EA045784B1 - FITTINGS FOR RAILWAY CAR VIBRATION DAMPER WEDGES - Google Patents

Description

Technical field

The invention relates to the field of vibration damper fittings for bolsters of railway car bogies.

State of the art

The problem of the running performance of railway freight cars has existed for many years. Previous attempts to solve this problem are presented in WO 2005005219 (Forbes and Hematian). As explained in this application, driving performance typically involves an often subtle interplay of dynamic responses across multiple degrees of freedom at multiple interacting interfaces.

The relationship between the self-steering reaction at the interface between the bearing adapters and the side frame axlebox supports and the coinciding lateral rocking movement at that interface, as well as the reaction at the interface between the bogie bolster and the side frames, contributes to the overall driving performance, while damping of all these movements is provided by friction dampers hesitation. Most often, these vibration dampers take the form of triangular wedges installed to operate between the bolster and the side frame columns. The type of friction coupling affects the characteristics. Those. For many years, friction was provided by the sliding interaction of steel (or cast iron) with steel. More recently, vibration damper wedges have been used with non-metallic running surfaces or pads that rest on the side frame column wear plates.

It was assumed that the force perpendicular to the friction surface of the vibration damper wedge can be reasonably approximated by a point load at the center of the friction surface or a distributed load that either has the same magnitude over the entire friction surface, or as a load distribution that actually has its center in the middle of the friction surface .

This assumption may have been adequate for previous approximations. However, later observations showed that the force on the friction surface is distributed completely differently than in the previous assumption. In other words, the pad tends to wear predominantly at the top edge as the bolster moves up relative to the side frame, and at the bottom edge as the bolster moves down. In addition, although it was assumed that the force distribution would be approximately the same from vibration damper wedge to vibration damper wedge, actual observation shows that the force distribution on the front surface of the friction vibration damper wedge is unexpectedly sensitive to manufacturing tolerances and variations on the rear surface of the vibration damper wedge.

Where the edge of a non-metallic friction lining is worn, the underlying metal parts may tend to scuff and scratch each other. This will most likely affect the performance of the cart, which is not desirable.

Given the inconvenience of replacing friction damper wedges undesirably frequently or at unexpected intervals, the uncertainty of changing vibration damper wedge parts, the variability of driving characteristics, and the uncertainty of wear pad life due to such variability pose a problem.

The essence of the invention

The invention relates to the design of a friction vibration damper wedge, which has a non-metallic friction surface that interacts with the wear plate of a side frame column during sliding movement. The friction vibration damper wedge has an inclined curved surface having an operating point in a certain place, which interacts with the corresponding pocket of the bolster. The location of the operating point is such that when it interacts with the bolster pocket, a more even distribution of the load on the friction surface is ensured.

In one aspect of the invention, a vibration damper wedge is provided for a railroad car bogie. The vibration damper wedge is formed to interact with the bolster pocket, which has a primary vibration damper alpha angle and a secondary vibration damper beta angle. The vibration damper wedge has a friction element, which, when in use, movably interacts with the wear surface of the side frame column of the railway car bogie. The friction element has a non-metallic wear surface. There is a spring socket that, when in use, interacts with the rail car bogie spring. There is an inclined surface of the vibration damper wedge having a primary angle alpha and a secondary angle beta. The inclined surface of the vibration damper wedge has a curvature. The curvature has a working point WP. The spring seat has an axial center line. The vibration damper wedge has a reference plane that is perpendicular to the non-metallic wear surface and that contains an axial center line. The axial center line intersects the inclined surface of the vibration damper wedge at the intersection point. The operating point is located in the central region of the inclined surface of the vibration damper wedge, adjacent to the specified reference plane, down the slope from the intersection point.

One indication of this aspect is that the operating point is offset from the non-metallic wear surface of the friction element by a distance ranging from _1/8 to ^5/8 inch (3 to 16 mm) further than the centerline. In another feature, the operating point is offset from the non-metallic wear surface of the friction element by a distance ranging from 1/4 to ^3/4 inch (6 to 19 mm) further than the centerline. According to another feature, the non-metallic wear surface is offset from the axial center line by a first distance x1; the operating point is shifted from the center line by a second distance x ₂ ; and the ratio x1:x ₂ is in one of the ranges: (a) from 21:2 to 21:8; and (b) from 10:3 to 40:3. According to another feature, this ratio is about 5:1. According to another feature, the non-metallic wear surface has an overall height y1, and the operating point is in the range of ^3/8 y1 to ^5/8 yi along the height of the non-metallic wear surface.

According to another feature, the curvature is a compound curvature. According to an additional feature, the curvature of the inclined surface is spherical. According to another feature, the curvature has a radius of curvature of less than 35 inches (89 cm). According to another feature, the specified radius is in the range of 15 to 30 inches (38 to 76 cm). According to another feature, the alpha angle is in the range from 30 to 50°. According to another feature, the beta angle is in the range from 5 to 20°. According to another feature, the inclined surface has a portion of the working surface, or contact patch, that has a radius of less than 2 inches (5 cm).

According to another feature, the vibration damper wedge has first and second end walls; the inclined surface is located between the first and second end walls, and the base plane is located midway between the first and second end walls. According to another feature, the inclined surface extends towards the first and second end walls. According to another feature, the vibration damper wedge includes an internal partition extending between an inclined surface of the vibration damper wedge and the friction element, and the operating point is aligned with the internal partition. According to another feature, the operating point is located along the base plane. According to another feature, the vibration damper wedge is at least partially hollow. According to another feature, the vibration damper wedge includes a mounting bracket.

According to another feature, the spring seat includes a downwardly projecting protrusion, the size of which allows it to be coaxially located within the spring of a railway car bogie. According to another feature, the vibration damper wedge is used in assembly with the spring group of a railway car bogie. The spring group has a vibration damper spring installed to interact with the spring socket for the vibration damper. There is at least a first main spring. The first vibration damper spring has a greater free height than the first main spring. According to another feature, the spring seat is formed in the lower wall of the vibration damper wedge; The vibration damper wedge has corners rounded at the rear, the radius of which exceeds _1/4 of the width of the vibration damper wedge, and the inclined surface of the vibration damper wedge ends at the outer edge of at least one of the rounded corners. According to another feature, the set for one end of the bogie bolster includes four vibration damper wedges, wherein the four vibration damper wedges include two vibration damper wedges with a left-hand beta angle, and two vibration damper wedges have a right-hand beta angle. According to another feature, the vibration damper wedge is designed to be assembled with the bolster pocket.

According to another feature, there is a railway car bogie that has an assembly of a vibration damper wedge and at least one self-steering device. According to another feature, the self-steering device includes a rocker arm located between the axle box support of the side frame and the bearing of the wheelset axle. According to another feature, a railway car bogie has a side frame mounted to swing laterally from side to side, wherein the side frame has a lateral rocking stiffness coefficient kpendulum; The railway car bogie has a bolster mounted on spring groups, and the spring groups have a coefficient of lateral stiffness k _springs h _e r, with k _pendulum being less than kspringshear.

According to another aspect of the invention, a vibration damper wedge is provided for a railroad car bogie. The vibration damper wedge is configured to interact with the bolster pocket having a primary vibration damper alpha angle and a secondary vibration damper beta angle. The vibration damper wedge has a friction element, which, when in use, movably interacts with the wear surface of the side frame column of the railway car bogie. The friction element has a non-metallic wear surface. There is a spring socket which, when in use, interacts with the railcar bogie spring. There is an inclined surface of the vibration damper wedge, wherein the inclined surface of the vibration damper wedge has a primary angle alpha, a secondary angle beta and curvature. The curvature has an operating point at which the inclined surface interacts with the bolster pocket at rest. The spring seat has an axial center line. The vibration damper wedge has a reference plane that is perpendicular to the non-metallic wear surface and that contains an axial center line. The axial center line intersects the inclined surface of the vibration damper wedge at the point

- 2 045784 intersections. The vibration damper wedge has a first end surface and a second end surface. The first and second end surfaces are located at a distance from each other and are opposite to each other. The reference plane is located midway between the first and second end surfaces. The operating point is located in a central region defined adjacent to the reference plane down the slope from the intersection point. According to another feature, the radius of curvature in the basal plane is less than 30 inches (76 cm). According to another feature, the vibration damper wedge has first and second end surfaces. The reference plane is defined at the midpoint between the first and second end surfaces. The vibration damper wedge is asymmetrical.

The features listed in relation to the preceding aspect of the invention are also applicable in relation to the previous aspect.

In yet another aspect, a vibration damper wedge for a railway car bogie is provided. The vibration damper wedge has a size that allows it to fit in the corresponding pocket of the bolster of a railway car bogie. The vibration damper wedge has a housing. The housing has a friction surface designed to interact with the column of the side frame of the railway car bogie. The friction surface has a non-metallic wear surface. The housing has a spring seat sized to mate with the upper end of the railway car bogie spring. The housing has an inclined surface that is formed to interact with a corresponding inclined surface of the bolster pocket. The inclined surface has a primary vibration damper wedge angle and a secondary vibration damper wedge angle. An inclined surface has a curvature. The housing has a first side surface and an opposite second side surface, as well as a central plane located between the first side surface and the second side surface. The first side surface is larger than the second side surface. The central plane is perpendicular to the friction surface. The central plane intersects the inclined surface. The inclined surface has a working point located on the central plane.

The features listed in relation to the preceding aspects of the invention also apply to the previous aspect.

In yet another aspect, a vibration damper wedge is provided compatible with a bolster pocket of a railroad car bogie. The vibration damper wedge has an inclined surface and has a primary alpha angle and a secondary beta angle. The inclined surface has an outwardly convex composite surface configured to interact with the bolster pocket at the operating point.

The secondary angle beta determines the direction of lateral displacement of the vibration damper wedge. The vibration damper wedge has a friction surface which, when in use, interacts with the wear surface of the side frame column of the railway car bogie. The friction surface has a non-metallic wear surface. The friction surface has a perpendicular plane passing through it. The perpendicular plane in use also extends in the direction of movement of the vibration damper wedge up and down. The perpendicular plane is also centered at the middle of the width of the non-metallic wear surface. The operating point lies on a perpendicular plane. The features listed in relation to the preceding aspects of the invention also apply to the previous aspect.

According to another aspect of the invention, a vibration damper wedge is provided for a railroad car bogie, wherein the vibration damper wedge is formed to cooperate with a corresponding bolster pocket having a primary damper angle alpha and a secondary damper angle beta. The vibration damper wedge has a friction element, which, when in use, movably interacts with the wear surface of the side frame column of the railway car bogie. The friction element has a non-metallic wear surface. There is a spring socket which, when in use, interacts with the railcar bogie spring. The vibration damper wedge has an inclined surface of the vibration damper wedge. The inclined surface of the vibration damper wedge has a primary vibration damper wedge angle and a secondary vibration damper wedge angle. The inclined surface of the vibration damper wedge has a curvature. Curvature has an operating point. The spring seat has an axial center line. The vibration damper wedge has a reference plane that is perpendicular to the non-metallic wear surface and that contains an axial center line. The axial center line intersects the inclined surface of the vibration damper wedge at the intersection point. The intersection point is the center of the work surface contact patch, which has a radius of less than 1.5 inches (3.75 cm). The operating point is located on the contact patch of the working surface.

The features listed in relation to the preceding aspects of the invention also apply to the previous aspect.

According to another aspect, a vibration damper wedge for a railway car bogie is provided. It has a friction surface which, when in use, interacts with the column wear plate of the side frame of the railway car bogie. It has a spring seat that

- 3 045784 interacts with a spring during use, the seat of the spring having an axial direction. It has an inclined surface which, when in use, interacts with the corresponding surface of the bolster pocket of the railway car bogie. There is a base plane perpendicular to the friction surface. The reference plane is parallel to the axial direction of the spring seat. The inclined surface has a spherical arc and a radius of curvature. The inclined surface has a primary vibration damper wedge angle and a transverse secondary vibration damper wedge angle. The inclined surface has an operating point which, in equilibrium, interacts with the bolster pocket at the rolling contact point. The radius of curvature of an inclined surface has its origin lying on one side of the base plane. The radius passes through the specified origin and operating point. The radius deviates from the reference plane at a bevel angle, the bevel angle being a secondary angle of the vibration damper wedge when said radius is viewed in a plane oriented at the primary angle of the vibration damper wedge.

