Patent Application
20200269887
The proposed technical solutions relate to the railway vehicles, especially to the railroad truck design that provides high dynamic characteristics of the freight railway car that meets the respective requirements of the Association of American Railroads' standards.
The dynamic characteristics of a railway car are determined, in many respects, by the parameters of its truck components, specifically, by the design of friction shock absorbers, the dimensions and performance of central swing suspension springs and the mechanical features of truck bolster side bearers.
The central swing suspension includes friction shock absorbers, each containing a pair: friction wedge-friction pad, and a set of springs made of composite two-row compression springs supporting the bolster and friction wedges.
The dynamic performance of freight railway car and its impact on the railway track are described with two key indicators: the design deflection safety factor Kres of central swing suspension set of springs and the relative friction coefficient φ (φ—“Relative friction coefficient”), which describes the efficiency of railway car body shock absorption by the friction shock absorber.
With inefficient shock absorption, resonance swaying of the railway car body may occur, i.e. the design deflection safety factor Kres is inadequate. With excessive shock absorption, which occurs when Kres value is high, the friction shock absorber rubbing parts may get locked as the railway car suspension becomes blocked with the impact on the railway track increased.
The physical sense of the design deflection safety factor Kres of central swing suspension springs is to demonstrate the spring deflection safety before any undesirable contact of the central swing suspension spring working coils in all possible regular operation modes of freight railway car. The central swing suspension spring flexibility, in this case, must be the highest provided the maximum allowable vertical movement of the springs under compression.
The required resistance to the freight railway car body vertical vibration is created by friction forces resulting from the relative movement of friction shock absorber parts rubbing between themselves in the pair “friction wedge-friction pad”. The friction forces between those parts have relative friction coefficient «qp». Coefficient «qp» must ensure effective reduction vibration amplitude of the freight railway car body with the highest possible central swing suspension spring flexibility when the conditions of allowable vertical movement of longitudinal axles of couplers of adjacent railway cars are met.
The railroad truck side bearers take forces resulting from the freight railway car body pivot pitch and transmit them to the bolster. Properly selected springing element of the side bearers ensure effective body shock absorption and improve the freight railway car dynamic characteristics.
With the purpose of improving the freight railway car dynamic characteristics, a great number of railroad truck designs have been developed.
Some of the most common trucks with an axial load of 32.5 tf are Motion Control trucks manufactured by Amsted Rail. Also, are known the following truck technical solutions with a swing suspension system made of S-335-compliant springs: U.S. Pat. No. 7,174,837 B2, Three-piece motion control truck system; Pub. No.: US 2013/0056919 A1, Damping device. The main disadvantage of these trucks is that their dynamic characteristics poorly match the respective requirements of the Association of American Railroads' standards.
The needs of an improved railway car truck, which comprises side frames with vertical columns and bolster openings, a central swing suspension, a bolster, side bearers, friction shock absorber friction wedges are met through the present invention, which, in one embodiment, describes a central swing suspension set of springs that boosts the truck's dynamic performance.
In another embodiment of the present invention, the dynamic performance of the railway car truck is improved by the described design of the friction shock absorber friction wedge of the railway car truck.
In yet another aspect of the present invention, the dynamic performance of the railway car truck is improved by the described design of the side bearings of the railway car truck.
In yet another embodiment of the present invention, the foregoing components are utilized in combinations for their complementary benefits.
The railroad truck has side frames 1, bolster 2 (
On bolster 2 end parts 2.1, guiding pockets 2.2 are implemented with sloping bearing surfaces 2.3 (
On the side frame 1 vertical columns 1.1, the friction shock absorber friction pads 4 are fastened (
Friction wedge 5 (
The front wall is implemented as two front bearing surfaces 5.1 sloping at an angle α to the base 5.3 and at an angle 2β to each other, designed to interact with mating bearing surfaces 2.3 of bolster 2 guiding pockets 2.2.
Front bearing sloping surfaces 5.1 in their bottom side part and base 5.3 of friction wedge 5 may be implemented with the respective recesses M to place in them the upper ends of outer springs 6.
The friction force arising between sloping bearing surfaces 2.3 of bolster 2 guiding pockets 2.2 and sloping front bearing surfaces 5.1 of the friction shock absorber friction wedges 5 in the proposed railroad truck is 2-3 times less than the friction force arising between flat surface 5.2 of friction wedges 5 rear wall and the mating surfaces of friction pad 4 fastened on the side frame 1 vertical columns 1.1.
