω1. At this time, the output shaft 35 of the gear transmission device and the input shaft 33 of the gear transmission device have the same angular velocity but opposite rotation directions. When the planetary gear transmission device 30 rotates, the output shaft 35 of the gear transmission device moves faster than the input shaft 33 of the gear transmission device, where the difference corresponds to twice the angular velocity ω of the rotating differential cage 34. D The rotation between the unbalanced mass blocks 19b, 19d, 19e of the central group and the unbalanced mass blocks 19a, 19c, 19f, 19g of the edge group is caused by the transmission between the shafts 33, 35 of the planetary gear transmission device 30 and the corresponding unbalanced mass blocks 19c, 19e. For example, if the transmission ratio between the shafts 33, 35 of the planetary gear transmission device 30 and the unbalanced mass blocks 19c, 19e is, for example, i = -1 (rotating in opposite directions at the same angular velocity), then the rotation (and subsequent fixation) of the differential cage 34 by an angle β / 2 will result in an angular difference β between the unbalanced mass blocks 19b, 19d, 19e of the central group and the unbalanced mass blocks 19a, 19c, 19f, 19g of the edge group. This angular difference β (phase shift) causes a reduction in the magnitude of the effective horizontal impact force F during horizontal operation, or a reduction in the magnitude of the effective vertical impact force F during vertical operation.

[0057] The exciting force F E (centrifugal force) of a single unbalanced mass block 19 is obtained by the product of the mass m, the eccentricity e, and the square of the angular velocity ω of the center of rotation U :

[0058]

[0059] During horizontal operation, any vertical components of the corresponding unbalanced mass block groups cancel each other out (for example, the vertical components of the unbalanced mass blocks 19a and 19g cancel each other out), which precisely achieves the maximum exciting force F of the unbalanced mass block group when all the unbalanced mass blocks 19a - 19g of the corresponding group are in the horizontal or vertical state. Rmax 、F Mmax 。 Without rotating the unbalanced mass block groups towards each other (angular difference β = 0, Figure 6 the position S1 of the unbalanced mass block on the left and Figure 7 the position S4 of the unbalanced mass block on the right), the exciting forces F E of the unbalanced mass blocks 19a, 19g and 19c, 19f of the edge group and the unbalanced mass blocks 19b, 19d, 19e of the central group are completely added together, where the exciting force F M of the central group and the exciting forces F R of the two synchronously operating edge groups have the same magnitude. In this case, the system has the maximum combined impact force F max(Maximum possible impact force amplitude) operation.

[0060] If the unbalanced mass blocks 19a - 19g rotate relative to each other, the maximum excitation force F Rmax 、F Mmax can never be fully superimposed at any time ( Figure 9 and Figure 10 ), and the impact force F red of the system decreases. Since the unbalanced mass blocks 19a - 19g rotate towards each other, the vibration generated by the central group, for example, precedes the vibration generated by the edge group (phase shift 40). This effect is shown in Figure 10 , where the vibrations of the central group and two synchronously operating edge groups are shown for two complete rotations of the unbalanced mass blocks 19a - 19g (starting from the horizontal positions of the edge group unbalanced mass blocks 19a, 19c, 19f, 19g).

[0061] Figure 10 All three diagrams in show the total excitation force F R of the edge group, the excitation force F M of the central group, and the curve of the resultant excitation force F err over the rotation angle α, where the initial positions correspond to Figure 9 the three unbalanced positions S1, S5, and S6 in . The trend of the total excitation force F R of the edge group is plotted with a dotted line, the trend of the excitation force F M of the central group is plotted with a dashed line, and the trend of the resultant excitation force F err is plotted with a solid line. The phase shift 40 of the amplitude of the entire system to the amplitude of the edge group or the amplitude of the central group corresponds to half of the angle difference β. This angle difference β is 0° in the upper diagram, 90° in the central diagram, and 180° in the lower diagram.