According to the feature of this aspect, the axial direction of the spring seat lies in the reference plane and the operating point also lies in the reference plane. According to another characteristic, the radius of curvature is in the range of 15 to 30 inches (38 to 76 cm). Another indication is that the radius of curvature is in the range of 20 inches ± 1 inch (51 ± 2.5 cm). In another feature, the operating point is offset from the non-metallic wear surface of the friction element by a distance ranging from 1/8 to 5/8 inch (3 to 16 mm) further than the centerline. In another feature, the operating point is offset from the non-metallic wear surface of the friction element by a distance ranging from _1/4 inch to ^3/4 _inch (6 to 19 mm) further than the center line. In yet another feature, the non-metallic surface is offset from the axial center line by a first distance x1; the operating point is shifted from the center line by a second distance x ₂ ; and the ratio x1:x ₂ is in one of the ranges (a) from 21:2 to 21:8; and (b) from 10:3 to 40:3. According to another criterion, the ratio is 5:1. According to yet another feature, the center line intersects said inclined surface at a reference point. The base point lies in the base plane. The operating point is located less than 1 inch (2.5 cm) from the reference point. The features of the previous aspects of the invention also apply to the above aspect.

These and other aspects and features of the invention may be understood by reference to the following description and illustrations.

Brief description of drawings

The principles of the invention may be better understood by reference to the accompanying drawings provided to illustrate an exemplary embodiment or embodiments incorporating the principles and aspects of the present invention, in which: FIG. 1a is an isometric view of an example embodiment of a railroad car bogie in accordance with one aspect of the present invention;

in fig. 1b is a top view of the railway car bogie of FIG. 1a;

in fig. 1c is a side view of the railway car bogie of FIG. 1a;

in fig. 1d is an exploded view of a portion of the cart of FIG. 1a;

in fig. 2a is an isometric rear and top view of the vibration damper wedge used in the cart of FIG. 1a;

in fig. 2b shows a bottom and rear view of the vibration damper wedge of FIG. 2a;

in fig. 2c shows a view of the vibration damper wedge of FIG. 2a front and top;

in fig. 2d shows an exploded view of the vibration damper wedge of FIG. 2c with wear pad before installation;

in fig. 3a shows a front view of the vibration damper wedge of FIG. 2a;

in fig. 3b is a rear view of the vibration damper wedge of FIG. 3a;

in fig. 3c is a large side view of the vibration damper wedge of FIG. 3a;

in fig. 3d is a small side view of the vibration damper wedge of FIG. 3a;

in fig. 3e shows a top view of the vibration damper wedge of FIG. 3a;

in fig. 3f is a sectional view along the vertical center plane of the spring seat, designated by section '3f-3f' in FIG. 3a;

in fig. 3g is a cross-sectional view in a horizontal plane, designated by section '3g-3g' in FIG. 3a;

in fig. 3h shows a section plane taken along a spherical radius through the operating point, as indicated by section '3h-3h' in FIG. 3c;

in fig. 4a is a partial sectional side view of the end of the side frame of the railway car bogie of FIG. 1a;

in fig. 4b is a sectional view through the side frame taken at '4b-4b' in FIG. 4a;

in fig. 4c is a sectional view of FIG. 4b, deflected laterally during rocking movement;

in fig. 4d is a longitudinal sectional view through the axle box housing and side frame bearing adapter assembly of FIG. 4a along section '4d-4d' in FIG. 4b;

- 4 045784 in Fig. 4e shows the axle box housing and bearing adapter assembly of FIG. 4c in a position tilted in the longitudinal direction;

in fig. 5a is an exploded perspective view of the side frame of FIG. 4a with bearing, bearing adapter and bearing adapter shock absorbers;

in fig. 5b shows the underside of the bearing adapter of FIG. 5a;

in fig. 6a is an isometric view of one of the shock absorbers of FIG. 5a;

in fig. 6b shows the opposite isometric view of the shock absorber of FIG. 6a;

in fig. 6c is a front view of the shock absorber of FIG. 6a;

in fig. 6d shows a bottom view of the shock absorber of FIG. 6a; and in fig. 6e is a sectional view of the shock absorber of FIG. 6c, taken along the section '6e-6e' in FIG. 6c.

Carrying out the invention

The following description and the embodiments described therein are provided to illustrate an example or examples of specific embodiments of the principles, aspects or features of the present invention. These examples are provided for the purpose of explaining, and not limiting, these principles and the invention. In the present description, like parts are designated by the same corresponding reference numerals in the description of the invention and in the drawings. Drawings may be to scale unless otherwise noted.

It is believed that the terminology used in this description corresponds to the generally accepted and customary meanings of these terms as understood by one skilled in the art of railroad transportation in North America. In this regard, applicant incorporates by reference the Rules and Standards of the Association of American Railroads (AAR), a private organization that sets rules for the exchange of railroad rolling stock in North America.

In addition, this description often gives many synonyms for a single object. There is no in haec verba requirement in patent law. The purpose of listing a set of synonyms is to indicate that any synonym can be used for a given part, regardless of whether that synonym is used in the description as presented, provided that it matches the meaning of the concept, function, or object passed in correct reading of the description, or that it is reliably shown in the illustrative figures, or both.

With respect to the general terminology for orientation and direction with respect to the railroad car bogies described herein, the longitudinal direction is defined to be the rolling direction of the railroad car or section of railroad cars when positioned on a tangential (i.e., straight) track. In a Cartesian coordinate system, it can be defined as the x-axis or x-direction. In the case of a railway car having a center beam, whether it is a cantilever center beam or a through center beam, the longitudinal direction is parallel to the center beam and parallel to the top chords and side beams. Unless otherwise specified, the terms vertical or up and down use the TOR (top of rail) as the basis. In a Cartesian coordinate system, it can be defined as the z-axis or z-direction. In the context of a railroad car bogie as a whole, the term lateral or laterally outward or transverse or transversely outward refers to a distance or orientation relative to the longitudinal centerline of the railroad car bogie or relative to the centerline of a center plate at the center of the cart. In a Cartesian coordinate system, this may be called the y-axis or y-direction. Given that a railway car bogie may typically have longitudinal and transverse axes of symmetry, a description of one half of the bogie can usually be generally intended to describe the other half, taking into account the differences between the right and left halves. A rotation about its transverse axis is an angular movement about a horizontal axis perpendicular to the longitudinal direction (ie, a rotation about an axis in the y direction). A yaw rotation is an angular movement about a vertical axis or z-axis. Rocking is an angular movement about a longitudinal axis or x-axis. In this description, the abbreviation kpsi, when used, means thousands of pounds per square inch. To the extent that this specification or the accompanying illustrations may refer to Association of American Railroads (AAR) standards, such as AAR placard sizes, such references should be understood to refer to the earliest priority date to which this document applies. Unless otherwise noted, it should be understood that the railroad cars described herein are of welded steel construction.

This description applies to railroad car bogies and bogie components. Several bogie sizes according to Association of American Railroad standards are listed on page 711 in the Car & Locomotive Cyclopedia, 1997. As stated, for a single railroad car having two bogies, the bogie category of 40 tons corresponds to the maximum gross weight of the car per rail. (gross car weight, GWR) 142,000 pounds (64,410.12 kg). Similarly, 50 tons equals 177,000 pounds, 70 tons equals 220,000 pounds, 100 tons equals 263,000 pounds, and 125 tons equals 315,000 pounds. .6 kg). In each case, the bogie load limit is half the maximum gross weight of the car per rail. The other two types of bogie are the 110 ton bogie for railroad cars, which has a GWR of 286,000 lb (129,727.42 kg), and the low profile 70 ton Special bogie, sometimes used for auto transporters. When applied to bogies, a wheel set includes an axle and a pair of steel wheels mounted on opposite ends of the axle.

This document applies to friction dampers or damper wedges for railway bogies and systems with multiple friction dampers. There are several types of vibration damper arrangements, some of which are shown on p. 715-716 of the 1997 Carriage and Locomotive Encyclopedia, which pages are incorporated herein by reference. Arrangements with two vibration dampers are shown and described in US Patent Application Publication No. 2003/0041772 A1, filed March 6, 2003, entitled Railroad Freight Car with Damped Suspension, also incorporated herein by reference. Each of the specified vibration damper arrangements shown on p. 715-716 of the 1997 Encyclopedia of Carriages and Locomotives, may be modified to use a quadrangular arrangement with two vibration dampers, i.e. internal and external vibration dampers in accordance with the principles of aspects of the present invention.

This document describes vibration damper wedges. In general terminology, vibration damper wedges are typically mounted within a corner bolster pocket formed at the end of the bogie bolster. Each vibration damper wedge may have a generally triangular cross-section, with one side of the triangle representing or containing a load-bearing surface; the second side, which may be called the bottom or base, forms the seat of the spring; and the third side is the inclined side or hypotenuse between said other two sides. The first side may generally have a substantially flat bearing surface for vertical sliding engagement with an opposing bearing surface of one of the side frame columns. The second surface may not be a surface per se, but may instead be in the form of a socket for receiving the upper end of one of the spring assembly springs. Although the third surface or hypotenuse may appear generally flat, it may have a slight convexity with a radius of curvature of approximately 60 inches (152.4 cm). The bulge may extend along the slope and may also extend across the slope. The side surfaces of the vibration damper wedges may be generally flat and may be coated, surface treated, skinned or have a low friction pad to provide smooth sliding interaction with the sides of the bolster pocket.

While the railroad car is in motion, the side frame may pivot or rotate through a small range of angular deflection around the end of the bogie bolster to provide equalization of the wheel loads. A small bulge on the sloping surface of the damper may tend to accommodate this rotational movement, allowing the damper to oscillate within a small range relative to the generally sloping surface of the bolster pocket while the flat bearing surface remains in flat contact with the side frame column wear plate. Although an inclined surface may have a slight convexity, for the purposes of this description it will be described as an inclined surface or as a hypotenuse and will be considered a substantially flat surface in general.

In the terminology of this document, vibration damper wedges have a primary angle α, which is the internal angle between (a) the inclined surface of the vibration damper pocket in the bogie bolster and (b) the wall surface of the side frame wedge opening, as viewed from the end of the bolster to the center of the bogie . In some embodiments, the secondary angle β is defined in the plane of the angle α, namely in a plane perpendicular to the vertical longitudinal plane of the (non-deviated) side frame deviated from the vertical at the primary angle. In other words, this plane is parallel to the (undeviated) long axis of the bogie bolster and runs along the rear side (hypotenuse) of the vibration damper. The secondary angle β is defined as the angle of lateral inclination visible when looking at the vibration damper wedge parallel to the plane of the angle α. During suspension operation, in response to path irregularities, forces affecting the secondary angle β can push the damper wedge either inward or outward according to the selected angle. The vibration damper wedges are driven or stressed by the corner springs or side springs of the spring groups. As used herein, spring groups include corner springs (or cushion springs, wedge springs, or side springs) and main springs. The main springs are located under the ends of the bolster. Accordingly, the main spring group is a subset of the spring group that includes springs mounted under the bolster, as opposed to springs mounted under the damper wedges.

- 6 045784

General description of the characteristics of the trolley.

