Self-contained frictionally damped resilient suspension system for railcars
This is a continuation of application Ser. No. 637,684, filed Dec. 4, 1975, which is a continuation-in-part of application Ser. No. 457,425 and application Ser. No. 457,426, both filed on Apr. 3, 1974, and both now abandoned. 
    
    
     BACKGROUND OF THE INVENTION 
     This invention relates to suspension systems for railcars and, more particularly, to suspension systems of this type which utilize elastomeric compression springs. 
     Most conventional suspension systems employ some sort of main spring element for supporting the sprung structure (e.g., the body of a vehicle in a vehicular suspension). In these suspension systems, conventional coil suspension springs are utilized most commonly as the main spring element; however, these springs deflect linearly in response to the application of compressive loads, or produce a linear load-deflection spring curve, and therefore are selected on the basis of the average anticipated suspension load. Above and below the average load, the spring provides insufficient and excessive spring force, respectively. In addition, conventional coil compression springs tend to bottom under high load or shock conditions. Typically, the compression limits of such springs are about one third their original length and, upon application of compressive loads or shock forces of sufficiently high magnitude, are compressed to the point that their coils bottom or come into metal to metal contact with one another. Coil springs, therefore, do not offer the most effective main sprin suspension means in many applications, especially railcar suspensions, subject to a wide range of load conditions, high magnitude shock forces or large displacements. Another disadvantage of coil springs is that they have little or no inherent damping, a property which is highly desirable in many vehicular suspension systems. 
     Elastomeric springs of the shear, or combination shear-compression type, have been used extensively in vehicular suspensions, particularly railcar suspensions; however, fears of environmental effects, cold flow, creep, fatigue and other failure phenomena have limited the usage of elastomeric springs to suspensions subject to small magnitude displacements or loads, or have required supplemental coil springs or load distribution devices to prevent over stressing of the elastomeric springs. Some known elastomeric compression springs, for example, employ end plates which are bonded to the opposed force bearing surfaces of the spring in order to increase the load bearing capacity of the spring. Destructive stress concentrations, however, tend to develop adjacent the end plate bonds under high compressive loads and may lead to bond breakage or rupture of the body of elastomeric material. Such bonded end plate elastomeric springs are disclosed in U.S. Pat. No. 3,461,816, issued to Beck and U.S. Pat. No. 2,154,586, issued to Stern. Elastomeric shear compression or sandwich springs, which typically employ flat inter-leaved rubber spring elements in V or chevron formations, suffer from similar problems. 
     Other types of known elastomeric compression springs employ solid bodies of elastomeric material of square or rectangular cross sectional configuration. In many practical applications, however, these springs tend to develop undersirable stress concentrations at the sharp corners between adjacent spring surfaces. Such stress concentrations may lead to rupture of the spring material under high compressive loads. Toroidal elastomeric springs of generally circular cross sectional configuration also have been proposed, as in U.S. Pat. No. 3,515,382, issued to Gallagher. Toroidal springs, however, are highly undersirable in many applications, especially vehicular suspension systems, in which the spring elements must be mounted and operated in a confined space. Furthermore, toroidal springs tend to develop destructive hoop stresses upon application of high compressive loads. 
     Another known type of elastomeric spring utilizes an elastic solid roller body, formed of natural or artificial rubber. The roller body rolls about its longitudinal axis between two spaced apart load application surfaces as they are moved relatively in rotational or translational fashion. The spring force obtained is produced as the roller body is rolled between the load application surfaces, in response to relative movement thereof, to a region of reduced spacing, where it is compressed radially. Such roller type elastomeric springs are disclosed in U.S. Pat. Nos. 2,712,742, 2,729,442, 2,819,063, 2,842,410, British Pat. No. 749,131 and German Pat. No. 2,189,897, all issued to Neidhart, and in U.S. Pat. No. 2,189,870, issued to Sluyter. The roller body or bodies employed in these roller type elastomeric springs, however, are subject to surface wear and destructive shear stresses which are produced by the rolling action between the load application surfaces. Additionally, destructive force couples, or torsional stresses, are produced as the force application axis, along which the compressive force of each load application surface is applied, shifts to one side or the other of the roller body centroid. 