Friction pads 4 are implemented with height ‘a’ in 245-260 mm range and their contact surface roughness of Rz40 to Rz80 ensuring friction coefficient fN=0.38 between friction pads 4 and flat surface 5.2 of friction wedges 5 rear wall. Height ‘b’ of each friction pad 4 lower edge position in reference to bearing surface 3.1 of side frame 1 bolster opening 3 is implemented in 160-175 mm range.
Ensuring friction wedge 5 contact on its entire flat surface 5.2 with friction pad 4 surface results in a uniform wear of friction wedge 5 and friction pad 4 contacting surfaces due to implementation of the following parameters:
The selection of above ranges ‘a’ and ‘b’ is governed by the central swing suspension spring deflection values to ensure friction wedge 5 rear wall contact on its entire flat surface 5.2 with friction pad 4. With empty freight railway car, the full-scale deflection f1 of central swing suspension set of springs equals 25 mm; with loaded car the full-scale deflection f2 of central swing suspension set of springs equals 83 mm.
Full contact of friction wedge 5 rear wall flat surface 5.2 with friction pad 4 of empty car is ensured at height ‘a’ value of more than 160 mm Full contact of friction wedge 5 rear wall flat surface 5.2 with friction pad 4 of loaded car is ensured at height ‘a’ value of less than 175 mm.
The increase of contact area of “friction wedge-friction pad” pair contacting surfaces required to enhance the efficient freight railway car body shock absorption can be reached at height ‘b’ value of more than 245 mm. The uniform wear of friction pad 4 contacting surface is ensured at height ‘b’ value of less than 260 mm, because the highest full-scale deflection of the central swing suspension set of springs under freight railway car gross weight will ensure reaching by friction wedge 5 of friction pad 4 lower edge.
The selection of friction pad 4 height ‘a’ is made based on the condition:
a≥c+f2,
where:
c height of friction wedge 5;
f2—full-scale deflection of central swing suspension set of springs in loaded freight railway car.
If a, b, f2 values get into the boundary values of their respective allowable ranges, it is guaranteed that the contact area of “friction wedge-friction pad” pair remains constant in the empty and loaded modes of operation.
The sloping front bearing surfaces 5.1 are inclined at an angle α to the base 5.3 (
In order to ensure a highly effective relative friction coefficient φ within the desired ranges (between 0.1 and 0.4 for empty railway car and between 0.07 and 0.13 for loaded railway car) the values of angles α and 2β must stay in the following ranges:
For effective interaction between friction wedge 5 and sloping bearing surfaces 2.3 of bolster 2 guiding pockets 2.2, contact area Z is implemented on sloping front bearing surfaces 5.1. The contact area Z is made with 1.0-5.0 mm deep thermal quenching, up to 350-450 HB hardness and with roughness of Rz 40 to Rz 80 to ensure friction coefficient fN=0.15 between the surface of contact area Z and sloping bearing surfaces 2.3 of bolster 2 guiding pockets 2.2. Width hz of contact area Z is 37-99 mm. In reference to base 5.3, the middle of contact area Z is at a distance ‘e’ equaling 68 mm assumed the centerline z-z of contact area Z.
Flat surface 5.2 of friction wedge 5 rear wall designed to interact with friction pad 4 contact surface is implemented with roughness of Rz 40 to Rz 80 to ensure friction coefficient fN=0.38 between flat surface 5.2 and friction pads 4.
The implementation of friction shock absorber friction wedges 5 and friction pads 4 with the above parameters ensures that the friction force difference between sloping bearing surfaces 2.3 of bolster 2 guiding pockets 2.2 and sloping front bearing surfaces 5.1 of friction wedges 5 is 2-3 times less than the friction force arising between flat surface 5.2 of friction wedges 5 rear wall and friction pads 4 to ensure improvement of the freight railway car dynamic characteristics.
On the front wall of friction wedge 5, between sloping front bearing surfaces 5.1, intermediate section 5.4 may be implemented (
Intermediate section 5.4, in its longitudinal direction, may be implemented with convex surface with radial distance R maintained in 650 to 920 mm range. Such implementation of intermediate section 5.4 provides more accurate and fluid mutual setting and interaction of friction wedge 5 and bolster 2 mating surfaces. Such fluid interfacing also improves the freight railway car dynamic characteristics.
Intermediate section 5.4, in its cross-section, may be implemented with an inner arc-like surface with radius r (
Width W of friction wedge 5 may be between 155 and 175 mm. This range of width W was chosen based on the condition of free positioning of friction wedge 5 in bolster 2 guiding pockets 2.2, as well as on the condition of optimal mounting of guiding pockets 2.2 on bolster 2 end parts 2.1. Height c of friction wedge 5 is between 135 and 160 mm.