This description is made in the context of railway car bogies. To establish this context, FIG. 1a shows an example of a cart 20. The cart 20 in FIG. 1a-1d are intended for general use and may have different swingarm lengths, spring rates, spring arrangements, wheelbases, window widths and heights, and so on. Those. the cart 20 may have a wheelbase ranging from 60 to 75 inches (150 to 187.5 cm). As discussed below, it has a spring group having a vertical spring rate and a quadrangular damper group having primary and secondary damper wedge angles. The cart 20 may have a 3x3 spring group arrangement, a 5x3 spring group arrangement, a 3:2:3 spring group arrangement, a 2x4 spring group arrangement, or any other possible arrangement. Although any cart may be suitable for a variety of general purpose applications, cart 20 may be optimized for the transport of relatively low-density and high-value cargo, such as automobiles or consumer goods, for example, for the transport of denser semi-finished manufactured goods. , which can be transported in freight railway cars for transporting paper rolls; or transportation of dense liquid or aggregated materials such as coal, crushed stone, grain, flour, potash, liquid petroleum products and so on. Thus, cart 20 is intended to represent a wide variety of cart types. The carriage 20 is symmetrical about its longitudinal center plane (x-z) and its transverse or lateral center plane (y-z). When referring to a side frame, it is understood that the cart includes first and second side frames, first and second spring groups, and the like.

The bogie 20 has a bogie bolster 24 and side frames 26. Each side frame 26 includes a generally rectangular side frame window 28 housing one of the ends 30 of the bolster 24. The upper boundary of the window 28 is defined by a side frame arc or compression member, identified as a top chord member 32, and the bottom of the window 28 is defined by a tension member identified as the bottom chord 34. The front and rear vertical sides of the window 28 are defined by the side frame columns 36. The ends of the tension element diverge upward until they meet the compression element. At each of the upwardly diverging ends of the side frame 26 are axlebox fittings or axlebox sockets 38 of the side frame. Each axle box socket 38 houses an upper fitting. The top fitting may be a flat plate socket or may be curved. The fitting may be called a rocker or socket, depending on the context, as described and discussed below. This top fitting, whatever it may be, is generally designated as reference numeral 40. The fitting 40, which may be referred to as a socket, engages or contacts a mating fitting 42 of the top surface of the bearing adapter 44. The fitting 42 may be a rocker arm that cooperates with the socket of the fitting 40, or the roles of the socket and the rocker arm may be reversed. The bearing adapter 44 interacts, i.e. sits on a bearing 46 mounted on one end of one of the bogie axles 48 adjacent to one of the wheels 50. The wheel pair includes an axle 48 and wheels 50 mounted on both ends of the axle. Bearing 46 may be, and as shown in the drawing, is a sealed roller bearing. Sealed roller bearings in standard sizes are widely used in North American practice. A fitting 40 is located in each of the front and rear axle box fittings 38, the fittings 40 being longitudinally aligned such that the side frame 26 can swing laterally relative to the rolling direction of the bogie.

The relationship of mating fittings 40 and 42 is described below with reference to FIGS. 4a-4e. The relationship of these fittings determines part of the overall relationship between the end of one of the axles 48 of one of the wheel sets and the axle box of the side frame. Those. When determining the overall response, the degrees of freedom of installation of the axle end in the side frame axle box support include a dynamic interface throughout the entire assembly of parts, such as what may be called a wheelset-side frame interface, which may include a bearing, a bearing adapter, an elastomeric a pad, if used, a rocker arm, if used, and an axle box socket mounted in the ceiling of the side frame axle mount, whether a cast, machined, or prefabricated socket is installed as a separate part. There are several different options for implementing this coupling unit between the wheelset and the side frame. To the extent that the bearing 46 has one degree of freedom, namely rotation about the axis of the wheel shaft, the analysis of the specified assembly can be focused on the bearing interface with the axle box socket or on the bearing adapter interface with the axle box socket. For the purposes of this description, elements 40 and 42 are primarily intended to represent the combination of features of the bearing adapter and axle box assembly defining the interface between the side frame axle box ceiling and the bearing adapter, and six degrees of freedom of movement at this interface, namely vertical, longitudinal and transverse translation (i.e., movement in the z, x, and y directions) and rotation about its transverse axis, rocking, and rotation about its vertical axis (i.e., rotational movement about the y, x, and z axes, respectively) in response to dynamic inputs signals.

- 7 045784

The lower chord or tension member of the side frame 26 may have a basket plate or a lower spring seat 52 rigidly attached to the frame. Although the carriage 20 may be free of unspring side cross braces, either in the form of a crossbar or side bars, in the event that the carriage 20 is taken to represent a rotary carriage with a crossbar or other cross-bar mounting, the lower rocker platform of the spring seat 52 may be mounted on the rocker arm. to provide lateral swing relative to the side frame 26. The spring seat 52 may have retainers to engage the springs of the spring set or spring group 54, sometimes referred to as the main spring group 56, whether internal lugs or a peripheral flange, to prevent the lower ends of the springs from coming out. The spring group 56 is captured between the distal end 30 of the bolster 24 and the spring seat 52, being compressed by the weight of the railroad car body and cargo that rest on the bolster 24 on top.

The bolster 24 has dual inner and outer pockets 60, 62 on each bolster surface at the outer end (i.e., a total of 8 pockets per bolster, 4 at each end). The bolster 24 is symmetrical relative to the central longitudinal vertical plane of the bolster (i.e., transversely relative to the bogie as a whole) and symmetrical relative to the vertical mid-span section of the bolster (i.e., the longitudinal plane of symmetry of the bogie as a whole, coinciding with the longitudinal centerline of the railway car ). Each surface of each end 30 of the bolster 24 has a pair of spaced apart bolster pockets 60, 62 for receiving vibration damper wedges 64, 66, 68, 70, respectively. The pocket 60 is generally positioned toward the center of the side frame 26 of the truck 20 laterally relative to the pocket 62. Wear plate inserts, such as a specially hardened machined material, may be installed in the pockets 60, 62 along the corner surfaces of the damper wedge. hesitation.

As can be seen in the drawing and as discussed in more detail below, the vibration damper wedges 64, 66, 68, 70 have a primary angle α between the vertical and angular rear apex of the larger surface. For embodiments of the invention described herein, the primary angle α may typically be in the range of 30 to 50°, possibly from about 40 to 45°. The same angle α corresponds to the front surface of the bolster pocket, be it pocket 60 or 62. The secondary angle β specifies the internal (or external) slope of the inclined surface of the vibration damper wedge. The true rake angle can be seen by looking along the plane of an inclined surface, measuring the angle between the inclined surface and the flat outer surface. The rake angle is the complement of the angle measured in this way. The rake angle may be greater than 5°, may range from 5 to 20°, and preferably from about 10 to 15°. A moderate rake angle may be desirable.

When the bogie suspension works in response to track irregularities, the vibration damper wedges can operate in their pockets. The rake angles form a force component that tends to move the inner surface of the outer wedge 68 (or 70) inward relative to the opposite inner surface of the bolster pocket 62. Likewise, the outer surface of the wedge 64 (or 66) may tend to move toward the outer flat surface of the inner bolster pocket 60. These inner and outer surfaces of the bolster pockets may be lined with a low friction surface or may be left as a metal surface as shown in the drawing. Left- and right-hand offsets of the wedges may tend to keep them closer together and keep them against the flat faces and may help prevent the dampers from twisting in their respective pockets.

The bolster 24 includes a middle platform 98 between the pockets 60, 62, against which a spring 96 operates. The middle platform 98 is such that it can accommodate a spring group three (or more) spring coils wide. However, regardless of the width of two, three, or more spring coils, and regardless of whether the center pad is used or not, the bolster pockets can have both primary and secondary corners, as shown in the drawing, with or without wear inserts. In cases where a center pad, such as pad 98, separates two vibration damper pockets, the wear plates of the opposite side frame column do not need to be solid. Those. two wear plate areas may be provided, one opposite each of the inner and outer vibration dampers, providing flat surfaces on which said vibration dampers can rest. The normal vectors to these areas may be parallel, the surfaces may be coplanar and perpendicular to the long axis of the side frame, and may provide a clear, continuous surface for the vibration damper friction surfaces.

As indicated above, in the pockets (60, 62) of the bolster there are located the front and rear pairs of the first and second, in the lateral direction of the inner and in the lateral direction of the outer wedges 64, 66 and 68, 70 of the friction vibration damper, respectively. Each pocket 60, 62 bolster

- 8 045784 has an inclined surface or vibration damper seat 72, which is mated to a similarly inclined surface 74 of the hypotenuse of vibration damper wedges 64, 66, 68 and 70. Each of the wedges 64, 66 is placed over the first inner corner spring 76, 78, and the wedges 68, 70 are placed over the second outer corner spring 80, 82. The corner surfaces 74 of the wedges 64, 66 and 68, 70 (described in more detail below) are adjacent to the corner surfaces of the corresponding pockets 72. The middle end spring 96 rests on the underside of a platform 98 located in the space between the bolster pockets 60 and 62. The upper ends of the center row springs 100 are located below the main center end portion 102 of the bolster 24. In this quadrangular arrangement, each vibration damper is individually spring-loaded by one or another spring in the spring group.

The static compression of the springs under the weight of the car body and load tends to act as a spring load to bias the vibration damper so that it acts along the slope of the bolster pocket, pressing the friction surface against the side frame. Friction damping is provided when the vertical sliding surfaces 90 of the friction damper wedges 64, 66 and 68, 70 move up and down friction wear plates 92 mounted on the inward facing surfaces of the side frame columns 36. Thus, the kinetic energy of motion is to some extent converted into heat through friction. This friction can dampen the movement of the bolster relative to the side frames. When lateral disturbance from the rails is transmitted to the wheels 50, the rigid axles 48 may tend to cause both side frames 26 to yaw in the same direction. The reaction of the side frames 26 is to swing like a pendulum on the upper rocker arms. The weight of the pendulum and the reaction force resulting from the twisting of the springs can then push the side frames back to their original position. The tendency for harmonic vibrations due to path irregularities can be reduced by friction of the vibration dampers against the wear plates 92.

Compared to a bolster with single dampers, which may be mounted on the centerline of the side frame, the use of dual dampers, such as spaced pairs of dampers 64, 68, can result in a larger moment arm to resist parallelogram deformation. 20 carts in total. The use of dual vibration dampers results in a greater restoring perpendicular force, which allows the trolley to bend, i.e. deflect in response to irregularities, but then elastically returns the cart to a rectangular orientation than in the case of a single vibration damper with restoring displacement, namely a rectangular force that increases with increasing deflection. In other words, in a parallelogram or diamond shape, the differential compression of one diagonal spring pair (e.g., inner spring 76 and outer spring 82 may be more compressed) compared to the other diagonal spring pair (e.g., inner spring 78 and outer spring 80 may be compressed less pronounced than springs 76 and 82) tends to produce a couple of restoring moments acting on the side frame wear plates. This pair of moments tends to rotate the side frame in the direction of bogie alignment (i.e., to a position in which the bolster is perpendicular to or at right angles to the side frames). Thus, the bogie is capable of flexing, but as it flexes, the vibration dampers cooperate to act as offset members operating between the bolster and the side frames to resist parallelogram or rhombic deformation of the side frame relative to the bogie bolster and return the bogie to its non-deflected position.

The above explanation is given in the context of the carriage 20, which contains a spring group 54 having three rows facing the side frame columns 36. The restoring moment in this case is M _R , the moment pair of one pair of diagonally opposite damper springs in the corner of the spring group without the moment pair of the other diagonally opposite pair, and taking into account the sloping rear side of the vibration damper wedges, the restoring moment is a function of kc, the vertical spring constant of the spring coil , on which the vibration damper is mounted and which displaces it.