     Still another known elastomeric spring which is generally similar to the roller type spring described previously except that the spring is compressed radially without rotational movement, is disclosed in U.S. Pat. No. 3,351,308, issued to Hirst. The arcuate load application surfaces between which the spring is squeezed or compressed radially surround and confine the spring, and hence prevent free bluging of diagonally opposed portions of the curved spring side surfaces under all load conditions, especially upon application of high compressive loads which tend to produce substantial flattening of the spring. Consequently, if restrained excessively from free deformation, the spring tends to rupture or fail. To limit total spring deflection, the top casting and base are formed so that they engage one another in response to application of a sufficiently high downward load or impact force. When engaged, however, forces are transmitted directly between the top casting and base, and hence the spring is ineffective. That is, the spring system comprising two load application surfaces and the elastic spring body, in effect, bottoms in much the same manner as a conventional coil spring. The end result is that this spring system is of limited usefulness, and is unsuitable for use in many practical spring applications, such as vehicular suspension systems, in which high load or impact forces are to be encountered. Further, in addition to being of arcuate configuration, the load application surfaces between which the spring is compressed initially are spaced apart diagonally. Consequently, the spring is squeezed therebetween or is deflected radially along a deflection axis which extends in a diagonally inclined direction between these surfaces in response to application of a downward load applied along a vertical load application axis. That is, the deflection axis (diagonal) does not coincide with the load application axis (vertical). The end result is that destructive torsional forces, force couples and shear stresses are produced in the elastic spring body. Futhermore, due to such non-symmetrical spring loading with respect to vertically applied loads, two parallel springs must be used. Another generally similar suspension for railcar useage is disclosed in U.S. Pat. No. 1,484,954, issued to Masury. 
     In addition to the main spring elements, another important factor in the design and selection of a suspension system is the damping or shock absorbing means. There is much concern in the mobile vehicle industry, particularly in the railcar and truck-trailer fields, regarding problems caused by dynamic forces which produce high frequency vibration, resonant motion, etc. Prior art damping systems which attempt to eliminate or minimize these problems have largely utilized hydraulic shock absorbers or constant force friction elements. Hydraulic damping is velocity responsive (rather than load responsive) and hence tends to produce damage to the lading for higher frequency forcing modes. Most prior friction dampers for railcar usage employ coil springs for actuating the friction shoe (see U.S. Pat. No. 3,517,620, issued to Weber) and thus produce a linear rate damping force. Examples of additional similar prior art dampers are disclosed in U.S. Pat. Nos. 3,338,183, 3,486,465 and 3,545,385. Consequently, such prior friction dampers suffer from many or all of the disadvantages of coil springs mentioned above and may even require supplemental damping by hydraulic shock absorbers. 
     One type of prior railcar suspension including a friction damper, which employs a rubber suspension pad for urging a friction shoe into frictional engagement with a friction surface, is disclosed in U.S. Pat. Nos. 2,356,743, and 2,357,264, both issued to Light, and in U.S. Pat. No. 2,295,554, issued to Cotrell. This suspension, however, due to the inclusion of coil springs as the main suspension spring elements, suffers from some or all of the above-mentioned disadvantages. 
     SUMMARY OF THE INVENTION 
     The invention utilizes one or more unique rod springs as the main suspension spring elements and preferably combines them with variable rate friction damping means to provide a railcar suspension system which achieves matched springing and damping of controlled variable rate. The preferred variable rate damping means also include a rod spring. 
     The rod spring of this invention provides a non-linear or variable rate, load-deflection spring curve which may be controlled selectively. The rod spring comprises an elongated solid body of elastomeric material, preferably natural rubber or its equivalent, having diametrically opposed longitudinal load bearing surfaces and having a cross sectional configuration selected to provide a shape factor which varies in response to application of compressive loads. Shape factor is defined as the ratio of the area of the spring load bearing surfaces to the area of its non-load bearing surfaces which are free to bulge in response to an applied compressive load. The higher the shape factor, the greater the amount of load required to produce a certain spring deflection, and hence the steeper the spring curve obtained. That is, the spring becomes stiffer, or its resistance to further compression increases porportionally as its shape factor increases. Thus, it is possible, by selecting a cross sectional configuration which provides certain shape factors at various compressive loads, to control the spring curve obtained in order to obtain a variable rate spring curve of desired spring rate or stiffness. 