[0062] The vibration of the entire system is generated by the superposition of the vibrations of the edge group and the central group. Since the horizontal or vertical excitation force F Mmax with the maximum amplitude of the central group has the same magnitude as the total excitation force F Rmax of the two edge groups (F Rmax = F Mmax = F max / 2), the reduced (horizontal) excitation force F err can be described as a function of the rotation angle α and the angle difference β as follows:

[0063]

[0064] Through extreme value analysis, the reduced impact force F max as a function of the maximum impact force F red, and it is obtained as a function of the angular difference β between the edge group and the central group:

[0065]

[0066] The reduction factor cos(β / 2) is shown in Figure 11 where the maximum impact force F max (maximum excitation amplitude) occurs at an angular difference β of 0°. Specifically, Figure 11 shows the trend of the reduction factor cos(β / 2) over the angular difference β between 0° and 180°.

[0067] The stabilizing assembly 11 according to the invention preferably operates in pairs, as Figure 1 shown. In the case of using two stabilizing assemblies 11 in series, the variable excitation directions result in various possible combinations for ballast compaction: both assemblies 11 operate horizontally simultaneously, both assemblies 11 operate vertically simultaneously, or one assembly 11 operates vertically while the other assembly 11 operates horizontally.

[0068] In terms of the compaction effect, the advantages of the present invention are that the center of gravity 22 is low or the action plane 14 where the horizontal force excitation application point is located is low. This enables mainly translational excitation of the track arrangement 4.

[0069] So far, vertical force excitation for compacting track ballast has only been provided in the sleeper boxes and on the shoulders of the ballast superstructure by sleeper box stiffeners and sleeper end stiffeners. Additionally, the present invention also enables vertical excitation of the ballast under each sleeper. Here, the only important thing is to ensure that the stabilizing assembly 11 does not detach from the rail head to avoid damage (rail head inspection, corrugation formation (Riffelbildung)). For safe operation, the vertically applied load A is set very high by the height actuator 13 so that the alleviating effect of the centrifugal force of the vibration generator 12 is still limited.

[0070] Since the vertical stiffness under the sleeper 5 is greater than the stiffness in the horizontal direction, the interaction between the rail vehicle 1, the track arrangement 4, and the ballast bed 3 is greater during vertical operation. Therefore, in the case of full vertical excitation, the machine parameters must be adjusted particularly carefully according to local conditions, especially according to the condition of the track ballast, the geometry of the ballast bed 3, and the existing subsoil.

[0071] The operating mode in which the front stabilizing component among the two stabilizing components 11 is vertically excited in a traveling direction and the rear stabilizing component among the two stabilizing components 11 is horizontally excited in a traveling direction simulates the driving condition of a rail vehicle on a track during normal operation. During such traveling on a track in normal operation, an antecedent undulation wave (Abhebewelle) (vertical excitation) usually appears in front of the rail vehicle, followed by a sinusoidal oscillation (horizontal excitation). Since this load from the compaction process thus approximates the subsequent load of rail traffic, it has a favorable impact on the persistence of the previous track position correction.

[0072] Through vertical force excitation, combined with the subsequent horizontal excitation, better compaction effect can be achieved by compacting the track ballast under the corresponding sleeper 5. Another main advantage of this operating mode is compaction control. For this purpose, the vertical force excitation is selected to be very low so that no compaction effect is generated. In this way, a judgment on the vertical stiffness can be made without disturbing the ballast structure and thus without disturbing the track geometry.

[0073] To determine the degree of ballast compaction, an acceleration signal needs to be measured on the stabilizing component 11, as described in AT 521481 A4. Since the measured acceleration is proportional to the force and the force generated from the dynamic excitation is known, the reaction force trend of the track panel 4 can be determined according to the difference. To evaluate whether the compaction is successful, a parameter is then obtained from the corresponding working diagram (using a constant excitation frequency) or via the impedance function (using a variable excitation frequency, dynamic stiffness). Figure 12 An example of the working diagram is shown. The vibration path 41 of the activated track panel area is shown on the abscissa. The ordinate represents the contact force 42 under the activated sleeper 5. From this working diagram, the stiffness (the relationship between the defined force difference 43 and the measured displacement 44 when loading the track panel 4), the damping of the system (the curvature of the curve), and the applied energy 45 (the defined area) can be inferred. The horizontal dash line represents the statically applied load 46, the minimum vertically applied load 47, and the maximum vertically applied load 48.