Although the embodiment shown is a 3x3 arrangement, there are various possible spring group arrangements such as 2x4, 3x3, 3:2:3 or 3x5. As shown, vibration dampers can be installed at each of the four corner locations. Groups of spring coils may have unequal stiffness if internal coils are used in some springs but not others, or if springs with different spring constants are used. In addition, the vibration damper springs may have a different undeformed length than the main spring coils. In other words, the vibration damper springs can be longer than the main spring coils. Thus, the preloaded deflection of the vibration damper springs will be greater than the preloaded deflection of the main springs. This is true for both a light car (i.e. empty) and a fully loaded car. Accordingly, the proportional difference (i.e. percentage change) in the spring excitation force in the damper springs will have a correspondingly smaller proportional change between the upper

- 9 045784 and the lower sections of the friction wedge stroke throughout its entire amplitude compared to the main springs. In this example, the corner springs 76, 78, 80 and 82 have a free height of 11 inches (27.5 cm), while the main springs are AAR D5 springs having a free height of 10.25 inches (25.6 cm).

The increased tendency to promote flexibly restorable squareness at the bolster-to-sideframe interface (i.e., by using four corner damper groups) tends to reduce the dependence on squareness at the axlebox-axle interface. This, in turn, may allow the use of a torsionally compatible (about a vertical axis) interface between the axle and axle box and provide a certain degree of self-steering.

The load-bearing plate, namely the side frame column wear plate 92 (FIG. 1a), is significantly wider than the end-to-end thickness of the side frames as a whole as measured, for example, at the axle bearings, and may tend to be wider than prior art plates. This additional width corresponds to the additional total damper span width measured entirely across the damper pairs, plus lateral movement as above, typically with a tolerance of 1 V ₂ (±) in. (3.75 cm) of bolster lateral movement relative to the side frame in any side from the undeviated central position. In other words, instead of having the width of one spring plus movement allowance, plate 92 may have the width of three springs plus 11/2 (±) inches (3.75 cm) movement allowance in either direction for a total movement with double amplitude 3 (±) inches. The bolster 24 has internal and external stops 106, 108, respectively, which limit the lateral movement of the bolster 24 relative to the columns 36 of the side frame. This movement allowance can range from ±11/8 to 1¾ inches and 1 ^3/16 to 1 9/16 inches ( ^2.8 to 4.4 cm and 3 to 3.9 cm) and can be specified, for example, by 11/2" (3.75 cm) or _11/4 " (3.1 cm) of lateral movement to either side of the neutral or centered position when the side frame is not deflected.

The lower ends of the springs of the entire group of springs, designated generally as 58, are located in the lower spring seat 52. The lower spring seat 52 may be arranged in the form of a tray with a rectangular peripheral projection upward. Although the cart 20 uses a spring group in a 3x3 arrangement, it is intended to be universal and present a number of options. They may be a 3x5, 2x4, 3:2:3 or 2:3:2 arrangement, or some other arrangement and may include a hydraulic shock absorber, or such other spring arrangement may be suitable for a given railcar application, for which this cart is intended.

Vibration damper wedges.

It was noted that the nature of the interaction of the friction lining with the wear plate of the side frame column changes the quality of movement of the bogie. To achieve a given ride quality, it is beneficial for the non-metallic wear surfaces of the non-metallic wear pads to wear relatively evenly instead of wearing disproportionately along one edge.

This wear is sensitive to the location of the contact point on the inclined side of the vibration damper wedge with the inclined surface of the bolster pocket. During operation, the damper wedges tend to move slightly in the pockets, such as when the side frames yaw, pivot, and roll relative to the bolster. These deviations may seem insignificant. In existing bogies, the radius of the convexity at the rear of the vibration damper is very small. It can have an effective value of about 60 inches (152.4 cm). In one type of cart it is known to be about 40 inches (100 cm) and the bulge is cylindrical in shape to create a line contact instead of a point contact. On the other hand, with point contact, the radius of the convex allows the vibration damper wedge to find its own fit in the vibration damper pocket and withstand the relative movement of the side frame during rotation about its vertical axis, rotation about its transverse axis and roll with less tendency to jam or seize. Using a radius of curvature of 60 inches (152.4 cm) was previously considered acceptable for this purpose, allowing the vibration damper wedge to find its own equilibrium position. Over time and use, a wear scar 182, which may also be referred to as a contact patch, may form in the area where the rear of the damper wedge repeatedly contacts the surface of the bolster pocket. This contact patch tends to wear as the surfaces are repeatedly pressed against each other. The contact patch reflects two degrees of freedom of the swinging surface. In other words, the contact patch extends along the rear part of the vibration damper wedge, which has a primary angle of inclination, and also in the transverse direction along the displacement at a secondary angle. A similar wear pattern forms in the bolster pocket. The wear of the non-metallic friction lining wear surface can be influenced by the forces it experiences, and the forces experienced by the non-metallic friction lining wear surface appear to correlate with the location and size of this 2DOF contact patch.

- 10 045784

Fig. 2a-2d and 3a-3h.

The vibration damper wedge is designated as reference numeral 120. Although the right vibration damper wedge is shown, the left vibration damper wedge has the same structure and is a mirror image of the right vibration damper wedge. Accordingly, the description of the right vibration damper wedge should be understood as describing both parts, taking into account their opposite directions. In this regard, the vibration damper wedge 120 is intended to provide an overview of the left and right vibration damper wedges 64 and 68.

The vibration damper wedge 120 has a housing 122. The housing 122 can be, and in the illustrated embodiment is manufactured from, a relatively common material such as ductile iron, cast steel, or cast iron. When viewed from the side, it has a generally triangular shape. There is a first surface or portion or element 124 extending vertically; a second surface or element or portion 126 extending horizontally; and a third element or surface or portion 128 extending generally along the slope and which may be considered as a hypotenuse element between elements 124 and 126, the three portions thus combining to form the above generally triangular shape. The vibration damper wedge 120 also includes a first end surface or end wall 132 and a second end surface or end wall 134. In this case, the first end surface 132 represents a large end surface (i.e., FIG. 3d), and the second end surface 134 represents represents a smaller end surface (Fig. 3c). The vibration damper wedge 120 has a primary angle α (alpha) visible in the side view of FIG. 3s. In the illustrated embodiment, the angle α is the same angle α as the angle of the coincident or corresponding or associated surface of the inclined face 74 of the bolster pocket, whether 60 or 62. It is possible that both planes do not have to be exactly parallel, but it is convenient to make them the same for both conceptual understanding and manufacturing. Angle α determines the primary angle of the bolster relative to the vertical plane when the vibration damper wedge is shown in a side view. The vibration damper wedge 120 also has a secondary vibration damper angle β. In the illustrated example, the secondary angle of the vibration damper wedge 120 is the same as the secondary angle β of the inclined surface of the bolster pocket, whether 60 or 62. It extends in the transverse direction and determines the lateral displacement of the vibration damper wedge 120 in the pocket. The true appearance of the secondary angle β is visible when viewed along the rear of the vibration damper wedge 120 in the inclined plane of the primary vibration damper angle α. This view is shown in Fig. 3h. Angle beta is the angle of the tangent plane at the point of contact, designated as the operating point WP, discussed below, relative to the perpendicular to the end walls or end surfaces 132, 134 in the plane of the angle α. Again, the angle β may differ slightly from the angle of the corresponding or associated bolster pocket, but for ease of conceptual understanding and ease of manufacture they can be generally considered the same.

Considering the secondary angle β, the end wall or end surface 132 is larger than the end wall or end surface 134, and the vibration damper wedge 120 is asymmetrical when viewed from behind or from above. The vibration damper wedge 120 also has a grip or holder or mounting member or holding device 130 that extends upward from the first portion or first member 124, the shape and purpose of which is as described below.

The vibration damper wedge 120 can be made as a solid casting. Alternatively, the vibration damper wedge 120 may be hollow, as shown. In other words, the housing 122 has an internal cavity 140 defined by members 124, 126, 128, 132 and 134. The internal cavity 140 may be, and as illustrated, divided into two subcompartments or chambers 136, 138 by a corner plate or baffle, or by an internal wall 150 The wall 150 may have a central opening or window 146. Each of the end surfaces 132 and 134 may have a triangular or generally triangular opening 142, 144, respectively.

When considering these elements, it will be seen that the front surface element or first element 124 is flat or generally flat and has a rectangular or generally rectangular flange 154 that extends peripherally around the panel or plate or wall 152. The plate or wall 152 extends from the side laterally between the end walls 132, 134, and upward and downward between the leading edge of the second member 126 and the leading and upper edge of the third member 128. The bead 154 and the wall 152 cooperate to form a pocket 156 into which the wear member 160 is mounted. This may be accomplished in various ways, namely, a cutout or notch or cavity or recess is created in the first member 124 to form a socket 156, the wall 152 forming the base or back of the socket 156, and the flange 154 forming a projection or stop formed by such way of deepening. The wear element 160 can be, and in this case is, a non-metallic friction lining. As can be understood, it has a non-metallic wear surface which, in use, slides up and down in frictional contact with the side frame column wear plate 92. The wear element 160 has a shape corresponding to, i.e. suitable for placement within the contour of the peripheral retaining bead 154. As shown, this shape is generally square or rectangular. The wear element 160 can typically be cast in place or held in place using epoxy resin or other bonding method. The wear element 160 has a vertical height h ₁₆₀ (in the z direction) and a lateral width w ₁₆₀ (in the y direction). It may be assumed that the locations of half of said height and half of said width coincide with the half-height and half-width locations of slot 156.

A mounting member 130 is formed on one side of the top edge or edge 162 of the bead 154 and projects or extends upward therefrom. It is in the form of an upwardly projecting member 164 in which a rearwardly extending pin 166 is formed by creating a semicircular recess or recess 168. Installing the vibration damper wedges in the bolster pocket may require some skill. To facilitate this process, the mounting member 130 is sized to be forward of the bolster pocket and project upward from the outer stop. The end of the bolster is located between the side frame column wear plates, so when the bolster is in position, a jig can be used to grip and lift the damper wedge in the bolster pocket while installing the springs. The clamping device is then removed to release the mounting element, and the vibration damper wedge is seated on the springs.

The second element 126 may, but not necessarily, be made in the form of a plate or wall 170 on which the spring seat 172 is mounted. In the example shown, the spring seat 172 is or includes a boss or downward extending protrusion 174 that is sized to approximate the inside diameter of the corner spring coil or vibration damper spring 176 in the spring group. In this description, reference numeral 176 is for general presentation. In other words, the vibration damper spring 176 may be any of the corner springs 76, 78, 80 or 82 above. For purposes of this description, although vibration damper spring 176 is referred to as a single spring, it should be understood that it can be, and in this case is, a double coil spring that has both an inner and outer coil. On the protrusion 174 there is a spring coil with orientation in the axial direction. That portion of plate or wall 170 that extends radially from protrusion 174 acts as a stop or limiter that determines the end of travel of the upper end of the spring and its vertical position depending on dynamic vertical load conditions. The protrusion 174 can be understood as a cylindrical boss having a vertical center line, which, when installed, is the same as the vertical center line of the vibration damper spring 176, designated CL ₁₇₆ . As may also be apparent from the previous discussion, the second element 126 is located at right angles to (ie, perpendicular to) the first element 124.