     The rod spring of this invention is positioned between, but preferably is not bonded to, two substantially parallel load application surfaces of generally planar configuration. The load application surfaces are so positioned that a compressive force can be applied to the spring in a radial direction along a deflection axis substantially perpendicular to the longitudinal axis of the body. That is, the load application axis, which is projected between and is normal to the faces of the load application surfaces, substantially coincides with the deflection axis. When compressed radially between such load applications surfaces, curved portions of the spring side and end surfaces adjacent the load application surfaces roll down freely into interfacing static contact therewith, thus increasing the area of the spring load bearing surfaces and increasing the spring shape factor. Consequently, the rod spring may be squeezed or compressed radially while its load bearing surfaces remain engaged in interfacing static fashion with their respective load application surfaces. Thus, shifting, sliding or rotational movement of the spring relative to the load application surfaces, and hence surface wear, destructive force couples or shear stresses are eliminated or minimized while achieving substantial deflections not possible with prior elastomeric compression springs. 
     Thus, it will be appreciated from the foregoing summary that this invention provides a highly effective railcar suspension system which offers matched spring and damping forces of controlled rates, the spring or damping rates being controllable most effectively by selection of an appropriate rod spring shape factor. It is possible, of course, to utilize one or more rod springs as the main spring elements in combination with conventional dampers, for example hydraulic dampers. Furthermore, the rod spring of this invention for the first time provides an elastomeric spring capable of withstanding substantial compressions -- from 30% to 50%, or higher, being practical -- without suffering compression set, cold flow, etc. Consequently, the suspension system of this invention is highly versatile and may be employed successfully in many railcar suspensions subject to large displacement between the sprung and unsprung structures which heretofore were ill-suited for usage with elastomeric springs. 
     These and other features, objects, and advantages of this invention will become apparent in the detailed description and claims to follow taken in conjunction with the accompanying drawings in which: 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a longitudinal vertical section taken through a suspension system embodying the principals of the invention as applied to a railcar truck; 
     FIG. 2 is a transverse vertical section of the FIG. 1 suspension; 
     FIGS. 3, 4 and 5 are fragmentary vertical sections of modified forms of the FIG. 1 suspension; 
     FIG. 3A is a fragmentary side elevation of a portion of the side frame and adapter for the FIG. 1 railcar truck suspension; 
     FIG. 6 is a side elevation of a double axis railcar truck equipped with the suspension system of FIG. 1; 
     FIG. 7 is a plan of the railcar truck of FIG. 6; 
     FIG. 8 is a perspective of the rod spring of this invention; 
     FIG. 9 is a transverse section, depicting the rod spring of FIG. 8 in its unloaded condition; 
     FIG. 10 is generally similar to FIG. 9, depicting the rod spring of FIG. 8 in its loaded condition; 
     FIG. 11 is a fragmentary longitudinal section depicting the rod spring of FIG. 8 in broken and solid lines in its unloaded and loaded conditions, respectively; 
     FIG. 12 is a graph of a load versus deflection of the rod spring of FIG. 8 formed of various cross sectional configurations; 
     FIGS. 13 and 14 are generally similar to FIG. 9, depicting a modified cross sectional configuration and various loading orientations of a modified rod spring. 
    
    
     DETAILED DESCRIPTION OF THE DRAWINGS 
     The suspension system of this invention is illustrated and described in the environment of a double axle railcar truck; however, other applications and uses of the suspension system of this invention will be apparent to one of ordinary skill in the art. 
     The illustrated truck of FIGS. 6 and 7 includes two spaced apart parallel side frames 40, each mounted at its ends by two suspension units, respectively. One such suspension unit is depicted in broken lines in FIG. 6 and in vertical cross section in FIGS. 1 and 2. A transverse bolster 48 (FIGS. 6 and 7) having an I-beam 49 and flanges 49a joins the opposite side frames in a torstionally compliant manner. Side bearings 42 support the car body 7 on the side frames and a pivot pin 27 transfers horizontal loads between the car body and the transverse bolster through a compliant center spring 44. 
     The suspension unit of FIGS. 1 and 2 includes a plurality of rod springs 84, 85, and 86 as the main suspension spring elements, and one or more friction dampers which preferably include rod springs 96, all contained in housing 82. 