[0074] Compaction control is based on the mechanical modeling of the track panel 4. Then, an optimization process is used to derive the deterministic parameters of the track panel 4, so as to obtain the measured response to a given excitation with a known power spectral density. The advantage of the measured variables determined in this way is that they can be directly physically interpreted and can also be used as the basis for planning track maintenance.

[0075] Figure 13Shows an improved form of the drive shaft 20a with unbalanced masses 19a, 19b, and 19c. Two outer unbalanced masses 19a and 19c are permanently connected to the drive shaft 20a. The central unbalanced mass 19a is supported to rotate on the drive shaft 20a and is coupled to the gear 49 (Zahnrad) via a coupling element 26 related to the direction of rotation. Here, the position of the central unbalanced mass 19b relative to the gear 49 depends on the direction of rotation. In the shown position, the coupling element 26 designed as a reversing pin is located in the upper drive recess 50 of the unbalanced mass 19b. Once the direction of rotation of the gear 49 changes, the gear 49 rotates 180° relative to the unbalanced mass 19b until the reversing pin rests in the lower drive recess 51 of the unbalanced mass 19b.

[0076] A centrifugal force locking mechanism 52 is arranged to prevent an unwanted reversal of the unbalanced mass 19b. Figure 14 This detail is shown in a side view. Rods 53 are assigned to the upper drive recess 50 and the lower drive recess 51. One end of the corresponding rod 53 is supported to rotate on the unbalanced mass 19b. When in a stationary state or when the unbalanced mass 19b is at a low rotational speed, the corresponding rod 53 is pressed inward by means of a designated spring 54. In this condition, the drive recesses 50, 51 freely receive the coupling element 26. As the speed increases, the centrifugal force pushes the two rods 53 outward. One of the rods 53 engages in the groove 26 of the coupling element 55, which locks the position of the unbalanced mass 19b relative to the gear 49. A sensor is effectively arranged to monitor the corresponding position of the unbalanced mass 19b.

[0077] In the shown embodiment variant, the unbalanced mass 19b is coupled via the gear 49 to two unbalanced masses 19d, 19e of the central group and the gear transmission output shaft 35 of the planetary gear transmission 30. Using other coupling elements 26 related to the direction of rotation, other gears 56 are arranged on the drive shaft 20a. This other gear 56 rotates 180° relative to the drive shaft 20a when the direction of rotation changes and couples the drive shaft 20a to two auxiliary shafts 20b and the gear transmission input shaft 33 of the planetary gear transmission 30. Thereby, these gear transmission elements 30, 49, 56 couple all the rotating shafts 20a, 20b and the unbalanced masses 19a - 19g together, where the drive shaft 20a is connected to a common drive 57.

[0078] For carrying out the method according to the invention, it is useful to actuate the stabilisation assembly 11 arranged on the rail vehicle 1 by means of a common control device 58. Here, the control device 58 is set up to actuate the vibration generator 12 and the height actuator 13 of the respective stabilisation assembly 11 individually. Preferably, an acceleration sensor 59 arranged on the front stabilisation assembly 11 transmits an acceleration signal to the control device 58 in order to subsequently evaluate the reaction force trend of the track panel 4.

Claims

1. A stabilizing assembly (11) for stabilizing an orbit (2), the stabilizing assembly (11) having a vibration generator (12), the vibration generator including rotation axes (20, 20a, 20b) aligned parallel to each other, the rotation axes having unbalanced mass blocks (19, 19a - 19g), the unbalanced mass blocks (19, 19a - 19g) being configured to generate impact forces (F, F max , F red ) with adjustable directions, and the stabilizing assembly having a flange wheel (15) and a squeezing roller (16), the flange wheel (15) and the squeezing roller (16) being configured to transfer the impact forces (F, F max , F red ) to a track panel (4) of the orbit (2) to be stabilized, the track panel being composed of sleepers (5) and rails (6) fixed to the sleepers, wherein each flange wheel (15) is supported to rotate about a wheel axis (17) and has a running surface (23) with a wheel diameter (d), characterized in that, The rotation axes (20, 20a, 20b) are arranged to generate the impact forces (F, F max , F red ) in the horizontal action plane (14) in such a way that: the horizontal action plane (14) of the impact forces (F, F max , F red ) is located at most 300 mm, in particular at most 260 mm, above the rolling plane (25) of the flange wheel (15).