The third element 128 is an inclined element. Nominally it slopes at a primary angle of alpha, but is convex. The location of the tangent point of this convexity, which is the neutral point of contact when the car is at rest, is defined as the operating point WP. A steel wall, the outer surface of which defines the working surface 200 of the third portion 128, is designated 180. Within the housing 122, an internal partition 150 extends from the front wall 152 to the rear wall 180 and from both of these walls to the bottom plate or bottom wall 170. In this position, the internal partition 150 reinforces all three elements. The partition 150 is assumed to have the thickness t ₁₅₀ shown in FIG. 3a. In the embodiment shown, wall 180 lies above the center line CL ₁₇₆ and in the same vertical plane as the operating point WP. This plane, conventionally designated as 190, is defined as the plane in which lies the center line CL ₁₇₆ of the spring and the normal vector to the friction surface of the non-metallic friction element, namely the lining 160, i.e. a vector perpendicular to the wall 152. In other words, said plane is perpendicular to the friction element. It is called the reference plane. In the specific example shown, the reference plane may also be a plane located midway between the first and second end surfaces of the housing 122. For purposes of this description, three spaces need to be considered. There is a large space that may be referred to as a central region or central zone of surface 200 adjacent to and containing plane 190, which may include material lying in plane 190 and in a region of surface 200 that is within two partition thicknesses t ₁₅₀ from center line CL ₁₇₆ of the partition in the transverse or y direction. There is a narrower space that lies within the projected thickness of the partition 150. Finally, there is a narrower space in which the rolling contact point lies in the plane 190 or within 1/8 inch (3 mm) to either side thereof, or such that the contact surface of the male and female elements at the point of rolling contact under load lies on or above the plane 190. In such circumstances, one skilled in the art could reasonably describe the operating point WP as lying in or approximately lying in the plane 190. The operating point WP is in the tangent plane of the inclined surface 200. In other words, assuming that the operating point WP is in the reference plane, and that the curvature of the surface 200 is spherical for simplicity, the tangential plane is constructed to pass through the operating point WP at an angle of in accordance with the primary angle α and in accordance with the secondary angle β. In the illustrated embodiment, this tangential plane, in the most conceptually simple example, will also be the plane of the inclined surface of the bolster pocket. Because it is a surface with a rolling contact point, adjacent regions of the surface 200 are located outside this tangent plane, and the normal to the tangent plane at the contact point defines the radius of the spherical surface. The center of curvature at the beginning of this radius will be on one side of the basal plane 190, with the radius offset from it by an angle β taken in the plane of the angle α. The location of WP can be considered to be within a radius of 1 inch (2.5 cm) from the reference point DP, measured on surface 200. In other words, in terms of scaling to the size of the damper wedge 120 itself, WP is within V ₄ of the width of the damper wedge 120 fluctuations from DP; or, again in other words, WP is within % height of non-metallic wear pad 160 from DP. In some embodiments, when these parts are at the rolling contact point under load, they are within the contact point width of the reference plane 190.

In terms of physical operation, the forces applied to the body 122 include the normal force applied to the side frame column wear plate 92, the friction force directed upward or downward in the plane of the wear element 92, the vertical reaction force in the spring seat, and the angular force reaction applied to the inclined surface 200. When the cart 20 is at rest, on a level path, in equilibrium, the point of application of the reaction on the inclined surface is at the operating point WP. During dynamic operation, when the bolster moves up and down relative to the side frame column and when it rotates around its lateral axis, rotates around its vertical axis and rocks the side frame, the actual instantaneous contact point deviates from the nominal operating point WP. The range of movement of the side frame when rotating around its transverse axis is small and can be of the order of ±2°. The range of deflection when rotating around its vertical axis is also small, about ±3°. The range of deviation during roll is also small, similarly of the order of ±3°. Deflection of the vibration damper wedge 120 during operation occurs in these ranges and creates a wear pattern, which may also be referred to as a contact patch 182, on the inclined surface 200 of the vibration damper wedge 120, where the rolling contact actually occurs, and forms a wear area as on the surface 200. and on the inclined contact surface of the bolster pocket. The contact within the wear scar varies randomly or largely randomly as the cart moves, with path discontinuities considered to be a white noise input function of time. The contact patch is a distinctive feature of the contact interaction with two degrees of freedom of the vibration damper wedge 120 and the bolster pocket 60 (or 62 depending on the specific case), extending both along the curvature of the inclined surface up the slope and down the slope, and in the left and right lateral or transverse directions, and has a circular or elliptical shape associated with the rolling contact point.

For a given angular deflection of the side frame 26 resulting from rotation about its vertical axis or rotation about its transverse axis, the deviation of the instantaneous rolling contact point from the operating point WP is a function of the curvature of the rear portion of the inclined surface 200. If the curvature has a large radius, such as the radius along default 60 inches (152.4 cm), as used in some prior art vibration dampers, lateral deflection in the y direction when rotating about its vertical axis, or arc displacement deflection in the upslope or downslope direction when rotating about its transverse axis will be relatively large. Where this radius is smaller, the deviation will be less. In this discussion, curvature along a slope does not have to be the same as curvature across a slope. They can be different, as in an ellipse. However, it may be convenient for them to be the same, so that the inclined surface is a partial spherical segment of the same radius. In any case, it has been found that the wear scar area becomes smaller as the radius of curvature becomes smaller, and that vibration damper performance and life improve when the radius is less than 45 inches (112.5 cm). This improvement is much more noticeable when the radius is less than 40 inches (100 cm). From observation, it has been found that it is useful for the radius to be in the range of 15 to 30 inches (38 to 76 cm). To this end, the illustrated embodiment is intended to represent a radius of 20 inches or approximately 20 inches ±1/2 inch or ±1 inch (50 cm or approximately 50 ±1.3 or ±2.5 cm), as appropriate. This can be expressed in different ways. In the illustrated embodiment, the radius r ₁₈₂ of the wear scar 182 is 2 inches (5 cm) or less. Expressed in parametric terms, the radius of the wear scar 182 is less than half the width of the vibration damper wedge 120. Alternatively, the radius of the wear scar 182 is less than 10% of the radius of curvature of the surface 200. In the case where the curvatures have different radii to produce an ellipse having a minor axis and a major axis, these axes can replace the radius r ₁₈₂ .

As stated above, in the example shown, the surface 200 is formed at a curvature. As shown in FIG. 3f, the vertical axis of the vibration damper spring 176 intersects the surface 200 at the slope reference point DP. The location of the WP point relative to the DP point may vary depending on geometries

- 13 045784 ries of surface curvature 200. The point contact WP may be in the range of '/ ₈ to 5/8 of the displacement rearward in the x direction from the non-metallic friction element 160. In this example, the term rearward from point DP also means downhill. This offset can also be expressed in terms of the distance along the arc length along the surface 200 from the point DP. It can also be expressed as a fraction of the displacement distance from the plane of contact of the friction element with the wear surface of the side frame column, i.e. this surface is in the same plane as the front surface of the non-metallic wear element. In the examples shown _, this parameter range may be from about _1/32 to ^5/32 of the total surface height of the non-metallic wear element 160.

In one embodiment, the operating point WP is offset rearward (i.e., downhill) in the x-direction (from the front surface of the non-metallic friction lining) by a distance ranging from about _1/4 inch (6 mm) to about ^5/8 inch ( 16 mm). In one particular embodiment, it is offset by a distance of about 0.56 inches or ^9/16 inches (1.4 cm). Assuming that these distances are proportional to the offset of the front surface of the friction lining, the leading edge of the bead 154 is offset forward by approximately 2 ^5/8 inches (6.6 cm). As shown in FIG. 3f, in another embodiment, the non-metallic surface is offset from said axial center line by a first distance x1; said operating point is offset from said center line by a second distance x2; and the ratio x1:x2 is in one of the ranges (a) from 21:2 to 21:8; and (b) from 10:3 to 40:3. In one embodiment, the ratio is about 5:1. In other words, the non-metallic wear surface has a total height of y160. In one embodiment _{, the operating point WP lies in the range of 3/8 to 5/8} ^{y 160} _along the height of said non-metallic wear surface.

In the mechanical system described above, a single-point rolling contact relationship is established between the inclined surface of the vibration damper wedge and the corresponding mating inclined surface of the bolster pocket. The same relationship can also be established by inverting the relationship such that the flat surface is the inclined surface of the vibration damper wedge 120 and the curvature surface is the surface of the bolster pocket. In other words, in a mathematical sense, it is to some extent indifferent which surface is the male surface and which is the female surface. In a further alternative, both surfaces may be formed on a curve, and one of the surfaces may be cylindrical rather than spherical. However, in the illustrated embodiments, the mating surfaces are machined surfaces, and manufacturing practicalities may cause a flat, planar surface to be formed in the bolster pocket and a curved surface to be formed on a smaller, lighter, less bulky, more easily machined vibration damper wedge. However, this description is intended to cover both possibilities as equivalents under the doctrine of equivalents.

The vibration damper wedge 120 may provide frictional vibration damping with little or no stick-slip behavior, but rather may provide frictional vibration damping for which the static and dynamic coefficients of friction are equal or differ by only a small amount (less than about 20%, perhaps , less than 10%). The wedge 120 may be used in the carriage 20 in combination with a bidirectional bearing adapter, as shown in FIG. 4a-4e described herein. Wedge 120 may also be used in a four-corner damper arrangement, such as in bogie 20. Wear member 160 may be made of brake lining material and the column wear plate may be made of high strength steel.

The vibration damper wedge 120 has a base area having a vertical extent slightly greater than the vertical extent of the inclined seat surface 74. The inclined seat surface 74 is inclined at a primary angle α and a secondary angle β. This allows for movement and wear. The mounting bracket of the mounting member 130 is mounted on the top edge and is visible from above after installation.

In this embodiment, the vertical surface of the first portion of the first element 124 of the friction damper wedge 120 has a load-bearing surface having a static friction coefficient μ _s and a dynamic or kinetic friction coefficient μ _k that may tend to exhibit little or no stick-slip behavior when interaction with the wear surface of the wear plate 92. In one embodiment, the friction coefficients differ from each other by an amount within 10%. In yet another embodiment, the friction coefficients are substantially equal and may be substantially free of stick-slip behavior. In one embodiment, dry friction coefficients of friction may be in the range of 0.10 to 0.45, may be in the narrower range of 0.15 to 0.35, and may be about 0.30. The friction damper wedge 120 may have a frictional surface coating or may be an associated pad, such as 160, having these frictional properties. The associated pad 160 may be a polymeric pad or cover. In yet another embodiment, the static and dynamic friction coefficients are essentially equal. The dynamic coefficient of friction can range from 0.10 to 0.30 and can be around 0.20.

Fig. 4a-4e.

The swing mating surface of the bearing adapter may have a convex or concave curvature, like that of a swing carriage, by which a rolling contact on the rocker arm allows the side frame to swing laterally. The interface of the bearing adapter to the axle box may also have an anteroposterior curvature, either a bulge or a dimple, and for a given vertical load this bulge or dimple may tend to produce a more or less linear resistance to longitudinal deflection similar to that can be created by a spring or an elastomeric pad.

The rigidity of the pendulum is directly proportional to the weight of the pendulum and for small angles of deflection can be taken as proportional to the angular deflection, which in geometric terms approaches /=kx. The pendulum may tend to maintain an overall proportionality between the weight carried by the wheel and the stiffness of the self-steering mechanism as the load increases, and thus provide proportional steering whether the car is empty or fully loaded. These stiffnesses are geometric stiffnesses, not spring stiffnesses.

In fig. 4a-4e show an embodiment of the bearing adapter and axle box assembly. The bearing adapter 44 has a lower portion 212 that sits on a bearing 46 of the axle 48. The bearing adapter 44 has a top portion 214 that includes a male bearing adapter mating portion 216. A mating female rocker socket interface portion 218 is mounted within the ceiling 220. The top fitting 218 may be a flat plate surface. When the side frames are lowered over the wheelsets, the end cutouts or channels 228 located between the bearing adapter miter stops 232 are located between the corresponding side frame axle box jaws 230. Thus, when the side frames are in place, the bearing adapter 44 is captured in position with the male and female mating adapter portions (216 and 218) in mating engagement.

The bearing adapter 44 may have a central adapter housing portion 254 that is shortened longitudinally and the internal distance between the corner stops slightly increased to allow installation of an auxiliary centering device or centering element, or a centrally offset restoring element in the form of, for example, elastomeric shock absorbing pads. , such as those identified as resilient pads or members 256. Features 256 may be considered a form of restorative centering member and may also be referred to as damper pads or bumper pads.