     Referring now in particular to FIGS. 8-14, each illustrated rod spring is made up of an elongated solid body of elastomeric material, preferably natural rubber or its equivalent, having a generally cylindrical cross sectional outline. Diametrically opposed upper and lower flats 3 and 4 extend in parallel planes substantially the entire length of the spring until merging with opposed end faces 6 and 8. To minimize or eliminate end stress concentrations, the peripheral edges of the end faces are rounded, as shown (FIG. 8). One or more generally cylindrical bosses 10 are upstanding from the upper and lower flats (lower bosses not shown). These bosses serve to locate and restrain rotational movement of the rod spring with respect to opposed load application surfaces between which the rod spring may be compressed, in a manner to be described presently. To prevent end stress concentrations, the bosses are spaced inwardly from the ends of the rod spring. 
     When loaded, the rod spring preferably is squeezed between parallel upper and lower load application surfaces 12 and 14, of generally planar configuration, as depicted in FIGS. 9-11. The faces of the upper and lower load application surfaces are parallel and respectively interface with, but are not bonded to, the upper and lower spring flats 3 and 4. In the examples of FIGS. 9-11, the load application surfaces are so positioned that when they are subjected to opposed vertical forces, indicated by arrows in FIGS. 9 and 10, directed along a force application axis, the rod spring is squeezed or compressed radially therebetween, without rotation, along a deflection axis 16 substantially perpendicular to the body longitudinal axis 18 and through its centroid C. (In the example of FIGS. 8-11, the centroid C coincides with the longitudinal axis 18). It will be understood, of course, that the direction of an applied load is not axial, that is, the rod spring is not loaded and thus does not function as an endloaded column. Furthermore, the rod spring is compressively loaded and hence does not function as a shear spring, nor as a combined shear-compression spring. 
     By virtue of the configuration, disposition and positioning of the load application surfaces of FIGS. 9-11, the interfacing portions of the load application surfaces 12 and 14 and spring flats 3 and 4 remain in continuous static interfacing engagement. That is, they are free of relative rotational, shifting or sliding movement during load application. Thus, as the rod spring is squeezed between the load application surfaces 12 and 14 in the manner described, its centroid C is moved along a line which substantially coincides at all times with the deflection axis 16. Further, the force vectors of the opposing compressive forces, indicated by arrows in FIGS. 9 and 10, substantially coincide at all times. Consequently, torsional forces or force couples which tend to twist of rotate the rod spring are eliminated or minimized, as are destructive shear stresses and surface wear of the elastomeric material which comprises the body. 
     In the unloaded condition, depicted in FIG. 9, the spring flats are disposed at spaced apart locations on the deflection axis 16 and constitute the spring load bearing surfaces; however, in the loaded condition, depicted in FIGS. 10 and 11, the spring flats are disposed at decreased spacing on the axis 16 and the spring load bearing surfaces are of increased area. This increased areas is provided by roll down of the upper and lower portions of the spring side and end surfaces into interfacing static contact with their respectively adjacent load application surfaces. More specifically, as illustrated in cross section in FIGS. 10 and 11, the rod spring includes four arcuate peripheral segments 20, respectively located adjacent the longitudinal edges of the upper and lower flats 3 and 4, which roll down laterally in response to an applied load. In their unloaded condition (FIG. 9), these segments approach and merge with their respectively associated flats at accute angles; however, when loaded (FIG. 10), they roll down freely until substantially coninciding with the faces of their respectively associated flats and thereby enter into interfacing static engagement with the load application surfaces 12 and 14. As depicted in FIG. 8, rounded portions of the spring ends roll down simultaneously with and in generally similar manner as the segments 20. Consequently, due to the free roll down of these spring side and end surfaces in response to an applied load, the resultant increased contact area obtained is free of destructive stress concentrations which commonly occur in bonded plate elastomeric springs in the vicinity of the rubber-plate interface or bond, or in conventional elastomeric shear compression springs. Additionally, the increased contact area obtained is free of corners, indentations or the like which tend to produce other destructive stress concentrations. Inasmuch as these portions of the spring side and end surfaces do not shift or slide relative to the load application surfaces during roll down, and thereafter remain in static interfacing engagement therewith, problems of surface wear, shear stress, force couples and torsional forces, such as are encountered in conventional roller-type and other prior elastomeric springs, are overcome or mitigated. 