Friction wedge 5 shown in
The dynamic characteristics of the freight railway car depend, in many respects, on the performance of the springs in the central swing suspension set of springs. The improvement of railway car dynamic characteristics can be accomplished by selecting the optimal parameters of the central swing suspension set of springs.
The central swing suspension set of springs, which support bolster 2 (
The height of each underwedge outer spring 8 is less than the height of underwedge inner spring 9 by a value of 6 to 10 mm.
To ensure that the design deflection safety factor Kres of central swing suspension set of springs is between 1.50 and 1.75, the ratio of height ‘h’ of each inner spring 7 from the set of springs for placing under bolster 2 to height ‘H’ of each underwedge inner spring 9 must be within 0.90 to 0.95 range. In this case, the height ‘h’ of inner spring 7 is less than the height of outer spring 6 within 2 to 6 mm range; the height of underwedge outer spring 8 is less than the height of underwedge inner spring 9 by a value within 6 to 10 mm range.
The two-row springs under the friction wedges are implemented with total stiffness of 353.7 kN/m; in addition, each outer spring 8 is implemented with outer diameter d8 of 123 to 125 mm, free height h8 of 282 to 286 mm, has stiffness of 251.4 kN/m and is made of a rod with diameter d8 of 20.5 to 21.5 mm; each inner spring 9 is implemented with outer diameter d9 of 78 to 80 mm, free height h9 of 290 to 294 mm, has stiffness of 102.3 kN/m and is made of a rod with diameter d9 of 13 to 14 mm.
The two-row springs under the bolster are implemented with total stiffness of 531.6 kN/m; in addition, each outer spring 6 is implemented with outer diameter d6 of 138 to 142 mm, free height h6 of 273 to 277 mm, has stiffness of 354.3 kN/m and is made of a rod with diameter d6 of 23 to 25 mm; each inner spring 7 is implemented with outer diameter d7 of 88 to 90 mm, free height h7 of 271 to 275 mm, has stiffness of 177.3 kN/m and is made of a rod with diameter d7 of 15 to 17 mm.
In the Table below, spring parameters are listed as an example of implementation of the central swing suspension compression springs with the stated parameters.
The implementation where the underwedge outer spring 8 height is less than the underwedge inner spring 9 height by an amount within 6 to 10 mm range results in:
If the height difference between underwedge springs 8 and 9 is more than 10 mm, this will lead to a decrease of their deflection safety and, accordingly, under the car gross weight load, closing of coils of underwedge springs 8 and 9 may occur.
If the height difference between underwedge springs 8 and 9 is less than 6 mm, this will lead to increase of their stiffness, rise of vertical accelerations and, as a result, to decrease of the safety factor of railroad truck against derailment.
If the ratio of minimum height spring 6 or 7 from the set of springs for placing under bolster 2 to maximum height spring 8 or 9 from the set of underwedge springs is less than 0.90, then deflection of underwedge springs 8, 9 increases with simultaneous decrease of the deflection safety factor resulting in a shorter service life of the central swing suspension springs.
If the above ratio is more than 0.95, this will result in decrease of deflection for empty car and friction wedge 5 contraction ratio, thus worsening the vibration damping of the freight railway car and decreasing the wear margin of friction wedge 5 rubbing surfaces.
If the design deflection safety factor Kres of central swing suspension springs is less than 1.5, this will lead to a decrease of deflection reserve before the closing the working coils of central swing suspension springs and, as a result, to loss of their strength. In addition, with high amplitudes of freight railway car body vibration, an increase of the vertical dynamics coefficient will cause closing of most spring working coils and their cutoff, which will result in shock impact on the railway track.
Use of springs with the design deflection safety factor Kres of central swing suspension springs less than 1.75 will lead, with the limited size of bolster opening 3.1 and due to large stiffness of the springs, to emergence of large accelerations under conditions when the freight railway car runs on uneven sections of railway track, which will result in increase of vibration frequency of the freight railway car and in worsening of its stability.
The graph analysis reveals that when the central swing suspension springs simultaneously come into operation, the graph lines sharply increase indicating that the set of springs reach the empty railway car mode with slight spring deflection. This does not provide the empty freight railway car with the vertical travel reserve, which causes large dynamic forces on the suspension and the freight railway car structures, leading to large dynamic impact on the railway track and a possibility of the empty car derailment.