As shown in FIG. 6a-6e, the elastic members 256 may have the general shape of a channel having a central or rear or transverse or septal portion 281 and a pair of left and right flanking wing portions 282, 283. The wing portions 282 and 283 may include downwardly and outwardly extending ends, which may have an arcuate bottom edge such that it can be seated on the bearing housing. The internal width of the wing portions 282 and 283 may be such as to provide a tight fit around the sides of the thrust blocks 280. A blade portion 285 extending laterally along the top edge of the baffle portion 281 may be positioned in a radially rounded recess 284 between the top edge of the thrust blocks 280 and the end of fitting 40 of the axle box socket. The inner side edge of the paddle portion 285 may tend to be tapered or lightened to accommodate and fit adjacent to the end of the axle box 40. In FIG. 5a and 5b show views of the bearing adapter 44 and the elastomeric shock-absorbing pads 256 as an insertion assembly between the bearing 46 and the side frame 26.

The bearing adapter 44 may also have an underside cutout or groove 292 in the form of a pair of laterally tapering bladed recesses, cavities or cutouts 294, 296 separated by a central bridge-like region 298 having a deeper section and sides that taper into the cutouts. 294, 296. The cutouts 294, 296 may have a major axis that extends transversely relative to the bearing adapter itself, but once installed, extends axially relative to the axis of rotation of the underlying bearing. The absence of material in the cutouts 294, 296 forms an overall H-shaped relief in the circumferential surface 300, which is seated on the outer side of the bearing 46, and in which the two side areas or legs of the letter H form pads or pads 302, 304, connected by a relatively narrow bridge, namely, the bridge-like region 298. To the extent that the lower surface of the lower portion of the bearing adapter 44 conforms to an arcuate profile such as that in which a bearing housing can be housed, the cutouts 294, 296 may tend to extend or extend predominantly along the top of the profile between the lying on both sides with overlays or platforms. This configuration may tend to spread the load of the rocker arm's rolling contact point onto the pads 302, 304 and therefore onto the bearing 46. Bearing life may depend on the peak load on the rollers. By leaving space between the bottom

- 15 045784 side of the bearing adapter and the upper center portion of the bearing housing above the bearing raceways, the cutouts 294, 296 may tend to prevent concentrated vertical load transfer to preferentially the upper rollers in the bearing. Instead, it may be advantageous to distribute the load across multiple rollers in each raceway. This may be facilitated by the use of spaced spacers or pads, such as pads 302, 304, that are mounted on the bearing housing. The central bridge region 298 may be located above a portion of the bearing housing that does not have a raceway beneath it, rather than directly above one of the raceways.

The male portion 216 is formed such that the overall upwardly facing surface 242 has both a first curvature r1 to provide longitudinal swing (FIGS. 4d, 4e) and a second curvature _r2 (FIGS. 4b, 4c) to provide swing ( i.e. rotational movement of the side frame) in the transverse direction. Likewise, in general, the female portion 218 has a surface having a first radius of curvature R1 in the longitudinal direction and a second radius of curvature R2 in the transverse direction. The interaction of these surfaces with the radii r1 and R1 of curvature provides the possibility of longitudinal rocking movement with resistance proportional to the weight on the wheel. In other words, the resistance to angular deflection is proportional to the weight, rather than being a fixed spring constant. This can lead to passive self-steering both in a light car and in fully loaded conditions. This relationship is shown in FIG. 4d and 4e. In fig. 4d shows the centered or non-deflected position of the longitudinal swing elements corresponding to the rest state. In fig. 4e shows the swing elements in their state of maximum longitudinal deflection. In fig. Figure 4d presents the local minimum potential energy condition for the system. In fig. 4e shows a system in which the potential energy is increased by the work done by the force F acting longitudinally in the horizontal plane through the center CB of the axle and bearing, will tend to gradually increase the height of the axle box. In other words, as the axle is forced to deflect by force, the swing motion may tend to lift the car and thereby increase its potential energy.

In general, the deflection can be measured by either the angular displacement Θ1 of the axial centerline or the angular displacement of the rocker contact point at a radius r1, denoted by θ ₂ . The end surface 234 of the bearing adapter 44 is inclined at an angle η relative to the vertical. A typical range for η may be about 3° of arc. A typical maximum value of δ _long may be about ±3/16 inch (0.5 cm) on either side of the resting vertical centerline.

Likewise, as shown in FIG. 4b and 4c, in the transverse direction, the interaction of these surfaces with radii r2 and R2 of curvature can lead to a lateral rocking movement, as can be the case with a rotary carriage. In fig. Figure 4b shows the rest-centered position of the side-swing system with minimum potential energy. In fig. 4c shows the same system in a lateral deflection state. In this case, δ2 is approximately equal to (L _pend u _l u _m -r ₂ )S _in φ, where for small angles S _inφ is approximately equal to φ. L _pend ulum can be taken as the difference in height at rest between the center of the lower spring seat 52 and the contact surface between the male and female portions 216 and 218.

This junction of the bearing adapter with the axle box socket is displaced under the influence of gravity acting on the pendulum towards the central position or rest position, in which there is a local minimum of potential energy in the system. The fully deflected position shown in FIG. 4c may correspond to a deviation from vertical of the order of less than 10° (and preferably less than 5°) on either side of the center, the actual maximum being determined by the distance from the stops 106 and 108 to the plate 92. Although in general R1 and R2 may vary in this way If the enclosing surface is an outer segment of the torus, it may be desirable for R1 and R2 to be the same, i.e. so that the supporting surface of the female fitting is made in the form of a part of a spherical surface that has neither a major nor a minor axis, but is simply formed on the basis of a spherical radius. R1 and R2 provide a tendency towards self-centering. This trend can be quite mild. Although r1 and _r2 may be the same, so that the convex surface of the bearing adapter (or axle box if the relationship is inverted) is part of the spherical surface, in general r1 and _r2 may differ. It may also be noted that, given that the system may tend to return to a state of local minimum energy (i.e., which is self-healing during normal operation), in the limiting case one or both of the radii R1 and R2 may be infinitely large, so that either a cylindrical segment is formed, or, when both are infinitely large, a flat surface can be formed. Another alternative could be that r1=r ₂ , and R1=R ₂ . In one embodiment, r1 may be the same as _r2 and may be about 40 inches (±5 inches) (100 (±12.5) cm), and R1 may be the same as R2, and both may be infinite, so that the inner surface is flat.

The radius r1 of curvature of the male longitudinal rocker arm may be less than 60 inches

- 16 045784 (152.4 cm) and can range from 5 to 50 inches (12.5 to 125 cm) and can be about 40 inches (100 cm). R1 may be infinite or may be less than 100 inches (250 cm) and may be in the range of 25 to 60 inches (62.5 to 152.4 cm) or in the narrower range of 30 to 40 inches (75 to 100 cm), depending on the radius r _b The radius r ₂ of curvature of the male side rocker can range from 30 to 50 inches (75 to 125) and can be about 40 inches (100 cm), as in the illustrated embodiment. R2 may be infinite, so that the plate is flat, or alternatively it may be about 60 inches (152.4 cm). If a flat female rocker surface is used and a spherical male surface is used, the radius of curvature of the male surface may be in the range of about 20 to about 50 inches (about 50 to about 125 cm) and can be in the narrower range of 30 to 40 inches ( from 75 to 100 cm). Various combinations are possible depending on the load, intended use and rocker materials. In each case, the mating male and female surfaces of the rocker arms can be selected to produce a physically based pairwise mating in terms of expected load, expected load history and service life. They may change.

The male and female surfaces may be reversed such that the female engagement surface is formed on the bearing adapter and the male engagement surface is formed on the axle box seat. One of the mating parts or surfaces is part of the bearing adapter, and the other is part of the axle box support.

The swing unit at the junction of the wheelset and the side frame tends to maintain itself in a centered state. There is a spatial relationship of the assembly formed by (a) a bearing adapter, such as a bearing adapter 44; (b) centering elements, such as, for example, elastic elements 256; and (c) axle box jaw thrust blocks 280. When the resilient element 256 is in place, the bearing adapter 44 may tend to be centered relative to the jaws or thrust blocks 280. Once installed, the shock absorber (element 256) fits snugly against the axle box jaw thrust lug and may be positioned adjacent to the end wall of the bearing adapter and between the corner bearing adapter stops with an easy interference fit. The shock absorber is sandwiched between the thrust lug and the bearing adapter and establishes a spaced relative position between them; provides initial central positioning of mating rocker elements; and also provides restorative displacement. Although the bearing adapter 44 may still swing relative to the side frame 26, such rocking may tend to deform (usually locally compress) a portion of the member 256, and the member 256 being elastic may tend to push the bearing adapter 44 toward the center position, regardless of whether there is a lot of weight on the swing elements or not. The elastic element 256 may have a restoring force-deflection characteristic in the longitudinal direction that is substantially less rigid than the force-deflection characteristic of a fully loaded longitudinal rocker arm, (perhaps one or two orders of magnitude less) so that in a fully loaded state of the car, element 256 may not have a tendency to significantly change rocking behavior. In one embodiment, element 256 may be made of polyurethane.

The rolling contact surface of the bearing is in a state of local minimum energy when centered under the corresponding seat, and it is preferable that the mating surface of the rolling contact has a radius that can promote self-centering of the male rolling contact element.

This can be expressed in different ways. In cylindrical polar coordinates, the long axis of the wheelset axle can be considered as the axial direction. There is a radial direction, measured perpendicular to the axial direction, and there is an angular circumferential direction, mutually perpendicular to both the axial direction and the radial direction. There is a place on the rolling contact surface that is closer to the axis of rotation of the bearing than any other place. This determines the position of the rest equilibrium or local minimum of potential energy. Since the radius of curvature of the rolling contact surface is greater than the radial length L between the bearing axis of rotation and the minimum radius location, the radial distance as a function of circumferential angle θ will increase to either side of the minimum radius location (or alternatively the location of the minimum radial distance from the bearing axis of rotation lies between areas of greater radial distance). Thus, the slope of the function r(θ), namely dr/dθ, is zero at the minimum point and is such that r increases with angular displacement from the minimum point to either side of the location of the minimum potential energy. If the surface has complex curvature, both dr/dθ and dr/dL are zero at the minimum point and are such that r increases on any side of the minimum energy location from all sides of the minimum energy location and is zero at that location. This may be true regardless of whether the rolling contact surface on the bearing is male

- 17 045784 surface or surrounding surface. The rolling contact surface has a radius of curvature or radii of curvature if complex curvature is used, i.e. is or are greater than the distance from the location of the minimum distance from the axis of rotation, and the rolling contact surfaces are not concentric with the axis of rotation of the bearing. Another way of expressing this is to note that there is a first location on the rolling contact surface of the bearing that is radially closer to the axis of rotation of the bearing than any other location thereon. The first distance L is defined between the axis of rotation and this nearest location. The bearing surface and the axle box surface have radii of curvature and are conjugate in a male and female relationship, one radius of curvature being the male radius of curvature r1, and the other radius of curvature being the female radius of curvature R2, (whichever it may be) . In this case, r1 is greater than L, R2 is greater than r ₁ , and L, r ₁ and R2 correspond to the formula L ^-1 -(rf1-R ₂ ^-1 )>0, and the surfaces of the rocker arms are designed to interact to ensure self-steering.

Geometry of a compound pendulum.

The rocker arms shown and described herein may employ swing elements that define composite pendulums, i.e. pendulums in which the radius of the covered rocker arm is not zero, and there is an assumption about the interaction of rolling (as opposed to sliding) with the covering swing element (rocker arm). The embodiment of FIG. Figure 4a shows a bidirectional compound pendulum. The characteristics of these pendulums can affect both the lateral stiffness and the self-steering of the longitudinal rocker arm.

Lateral suspension stiffness may tend to reflect (a) side frame stiffness between (i) the bearing adapter and (ii) the lower spring seat (i.e. when the side frames rock laterally); (b) the lateral deflection of the springs between (i) the lower spring seat and (ii) the upper spring seat attached to the bogie bolster; and (c) the moment between (i) the spring seat in the side frame and (ii) the upper attachment of the spring to the bogie bolster. The lateral stiffness of the spring groups can be approximately 1/2 the stiffness of the vertical spring.