     The cross sectional configuration of the rod spring provides a shape factor which varies in response to application of compressive loads along the deflection axis 16. The term &#34;shape factor&#34; as used herein may be defined as the ratio of the area of the spring load bearing surfaces to the area of the unloaded surfaces which are free to bulge in response to an applied load. In the unloaded condition, depicted in FIG. 9, the upper and lower flats constitue the load bearing surfaces, and hence the shape factor is small. When loaded, as depicted in FIGS. 9 and 10, portions of the spring side and end surfaces roll down to increase the area of the load bearing surfaces, as described previously. consequently, the shape factor is correspondingly increased. As the shape factor increases, increasing compressive loads are required to attain a given deflection. That is, the spring becomes stiffer, or its resistance to further compression increases proportionally as its shape factor increases. Thus, it is possible, by forming the rod spring of a cross sectional configuration which provides a load variable shape factor, to obtain a rod spring having a non-linear or variable rate load-deflection curve. The preferred cross sectional configuration is symmetrical relative to the load application axis and most preferably is generally cylindrical. 
     FIG. 12 depicts typical load-deflection spring curves for the rod spring of this invention formed of various cross sectional configurations. These curves may be determined experimentally, or by calculation using known formulae for compression of rubber spring bodies. Curve (a) depicts the load-deflection curve of the rod spring of substantially cylindrical cross sectional configuration generally similar to that indicated by broken lines 22 in FIG. 9. The lower generally horizontal portion of the curve represents a low shape factor produced by substantially line contact between the spring load bearing and load application surfaces. The steeper, generally vertical portion of the curve depicts a higher shape factor in which the sides and ends of the rod spring have rolled down to thereby increase the load bearing contact surfaces thereof. Curve (b) represents the load-deflection curve of a rod spring formed with diametrically opposed upper and lower flats such as those shown in FIG. 8. These flats serve to preload or stiffen the spring during initial deflections, and hence the lower portion of the curve (b) is offset from the corresponding portion of curve (a). As will be appreciated from FIG. 12, curve (b) is somewhat steeper than curve (a) for light loads; however, the two curves (a) and (b) substantially coincide at higher loads. Curves (c) and (d) depict the load-deflection curves of rod springs of generally eliptical cross sectional configuration which are loaded such that the force application axis substantially coincides with their minor and major axes, respectively, as illustrated in FIGS. 13 and 14. Inasmuch as the rod spring of FIG. 13 has a higher shape factor than that of the rod spring of FIG. 14 for most applied loads in the lower to mid-range, the lower portion of curve (c) is steeper than that portion of curve (d); however, at higher loads in which each spring is squeezed further, curves (c) and (d) are of generally similar steepness. 
     Thus, it is possible, by varying the cross sectional configuration of the rod spring to control the load-deflection spring curve obtained. It will be recognized that other cross sectional configurations may be used and that the configurations illustrated and described herein are not considered as limiting. It is possible, of course, to control the spring curve obtained by modifying the area and/or configuration of the flats 3 and 4. Additionally, the spring curve obtained may be controlled by using elastomeric material of varying hardness. 
     The rod spring of this invention may be compressed as described herein to obtain substantial deflections, without overstressing, or appreciable compression set, or cold flow of the elastomeric material used, under both intermittant and repeated loading conditions. Preferably, for a rod spring having the flattened cross sectional configuration illustrated, percentage deflection, or the percentage ratio of deflection (x) to diameter (d) (See FIGS. 10 and 9 respectively), is about 30% to 50%; however, a generally similar rod spring has been tested successfully at substantially higher percentage deflections. 
     To reduce the size and weight of the rod spring without affecting the spring curve obtained, parallel diametrically opposed relief surfaces 24 may be formed by lateral relief cuts in the body of elastomeric material between the upper and lower flats, as shown (FIGS. 8 and 9). The depth of these relief cuts, and hence the amount of rubber removed, is dependent upon the amount of roll down of the spring sides. These cuts must leave the arcuate segments 20 of sufficient length S (FIG. 9) to roll down under maximum load and deflection. As illustrated in FIGS. 8 and 11, the ends of the rod spring also are relieved to reduce the size and weight and to enable the rod spring to be positioned in a confined space. The ends, of course, may be of generally spherical configuration, as depicted in broken lines 26 in FIG. 11, to further minimize or eliminate end stresses. 