2. The stable component (11) according to claim 1, characterized in that, The horizontal action plane (14) is located at a distance of less than half of the wheel diameter (d / 2) above the horizontal plane (18) passing through the respective wheel axis (17).

3. The stable component (11) according to claim 1 or 2, characterized in that, At least two rotating shafts (20, 20a, 20b) and / or unbalance mass blocks (19, 19a - 19g) are coupled to the gear drive elements (30, 49, 56) and are driven by a common drive (57).

4. The stabilizing component (11) according to any one of claims 1 to 3, characterized in that, At least one unbalance mass block (19, 19a - 19g) is rotatably supported on each rotating shaft (20, 20a, 20b).

5. The stable component (11) according to claim 4, characterized in that, At least one unbalance mass block (19b) is coupled to a specified rotating shaft (20a) by a rotation - direction - related coupling element (26) such that when the rotation direction changes, the unbalance mass block (19a) rotates relative to the rotating shaft (20a), in particular by 180°.

6. The stable component (11) according to claim 5, characterized in that, The at least one unbalance mass block (19b) is coupled to the specified rotating shaft (20a) by a centrifugal locking mechanism (52).

7. The stabilizing component (11) according to any one of claims 1 to 6, characterized in that, A central rotating shaft (20a) and lateral rotating shafts (20b) respectively on the left and right sides of the central rotating shaft (20a) are arranged parallel to the longitudinal direction (7).

8. The stable component (11) according to claim 7, characterized in that, The unbalance of the unbalance mass blocks (19a, 19b, 19c) assigned to the central rotating shaft (20a) is twice as large as the unbalance of the unbalance mass blocks (19d, 19e, 19f, 19g) assigned to the respective lateral rotating shafts (20b).

9. The stabilizing component (11) according to claim 7 or 8, characterized in that, The directly - driven unbalance mass blocks (19a, 19c, 19g, 19f) are coupled to the common drive (57), and the indirectly - driven unbalance mass blocks (19b, 19d, 19e) are coupled to the directly - driven unbalance mass blocks (19a, 19c, 19g, 19f) via a planetary gear drive (30).

10. The stabilizing component (11) according to claim 9, characterized in that, The cage (34) of the planetary gear drive (30) is rotatably supported and coupled to a rotary drive (39).

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

12. An orbital vehicle (1), said orbital vehicle having a frame (9) which can move on a track (2) in such a way as to be supported on an on-rail running mechanism (10), characterized in that, At least two stabilizing assemblies (11) according to any one of claims 1 to 11 are arranged such that a front stabilizing assembly (11) having a first height actuator (13) is fastened to the frame (9), and a rear stabilizing assembly (11) having a second height actuator (13) is fastened to the frame (9).

13. The rail vehicle (1) according to claim 12, characterized in that, The vibration generator (12) and the height actuator (13) are actuated by a common control device (58), and the control device (58) is arranged to actuate the respective vibration generator (12) and the respective height actuator (13) individually.

14. A method for operating a rail vehicle (1) according to claim 12 or 13, characterized in that, The rail vehicle (1) moves forward, and the front stabilizer assembly (11) operates with vertical impact forces (F, F max , F red ), and the rear stabilizer assembly (11) operates with horizontal impact forces (F, F max , F red ).

15. The method according to claim 14, wherein Vertical acceleration is recorded at the front stabilizing assembly (11) by the acceleration sensor (59) in order to derive the reaction force trend of the track panel (4) based on the vertical acceleration.

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