To estimate the lateral stiffness, the formula ktruck ⁼ 2 X [(ksideframe) ¹ + (kspring shear) \ where ksideframe [kpendulum+kspring moment] can be used;

k _spring shea - lateral spring constant for the spring group during shear;

k _pendulum is the force required to deflect the pendulum per unit of deflection, measured at the center of the lower spring seat;

k _spring moment - the force required to deflect the lower spring seat by one unit of lateral deflection against the torque caused by the unequal compression of the inner and outer springs.

In a pendulum, the relationship between weight and deflection is approximately linear for small angles, similar to F=kx in a spring. The lateral constant can be defined as k _pendulum =W/L, where W is the weight and L is the length of the pendulum. The approximate equivalent length of the pendulum can be defined as L _e q=W/k _pendulum . W is the spring weight on the side frame. For a cart with L=15 inches (37.5 cm) and a bulge radius of 60 inches (152.4 cm), L _eq may be about 3 inches (7.5 cm). For a rotating cart, L _eq can be more than twice this value.

The formula for the longitudinal (i.e. self-steering) rocker arm shown in FIG. 4a can also be defined as

F/6i _ong = kiong = (W/L)[[(l/L)/(l/n - 1/Ri)]-1] where k _long is the longitudinal constant of proportionality between the longitudinal force and the longitudinal deflection of the rocker arm;

F is the unit of longitudinal force applied to the center line of the axle;

δ _long - unit of longitudinal deviation of the center line of the axis;

L is the distance from the center line of the axis to the top of the male part 216;

R1 is the longitudinal radius of curvature of the enclosing cavity in the axle box socket 38;

r ₁ is the longitudinal radius of curvature of the convexity of the male part 216 on the bearing adapter.

In this relationship, R1 is greater than r _b and (1/L) is greater than [(1/r ₁ )-(1/R ₁ )], and, as shown in the drawing, L is less than either r ₁ or R ₁ . In some embodiments described herein, the length L from the center of the axle to the top of the bearing adapter surface at the center rest position may typically be from about 5% to 6 inches (about 14.4 to 15 cm) (±) and may be ranging from 5 to 7 inches (12.5 to 17.5 cm). Bearing adapters, axle bearings, side frames and bolsters are typically made of steel. The present inventor believes that the rolling contact surface may preferably be made of tool steel or the like.

In the lateral direction, the approximation for small angular deviations looks like kpendulum ⁼ (Rg/bg) ⁼ (W/Lpend.)[[(l/Lpend.)/((l/RRocker) ^- (1/Rseat))] + 1] ,

- 18 045784 where k _pen d _ulum is the lateral rigidity of the pendulum;

F2 is the force per unit of lateral deflection applied to the lower spring seat;

δ ₂ - unit of lateral deviation;

W is the weight on the pendulum;

Lpend. - the length of the undeflected pendulum between the contact surface of the bearing adapter and the lower part of the pendulum in the spring seat;

R _Rocker =r ₂ - lateral radius of curvature of the rocker surface;

R _Seat =R ₂ - lateral radius of curvature of the rocker socket.

In cases where R _Seat and RR _ocker are of the same magnitude and are not excessively small relative to L, the pendulum can have a relatively large lateral deflection constant. If the radius RSeat is large compared to L or RR _ocker , or both, and can be approximated as infinite (i.e. in the case of a flat surface), this formula simplifies to

Using this number in the denominator and the calculated weight in the numerator gives the equivalent pendulum length Leq=W/kpendulum.

The cart may not have lateral unsprung crossbar braces, laterally extending parallel bars or diagonally crossing frame braces or other unsprung stiffeners. In these embodiments, the carts may have four corner groups of vibration dampers actuated by each spring group.

Friction surfaces.

As explained in WO 2005005219, the dynamic response can be quite weak. It may be desirable to replace a physically locked relationship with a relationship that allows the cart to flex in a non-rectangular manner, provided that the displacement tends to return the cart to its rectangular position, which can be achieved by using a pair of dual vibration dampers with a higher moment of resistance compared to single dampers. hesitation. While the use of side soft rockers, low-slip dampers, quad dampers and self-steering may be beneficial in their own right, they also appear to be linked in subtle and unexpected ways. Self-steering may work better where there is a reduced tendency for slippage in the vibration dampers. Lateral rocking in swing mode may also function better where vibration dampers have a reduced tendency to stick-slip. Lateral sway in swing mode may tend to perform better if the sway dampers are installed in a quad arrangement. Paradoxically, bogie wobble may not be significantly worsened when the rigid linkage of the cross member or frame brace is replaced by four corner dampers (obviously making the bogie softer rather than stiffer), and when the dampers are less prone to stick-slip. The combined effect of these traits may have unexpected relationships.

As described herein, there is a friction damping surface between the bolster and the side frames. Either the side frame columns or the vibration damper (or both) may have a low or controlled friction bearing surface, which may include a hardened wear plate that can be replaced if worn or broken, or which may include a consumable covering or pad or lining. This load-bearing surface of the motion-calming friction damping element can be obtained by treating the surface to achieve the required coefficients of static and dynamic friction by applying a surface coating, insert, lining, brake pad or brake lining, or other treatment. Such a pad or lining may have a polymer-based matrix or a composite matrix filled with a mixture of metal particles or other materials to provide specified friction characteristics.

This friction surface may, when used in combination with an opposing load-bearing surface, have a static friction coefficient μ _s and a dynamic or kinetic friction coefficient μ _k . Ratios may vary depending on environmental conditions. For the purposes of this description, the friction coefficients shown will be considered under dry day conditions at 70°F. In one embodiment, for dry friction, these coefficients of friction may range from 0.15 to 0.45, or may range more narrowly from 0.20 to 0.35, and in one embodiment, may be approximately 0.30. In one embodiment, the cover or pad, when used in combination with the opposing side frame column load-bearing surface, can provide static and dynamic friction coefficients at the friction surface that differ from each other within 20%, or in a more restricted approach, within 10%. In yet another embodiment, the static and dynamic coefficients of friction are substantially equal.

- 19 045784

The vibration damper may be subjected to a friction-specific treatment, such as a coating, lining or lining on the vertical friction surface. In one embodiment, the coefficients of static and dynamic friction at the friction surface may be about 0.3 and may be approximately equal to each other.

Spring groups.

Spring groups can have different spring arrangements. It may be useful to have upward and downward damping forces that are not too different from each other, and which in some cases may tend to be approximately equal. Frictional forces on vibration dampers can vary depending on whether the vibration damper is loaded or unloaded. The damper wedge angle, friction coefficients and spring force under the vibration damper wedges can vary. The damper wedge is loaded as the bolster moves down in the side frame window as the force in the springs increases and therefore the force on the damper wedge increases. Likewise, the vibration damper wedge is relieved when the bolster moves toward the top of the frame side window as the force in the springs decreases. The equations can be written as under loading

Fd = F _s (SoYF)-sch) (1 +(μ ₈ -με)€θί(φ)+μ ₈ με) during unloading

Fd = gcFs (Cot(dO+n _s ) (1+(Is-I ₈ )Soff)+i8Is), where Fd is the friction force on the side frame column;

Fs is the force in the spring;

μ _s - coefficient of friction on the angular inclined surface of the bolster;

μ _c - friction coefficient on the side frame column;

f - internal angle between the corner surface of the bolster and the friction surface resting on the column.

For a given angle, the friction load coefficient Cf can be defined as Cf=F _d /F _s . This load factor C _f will tend to differ depending on whether the bolster is moving up or down.

In some embodiments, the empty and fully loaded states may have different vertical spring rates. For this purpose, springs of different heights can be used, for example, to obtain two or more vertical spring rates for the entire spring group. Thus, the dynamic response of a light car may be different from the dynamic response of a fully loaded car that uses two spring rates. Alternatively, if three (or more) spring rates are used (outer, inner, inner-inner), there may be an intermediate dynamic response at half load. In one embodiment, each set of springs may have a first combination of springs that have a free length of at least a first height, and a second set of springs in which each spring has a free length that is less than the second height, wherein the second height is less than first height by a distance δ1, so that the first group of springs will have a compression range between the first and second heights, in which the spring constant of this group has a first value, namely the sum of the spring constants of the first group of springs, and a second range, in which the spring constant of this group greater, namely, equal to the spring stiffness of the first group plus the stiffness of at least one of the springs, the free height of which is less than the second height. Different spring rates can provide corresponding different damping modes.

For example, in one embodiment, a car having its own sprung weight (i.e., the weight of the car body without cargo, excluding the unsprung weight of elements below the main spring, such as the weight of side frames and wheel sets) in the range of about 35,000 to about 55,000 pounds ( from about 15,900 kg to about 25,000 kg) (±5,000 lb (2,270 kg)), may have spring groups in which the first portion of the springs has a free height greater than the first height. The first height may range, for example, from about 9% to 101/4 inches (24.4 to 25.6 cm). When the car sits empty on its bogies, the springs are compressed to this first height. When the car is operated in an unloaded condition, this first part of the springs may tend to determine the dynamic response of the car during vertical rebound, rotation and lateral rebound, and side-to-side rocking and can influence the wobble behavior of the car. The spring rate in this first mode can range from 12,000 to 22,000 lb/in (844 to 1547 kgf/cm ² ) and can range from 15,000 to 20,000 lb/in (1055 to 1406 kgf/cm ² ).

When the car is heavily loaded, for example, when the combination of the dead weight and the cargo spring weight exceeds a threshold value, which may correspond to a value per car ranging from perhaps 60,000 (27,125 kg) to 100,000 lb (45,360 kg) (i.e., 15,000 to 25,000 pounds

- 20 045784 (6804 to 11430 kg) per spring group with symmetrical load at rest), the springs can be compressed to the second height or lower. This second _height may range from perhaps 81/2 to 9 ^3/4 inches (21.3 to 24.4 cm). At this point, the spring mass is sufficient to begin to deflect the other portion of the springs in the common spring group, which may be some or all of the remaining springs, and the spring constant of the combined group of currently compressed springs in this second mode may differ and may be greater than spring stiffness in the first mode. For example, this higher spring rate may be in the range of about 20,000 to 30,000 lb/in (1406 to 2109 kgf/cm ² ) and may be designed to provide a dynamic response when the sum of the dead weight and the load weight exceeds a mode change threshold.

In various embodiments of bogies, such as bogie 20, the resilient interface between each side frame and the end of the associated bogie bolster may include a quadrangular wedge damper arrangement and a 3x3 spring assembly. These groups may have damper wedges having primary angles ranging from 30 to 60°, or a narrower range from 35 to 55°, an even narrower range from 40 to 50°, or may be selected from a set of angles 32, 36, 40 or 45°. Vibration damper wedges have modified friction surfaces, such as non-metallic surfaces.

The combination of damper wedges and side springs may be such that the spring rate under the side springs is 20% or more of the total spring rate of the spring groups. It can range from 20 to 30% of the total spring stiffness. In some embodiments, the combination of wedges and side springs may be such as to provide a total frictional force for the vibration dampers in the group for a fully loaded car when the bolster moves downward that is less than 3,000 pounds (1,360 kg). In other embodiments, the arithmetic sum of the frictional forces in the upper and lower directions of the vibration dampers in the group is less than 5500 pounds (2495 kg).

Combinations and permutations.

Features of various figures may be mixed and matched without departing from the essence or scope of protection of the invention. To avoid redundant description, it should be understood that various vibration damper wedge configurations may be used with 2x4, 3x3, 3:2:3, 2:3:2, 3x5, or other spring groups. Likewise, there are a large number of possible combinations and permutations of vibration damper wedge arrangements and bearing adapter arrangements. In this regard, it may be understood that various features can be combined without further increasing the number of drawings and descriptions.