Graph C, illustrating the work of proposed set of springs, reveals that sequential placement in operation of the central swing suspension springs facilitates sooner placement in operation of the friction wedge (areas 1-2) ensuring its optimal operation, and after subsequent placement in operation of the springs located under the bolster (areas 2-3), the central swing suspension set of springs reaches the working mode of operation with the last, shortest, springs placed in operation before reaching the loaded railway car mode. Such a displacement of graph C to the right from zero ensures an increase of the deflection safety and spring travel for empty car while maintaining the friction shock absorber efficient operation. The increased travel reserve for empty car guarantees the optimal dynamic properties while ensuring the necessary spring deflection safety for loaded car.
To ensure the optimal dynamic properties, the ratio of height ‘h’ of inner springs 7 placed under bolster 2 to height ‘H’ of underwedge inner springs 9 is within 0.90 to 0.95 range. This ratio ensures that even before springs 6, 7 located under bolster 2 start working, underwedge springs 8, 9 will be placed in operation sooner, which guarantees constant pressing of friction wedge 5 to friction pad 4.
Areas 1-2-3 (
The claimed h/H ratio within 0.90 to 0.95 range and selection of springs 6, 7 ensure that the value of the design deflection safety factor Kres of central swing suspension springs is in 1.50 and 1.75 range. In turn, this ensures increased deflection safety of the central swing suspension springs under static and dynamic operating loads including dynamic loading, which in the physical sense means a lesser contraction ratio of springs. This leads to a lesser deformation of spring coils, drop of stress and, as a result, to improvement of the coil reliability, increase of their service life, decrease of dynamic loads, improvement of empty car's vibration damping and to its high dynamic properties.
In the first area of the graph, from 0 to point 1 (line Cavg), only underwedge inner springs 9 work; the full-scale deflection f1 of central swing suspension equals 8 mm.
In the second area of the graph, between points 1 and 2, friction wedges 5 bias against both inner 9, and outer 8 underwedge springs. In area 0-2, the values of full-scale deflections f 9-2 and f 8-2 of inner 9 and outer 8 underwedge springs are: f 9-2=17 mm and f 8-2=9 mm.
In the third area, from point 2 to point 3, between points 2 and 3, outer springs 6 located under bolster 2 begin operating. In area 0-3, the values of full-scale deflections f 9-3 and f 8-3 of inner 9 and outer 8 underwedge springs are: f 9-3=19 mm, f 8-3=11 mm; the value of full-scale deflection f 6-3 of outer spring 6 located under bolster 2 is: f 6−3=2 mm.
In the fourth area, inner springs 7 located under bolster 2 begin operating, in addition, all springs from the central swing suspension set of springs also begin operating.
The ratio of maximum height of springs 6, 7 located under bolster to their minimum height is within 0.90 to 0.95. This helps to maintain strength and reliability of the springs under static and dynamic operating loads.
A major contribution to obtaining optimal dynamic performance of the freight railway car is provided by the railroad truck side bearers. Properly selected design and matching springs of side bearers exert a significant impact on the improvement of this performance.
Side bearer 10 (
Springing element 11 may be implemented as a single compression spring 11.1 or made of multiple compression springs 11.2, 11.3, 11.4, 11.5, 11.6 (
For use under a car with the least tare weight of 18 t, springing element 11 may be implemented as a single spring 11.1 (
With car tare weight of at least 22 t for truck side bearers springing element 11 may be used with total stiffness of 824.64 kN/m, made of two compression springs 11.2 and 11.3 located one inside the other (
For car tare weight of at least 25 t for truck side bearers springing element 11 may be used with total stiffness of 1073 kN/m, made of three compression springs 11.4, 11.5, 11.6 (
The ranges described in the foregoing paragraphs are provided in the context of certain embodiments. The described ranges and may be different for alternative embodiments or if non-standard materials or railcar designs are used.
On the inner surface of the stop sleeve 12 of the railroad truck side bearers, intermediate support plate 13 may be placed for use in supporting springing elements 11 and protecting stop sleeve 12 inner surface from wear.
Consequently, the proposed technical solutions for the group of inventions within the railroad truck and its components significantly improve the freight railway car dynamic properties to ensure compliance with the requirements of the Association of American Railroads' standards.
While at least one exemplary embodiment of the present invention(s) is disclosed herein, it should be understood that modifications, substitutions and alternatives may be apparent to one of ordinary skill in the art and can be made without departing from the scope of this disclosure. This disclosure is intended to cover any adaptations or variations of the exemplary embodiment(s). In addition, in this disclosure, the terms “comprise” or “comprising” do not exclude other elements or steps, the terms “a” or “one” do not exclude a plural number, and the term “or” means either or both. Furthermore, characteristics or steps which have been described may also be used in combination with other characteristics or steps and in any order unless the disclosure or context suggests otherwise.