     Although the illustrated rod spring is made up of a solid body of elastomeric material, it could include appropriate cavities, holes, etc. to aid in fabrication, as well as to provide air circulation for heat dissipation. The construction and arrangement of such cavities, holes, etc., of course, should be selected to maintain the above-mentioned operation of the rod spring. 
     The rod spring of this invention is particularly suitable for use in vehicular suspension systems for railcars or rubber tired vehicles. In such applications, several springs may be arranged to combine their spring forces in series (i.e., between a plurality of load application surfaces; e.g. see FIG. 1 -- three tires designated 84, 85 and 86) or in parallel (i.e., between common load application surfaces; e.g. see FIG. 1 -- intermediate tier designated 85) with their longitudinal axes parallel, perpendicular or askew to one another. Thus, it will be understood that the suspension system spring curve obtained may be controlled by controlling the spring curves of the individual rod springs, as described previously, and additionally by controlling their numbers, orientations, arrangements, etc. 
     Referring again to FIG. 1, centrally positioned in the housing is a wheel axle 2 having a conventional roller bearing 5 and a wheel 1. Within the housing is a vertical stack of spring elements designated as 84, 85, and 86 each generally similar to the rod spring of FIGS. 8-11. The top longitudinal element 84 abuts against housing 82 and rests on a horizontal spacer plate 87. The upper and lower surfaces of the spring element 84 are provided with flats 84a. The second tier of spring elements 85 is also provided with flats 85a at top and bottom and includes three such elements, one centered over the axle 2 and the others spaced equidistantly on either side of the axle. The spring elements abut against the spacer plate 87 and a second lower spacer plate 88. Finally, the lowermost spring elements 86 are two in number transversely located and equidistantly spaced on either side of the axle. The elements 86 are also provided with upper and lower flats 86a and abut against the lower spacer plate 88 and a spacer plate 90. The spacer plate 90 is bonded to an elastomer pad 91 which mainly provides lateral shear spring return motion for the truck and railcar between the wheels and sideframe. The spring pad 91 is carried on a sub housing carrier or adapter 93 which is mounted on the bearing 5. A bumper or stop 94 may be provided for limiting downward movement of the housing 82 to prevent over-compression of the spring elements 86. It will be understood, of course, that the suspension housing 82 and plates 87, 88 and 90 provide sets of opposed substantially flat load application surfaces which correspond to surfaces 12 and 14 of FIGS. 8-14. 
     The springs 86 (lowest tier) may be designed to be the stiffest springs with the intermediate and upper springs 85 and 84 being softer. With this construction, the suspension may be precompressed to the initial limit of the three inch travel allowed within the coupler vertical limits (preferably precompressed about 3 inches) and thus provides good spring effect even in the unloaded condition of the vehicle. The relative sizes and shapes of the rod springs, of course, may vary depending on the application and the shape factor selected. In a preferred embodiment of the invention shown in FIGS. 3 and 4, the intermediate springs 85 are replaced by a single rod spring 100 which is arranged in parallel alignment with spring 84. Each of the springs 84 and 100 has a generally cylindrical shape and again is softer than the lower springs 86. These lower springs, although stiffer than the upper two layers, provide, to a large extent, the soft portion of the spring rate in the suspension system. Compression with light lading in the car will allow the springs 86 to absorb some of the loads, but for most loads and all heavy lading, the upper spring elements come into play. It should be understood that the size, stiffness, number and arrangement of the springs is dependent upon the load capacity and type of application required, therefore the examples are to be considered as illustrative only. 
     Although there is inherently in the rod springs a certain amount of internal damping, dependent upon the elastomeric material used, it is insufficient to provide total damping necessary for most conditions experienced by vehicles. As an example, a rubber having approximately 15% internal damping in the embodiment illustrated may amount to less than 8% of critical (effective) damping for the total system for an application which may require as much as 20% of critical damping. The additional required damping is obtained from a unique variable rate damping system again best illustrated in FIGS. 1 and 2 of the drawings. 