In various bogie embodiments, stops may be shown herein mounted on the bolster inside and outside the wear plates on the side frame columns. In the embodiments shown herein, the clearance between the stops and the side plates is desirably sufficient to allow at least ^3/4 inch (1.9 cm) of lateral movement of the railroad car bogie bolster relative to the wheels in either direction. from the neutral position, preferably allows movement of more than 1 inch (2.5 cm) to either side of the neutral position and may allow movement ranging from about 1 or 1 ^3/8 inches _to about 15/8 inches or 1 ^9/16 inches (from about 2.5 or 3.4 cm to about 4.1 or 3.9) either side of the neutral position.

In each of the bogies shown and described herein, the overall ride quality may depend on the relationship between the design and physical properties of the spring group or the design and properties of the vibration damper, or both, in combination with the dynamic properties of the bearing adapter interface with the axle box. . It may be advantageous if the lateral stiffness of the side frame acting as the pendulum is less than the lateral shear stiffness of the spring group.

The embodiments of the carts shown and described herein may vary depending on their suitability for different types of service. Bogie performance can vary greatly depending on expected load, wheelbase, spring rates, spring layout, pendulum geometry, vibration damper layout, and vibration damper geometry.

Various embodiments of the present invention have been described in detail above. Since changes in or additions to the best practice described above may be made without departing from the spirit, principle or scope of the invention, the present invention is not to be limited by these details but only by the appended claims.

-

Claims (35)

CLAIM 1. Wedge (64, 66, 68, 70, 120) of a vibration damper for a bogie (20) of a railway car, designed to interact with the pocket (60, 62) of the bolster, wherein said wedge of the vibration damper contains a friction element (90, 160 ) which, when in use, movably interacts with the wear surface (92) of the column (36) of the side frame of the bogie (20) of the railway car, and the specified friction element (90, 160) has a non-metallic wear surface; a spring socket (172, 174) which, when in use, interacts with a spring (76, 78, 80, 82, 176) of the railway car bogie (20); an inclined surface (74, 128, 200) of the vibration damper wedge having a primary angle alpha (α) and a secondary angle beta (β), and the specified inclined surface of the vibration damper wedge has a curvature, and the specified curvature has an operating point (WP); wherein said spring seat has an axial center line (CL ₁₇₆ ); wherein said vibration damper wedge (64, 66, 68, 70, 120) has a reference plane (190) which is perpendicular to said non-metallic wear surface and which contains said axial center line (CL ₁₇₆ ); wherein said axial center line (CL ₁₇₆ ) intersects said inclined surface (74, 128, 200) of the vibration damper wedge at a point (DP) of intersection, and said operating point (WP) is located in the central region of said inclined surface (74, 128, 200 ) vibration damper wedge adjacent to the specified reference plane (190). 2. Wedge (64, 66, 68, 70, 120) of the vibration damper for the bogie (20) of a railway car, designed to interact with the pocket (60, 62) of the bolster, wherein said wedge of the vibration damper contains a friction element (90, 160 ) which, when in use, movably interacts with the wear surface (92) of the column (36) of the side frame of the bogie (20) of the railway car, and the specified friction element has a non-metallic wear surface; a spring socket (172, 174) which, when in use, interacts with a spring (76, 78, 80, 82, 176) of the railway car bogie (20); an inclined surface (74, 128, 200) of the vibration damper wedge, wherein said inclined surface of the vibration damper wedge has a primary angle alpha (α), a secondary angle beta (β) and curvature, wherein said curvature has an operating point (WP) at which said the inclined surface (74, 128, 200) interacts with the pocket (60, 62) of the bolster at rest; wherein said spring seat (172, 174) has a center line (CL176); wherein said vibration damper wedge (64, 66, 68, 70, 120) has a reference plane (190) which is perpendicular to said non-metallic wear surface and which contains said axial center line (CL176); wherein said axial center line (CL ₁₇₆ ) intersects said inclined surface (74, 128, 200) of the vibration damper wedge at the intersection point (DP); wherein said vibration damper wedge has a first end surface (132) and a second end surface (134), wherein said first and second end surfaces are spaced apart from and opposite each other; wherein said reference plane (190) is located midway between said first and second end surfaces (132, 134), and said operating point (WP) is located in a central region adjacent to said reference plane (190). 3. A vibration damper wedge (64, 66, 68, 70, 120) for a railway car bogie (20), having a friction element (90, 160) having a friction surface that, when in use, interacts with a wear plate of a side frame column (36) a railway car bogie (20), said friction surface being a non-metallic wear surface; a spring seat (172, 174) which, in use, cooperates with a spring (76, 78, 80, 82, 176), said spring seat having an axial center line (CL ₁₇₆ ); and an inclined surface (74, 128, 200), which in use interacts with a corresponding pocket surface (60, 62) of the bolster (24) of the railway car bogie (20); a reference plane (190) that is perpendicular to said friction surface and parallel to said axial center line (CL ₁₇₆ ) of said spring seat (172, 174); wherein said center line intersects the inclined surface (74, 128, 200) at the intersection point; wherein said inclined surface (74, 128, 200) has a spherical arc and a radius of curvature; wherein said inclined surface (74, 128, 200) has a primary angle (α) of the vibration damper wedge and a transverse secondary angle (β) of the vibration damper wedge; wherein said inclined surface (74, 128, 200) has a working point (WP), which in an equilibrium state interacts with said bolster pocket (60, 62) at the point of rolling contact; and wherein said radius of curvature of said inclined surface (74, 128, 200) has an origin lying on one side of said reference plane (190), a radius passing through said origin, and said operating point (WP) diverging from said reference plane under a bevel angle, wherein said bevel angle is said secondary angle (β) of the vibration damper wedge when said radius is considered in a plane oriented at said primary angle (α) of the vibration damper wedge. 4. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 3, wherein said operating point (WP) is located downhill from said intersection point (DP). 5. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 4, wherein said curvature has a radius of curvature in said reference plane (190) of less than 76 cm. 6. A vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 4, having first and second end surfaces (132, 134), wherein said reference plane (190) is defined midway between said end surfaces , and the specified vibration damper wedge is asymmetrical. 7. Wedge (64, 66, 68, 70, 120) vibration damper according to any one of claims 1 to 6, in which the specified operating point (WP) is offset from the specified non-metallic wear surface of the specified friction element at a distance of 3 to 16 mm further than the specified center line (CL ₁₇₆ ). 8. Wedge (64, 66, 68, 70, 120) vibration damper according to any one of claims 1 to 7, in which the specified operating point (WP) is offset from the specified non-metallic wear surface of the specified friction element at a distance of 6 to 19 mm further than the specified center line (CL ₁₇₆ ). 9. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 8, wherein said non-metallic wear surface is offset from said axial center line (CL ₁₇₆ ) by a first distance x ₁ ; the specified operating point (WP) is offset from the specified center line by a second distance x ₂ ; and the ratio x1:x ₂ is in one of the ranges (a) from 21:2 to 21:8; and (b) from 10:3 to 40:3. 10. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 8, wherein said non-metallic wear surface is offset from said axial center line (CL ₁₇₆ ) by a first distance x1; the specified operating point (WP) is offset from the specified center line by a second distance x ₂ ; and the ratio x1: _x2 is 5:1. 11. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 10, wherein said non-metallic wear surface has an overall height y1 and said operating point (WP) is in the range of ^3/8 y1 to ⁵ /8 y1 along the height of the specified non-metallic wear surface. 12. The vibration damper wedge according to any one of claims 1 to 11, wherein said center line (CL ₁₇₆ ) of said spring seat (172) is in said reference plane (190) and said operating point (WP) is also in said reference plane . 13. Wedge (64, 66, 68, 70, 120) of the vibration damper according to any one of claims 1 to 12, in which the specified radius of curvature is in the range of 51 ± 2.5 cm. 14. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 13, wherein said center line intersects said inclined surface at a reference point (DP), said reference point (DP) is at said reference plane (190), and the specified operating point (WP) is located less than 2.5 cm from the specified reference point (DP). 15. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 14, wherein said curvature is a composite curvature. 16. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 15, wherein said curvature of said inclined surface (74, 128, 200) of the vibration damper wedge is spherical. 17. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 16, wherein said curvature has a radius of curvature of less than 89 cm. 18. Wedge (64, 66, 68, 70, 120) of the vibration damper according to claim 17, in which the specified radius is in the range from 38 to 76 cm. 19. Wedge (64, 66, 68, 70, 120) of the vibration damper according to claim 17, in which the specified radius is 50 ± 1.3 cm. 20. Wedge (64, 66, 68, 70, 120) vibration damper according to any one of claims 1 to 19, in which the alpha angle is in the range from 30 to 50°. 21. Wedge (64, 66, 68, 70, 120) of the vibration damper according to claim 20, in which the beta angle is in the range from 5 to 20°. 22. Wedge (64, 66, 68, 70, 120) vibration damper according to any one of claims 1-21, in which the specified - 23 045784 the inclined surface (74, 128, 200) of the vibration damper wedge has a contact patch having a radius of less than 2.5 cm. 23. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 22, having first and second end walls, wherein said inclined surface of the vibration damper wedge is located between said first and second end walls, and said reference plane is located midway between said first and second end walls. 24. The vibration damper wedge (64, 66, 68, 70, 120) of claim 23, wherein said inclined surface of the vibration damper wedge extends to said first and second end walls. 25. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 24, including an internal partition (150) extending between said inclined surface (74, 128, 200) of the vibration damper wedge and said friction element (90, 160), and said operating point (WP) is aligned with said internal partition (150). 26. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 25, wherein said operating point (WP) is located along said reference plane (190). 27. A vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 26, which is at least partially hollow. 28. The vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 27, wherein said spring seat (172) includes a downward protruding protrusion (174) having a size such that it can be aligned location inside the spring (76, 78, 80, 82) of the bogie (20) of the railway car. 29. Wedge (64, 66, 68, 70, 120) of the vibration damper according to any one of claims 1 to 28, configured to be assembled with a spring group of a railway car bogie, wherein the spring group has a first vibration damper spring installed to interact with said spring seat for said vibration damper wedge, and at least a first main spring, wherein said first vibration damper spring has a greater free height than said first main spring. 30. A vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1 to 29, wherein said spring seat (172) is defined in a lower wall (170) of said vibration damper wedge, wherein said lower wall has a width and rear-rounded corners, said corners having a radius that is greater than 0.25 of the width of said bottom wall of the vibration damper wedge, and said inclined surface (74, 128, 200) of the vibration damper wedge ends at an outer edge of at least one of the indicated rounded corners. 31. Wedge (64, 66, 68, 70, 120) of the vibration damper according to any one of claims 1 to 30, configured to form a set of four said vibration damper wedges for one end of the bolster (24) of the bogie (20) of a railway car , wherein said four vibration damper wedges include two said left-hand beta damper wedges (64, 68) and two said right-hand beta damper wedges (66, 70). 32. Wedge (64, 66, 68, 70, 120) of the vibration damper according to any one of claims 1 to 30, configured to be assembled with a pocket (60, 62) of the bolster. 33. Trolley (20) of a railway car, having a combination of a vibration damper wedge (64, 66, 68, 70, 120) according to any one of claims 1-32 and at least one self-steering device (42, 44). 34. The trolley (20) of a railway car according to claim 33, in which the self-steering device (42, 44) includes a rocker arm (42) located between the axle box support (40) of the side frame and the bearing (46) of the wheelset axle. 35. The trolley (20) of a railway car according to claim 34, having a side frame (26) installed with the possibility of swinging in the lateral direction, while the specified side frame has a lateral rigidity k _pendulum of swing, and the specified trolley (20) of the railway car has a bolster beam (24) mounted on spring groups (54, 56), and said spring groups have a lateral stiffness kspringshear, ^while kpendulum is less than kspringshear. -