Self-contained frictionally damped resilient suspension system for railcars

A railcar shock insulated suspension system is provided with unique variable rod springs and co-related variable rate friction damping means. Each rod spring includes an elongated solid body of elastomeric material which, when subjected to radially directed compression loads and allowed to deform freely without rotation, produces a non-linear or variable spring rate proportional to the ratio of the areas of its loaded to unloaded surfaces. Rod springs constitute the main suspension spring elements, and additionally may constitute the damping means spring elements, if desired.

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:

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 "shape factor" 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' 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.