     The variable damping system of this invention produces a damping force which varies with vertical distance on a varying exponential curve which is designed to closely match that of the main suspension rod springs and add cumulatively to their inherent damping. The varying exponential spring and damping rate provides a slowly changing rate under low loads and a rapidly increasing rate in the upper load range. The system of FIGS. 1 and 2 includes two friction dampers which also provide load responsive lateral damping for the lateral pad springs where used. For this purpose, a horizontal lateral gap of about 1/4 inch is provided on either side of the friction shoe 98 to allow lateral movement. 
     As is also readily apparent, the illustrated symmetric spacing of the dampers balance the damping and springing moments on either side of the axle 2 thus reducing undesirable moments and tending to center the carrier at all times. 
     In the damping system of FIG. 1, each damper includes a rod spring 96 arranged in a cylinder 95, the two cylinders 95 being spaced equidistantly from and on each side of the rotatable axis (in FIG. 1, the vehicle axle). A friction shoe element 98, of a conventional brake shoe type material, is pushed by a rectangular piston 99, which abuts against the side of spring element 96, against a sloped surface 46, which latter is parallel to the face of element 98. It can thus be readily recognized that lowering the housing 82 due to increased loads or dynamic forces acting on the car will move the sloped surface 46 downwardly, sliding it past the friction show 98 and compressing the rod spring 96 as described above. The greater the downward movement, the greater the outward force provided by the spring element 96 on the friction shoe thus increasing the damping force applied through the system. It is, of course, understood that the angle of the sloped surface and/or the stiffness of the rod spring 96 can be varied to control the damping force obtained, and that carrier 93 and piston 99 provide load application surfaces generally similar to surfaces 12 and 14 of FIGS. 8-14. 
     Vertical guidance of the carrier 93 to prevent fore and aft movement of the wheels relative to the side frame is provided first by the friction shoes 98 and under more severe loading by the replacement steel wear plates 97 or their mountings 97&#39; secured to fore and aft surfaces of the adapter for engagment in the axle slot 99 of the side frame 40. 
     Preferably, the damping force provided by the FIG. 1 system is directly related to the spring force provided by the spring elements 84, 100, 86 (FIGS. 3 and 4) such that both spring and damping forces increase at approximately the same rate under loading dynamic or static. Thus, where loads are high, the spring force is high and the damping force is high which is a desirable situation. When the loads are light, the spring force is low and the damping force is low, thus cushioning the shock to fragile light lading such as fruit in a refrigerator car. 
     In some instances, it is desirable to provide additional stability to the cylindrical spring elements, particularly when the generally cylindrical elements shown in FIGS. 1 and 2 are not stacked in vertical alignment with one another. For this purpose the modifications shown in FIGS. 3 and 4 are provided with bosses 100a that locate in recesses 100b of the plates 87 and 88. The bosses seated in the recesses provide locating and stability functions as well as allow ease of assembly and movement of the stacked springs to or from the housing 82. The bosses are spaced in from the longitudinal ends of the plates to minimize stress concentration at the ends of the elastomers. 
     In FIG. 5 still additional spring elements 101 of a different arrangement are shown. In this embodiment, the flats are provided with recesses 101a. The shape of the spring element is a balance between the amount of rubber removed at top and bottom to provide economy of material plus proficiency of performance. The recess type spring (still generally cylindrical) has basically the same volume as the more generally cylindrical spring. The shape of the curve of this modified form in cross sectional quadrant is approximately the same as for 84 and 100. The shape in all the modified forms is quite important since too great a change from the generally cylindrical configuration changes the spring characteristics too severely to achieve the desired results. 
     While the forms illustrated provide the best configurations, they are not to be considered as limiting. Furthermore, various flats can be added to assure a better bond (if required) to the spacer plates and the ends of the generally cylindrical elements can be relieved to prevent high stress concentrations provided they are not changed too drastically to preclude their meeting the requirements of shape factor, size, spring travel, and desired variable spring rates necessary to a fully operative suspension system. 
     In addition, and as best shown in FIGS. 3-5, the configuration or arrangement of the spring elements in the spring housing can also be varied. For example, in FIGS. 1 and 2, three cylindrical spring elements 85 are arranged with their longitudinal axes the same as the longitudinal axis of the axle 2. 
     While preferred embodiments of the invention have been illustrated and described, it should be understood that variations will be apparent to one skilled in the art. Accordingly, the invention is not to be limited to the specific embodiments illustrated.