JPS6234581B2 - - Google Patents

Description

[Detailed description of the invention]

The present invention relates to a vehicle suspension system, a diesel car equipped with the vehicle suspension system, and a vehicle with rubber tires. Although the present invention is described and illustrated herein with reference to a particular diesel rail car and rubber-tired vehicle suspension system, the invention is not limited to the particular suspension systems illustrated and described, but is otherwise Can be used for equipment with vehicle suspension systems. Vehicle suspension system with rubber tires disclosed in U.S. Pat. No. 3,984,125 and U.S. Pat. No. 3,961,582
The diesel vehicle suspension disclosed in No. 1 and No. 3,961,584 has multi-rod springs of generally circular cross section as the spring elements of the main load support and the friction damper. A primary object of the present invention is to provide a vehicle suspension system with improved rod springs and friction dampers. Another object of the present invention is to provide a vehicle suspension system that is lighter in weight than suspension systems that utilize circular cross-section rod springs or conventional elastomeric compression springs for a given load and frequency condition. Another object of the invention is to provide a vehicle suspension of the above type having a single rod spring as the spring element for both the main load support and the friction damper. Yet another object of the invention is to provide a vehicle suspension of the above type with a rod spring that produces a lower suspension frequency with a given amount of elastomeric material compared to conventionally used rod springs of circular cross section or ordinary elastomeric compression springs. The purpose is to provide equipment. Another object of the invention is to provide a vehicle suspension system of the above type that produces a substantially constant frequency over a predetermined load range. Yet another object of the invention is to provide a vehicle suspension system of the type described above which allows automatic steering when passing through curves. Another object of the present invention is to provide an automatically steered railcar truck having a dual axle (biaxle), single axle or articulated single axle construction. Yet another object of the invention is a vehicle of the above type having a friction damper, preferably of a variable type, which automatically adjusts to wear so as to maintain substantially the same rate over the service life of the wear material. An object of the present invention is to provide a suspension device. The present invention provides a spring element and the spring element such that a first component of the load supporting force is applied to the spring element as a load bearing compression force, and a second component of the load supporting force is applied to the friction damper and controls the friction damper. The present invention relates to a vehicle suspension system having a friction damper operatively associated with a spring element. According to a preferred embodiment of the invention, in which the spring element is a rod spring having an oval cross-section, the rod spring and the friction damper and the load transmission operator can simultaneously apply forces thereto. The operator is supported by a load-supporting support surface, and the first component of the load-supporting force applied to the support surface is applied to the rod spring along a fixed position force vector as a load-supporting compressive force, and at the same time the load-supporting force is is configured and arranged with respect to the support surface such that the second component of is applied as a normal force to the friction damper along the normal force vector. In most practical systems, the upper rod spring load bearing surface is supported by a suitable load application surface formed by the vehicle body or by a suitable load application member attached to the vehicle body. The operator and support surface are arranged in a downwardly disposed load-bearing vertical alignment with respect to the rod spring, with the operator interposed between the rod spring and the support surface even if other alignments are utilized. The supporting surface is
by a load arm constituting a load-bearing connection between the vehicle body and the running gear, by a member attached by the vehicle axle, or by other suitable means according to the alignment of the rod spring operator and damper. It is formed by a device. A preferred operator forms upper and lower contact surfaces that are tapered in profile and diverge along a force vector associated with the component of the damper normal force applied thereto. The support surface slopes parallel to the lower contact surface. The upper operator contact surface is parallel to and contacts one of the rod spring loaded support surfaces, and the lower contact surface slidably engages the support surface. As a result, a shift or lateral movement of the support surface in response to an applied load-bearing force occurs only as a sliding movement with respect to the lower contact surface and is therefore not transmitted to the upper contact surface. Additionally, the operator is coupled to the friction damper such that the operator maintains a substantially fixed position with respect to the friction damper during force application. The upper operator contact surface thus remains in static contact with the rod spring during force application. A preferred friction damper is made of two relatively movable parallel friction surfaces in a substantially perpendicular relationship with respect to a vector, said normal force component along which vector being applied by an operator. One friction surface is movably connected in cooperation with an operator such that in response to the application of a normal force component, said friction surface is pressed against the other friction surface to produce a variable friction damping force. It is therefore clear that the present invention provides a vehicle suspension system that is economically less expensive, more reliable, and more flexible than anything previously known in the art of vehicle suspension systems. For example, for use in a rubber-tired vehicle tandem axle suspension, the vehicle suspension of the present invention has two pairs of independently movable intermediately pivoted load arms, each pair of load arms It is associated with a single rod spring and friction damper of the type described herein. When used on this and other rubber-tired vehicles, the suspension system of the present invention equalizes loads and reduces individual wheel shock by about 50%, controlling vibration through balanced load variable damping. to provide the desired mechanical properties, reduce shock loads by providing variable load damping, improve roll resistance, extend service life in the use of rod springs, and provide automatic steering when negotiating curves. performance, thus reducing timer wear due to scuffing occurring at cornering speeds above equilibrium speed and reducing rolling resistance drag, stable under acceleration and braking conditions without wheel bounce, The effective unsprung weight can be lowered,
Furthermore, a substantially constant frequency can be provided over most practical load ranges. For use in railcars, the suspension system of the present invention has many of these and other advantages.
Suspension systems that utilize independently movable load arms and their associated rod springs and friction dampers, for example, allow four-wheel trucks or articulated trucks to maintain a relatively low angle of attack relative to a curved track and avoid It can follow rail curves when traveling at equilibrium speed. The truck can automatically steer around spiral rail curves, other curves on which diesel cars can roll, or siding or yard curves. For use on railcars, the suspension system of the present invention provides additional cushioning and absorption of shock loads in close proximity to the rails, thus providing protection for the railcar structure, supported loads and brake rigging and couplings. A very favorable spring action for auxiliary parts such as cushioning. When used in rubber-tired vehicles and diesel railcars, the rod springs of the present invention with an oval cross section provide lower suspension vibrations at a given load condition at a lower weight than could conventionally be used with rod springs with a circular cross section or ordinary elastomeric compression springs. This is preferable because the number can be maintained. Additionally, the spring can provide a substantially constant suspension frequency over a predetermined load range. In the most practical use of suspension systems in connection with rubber-tired vehicles and diesel vehicles, they are also virtually indestructible, do not bottom out, bounce non-linearly, and are therefore capable of handling from light loads to heavy or equivalent loads. It is possible to provide a load bearing that is extremely effective and highly flexible in response to a wide range of load conditions, including shock loads. As a result, damaging vertical accelerations are reduced or substantially minimized and loads can be best maintained under adverse road or track conditions and at high speeds. Of course, it is clear that rod springs of other cross-sectional shapes or other types of springs can also be used in place of the rod springs of oval cross-section. The variable friction damper according to the invention can automatically adjust for the wear during the service life of the friction elements, thus eliminating the efficiency losses associated with the wear of the damping friction surfaces that occur in hydraulic dampers or conventional friction dampers. No performance degradation occurs. These and other objects, features, and advantages of the present invention will be apparent from the following detailed description and claims, which refer to the drawings in which like numerals refer to like parts. Referring initially to FIGS. 1-3, a vehicle suspension system according to the present invention includes a rod spring 10, a friction damper (indicated generally by the numeral 12), and a pivot 1 disposed below the rod spring 10 and a pivot 1.
It has an operator 14 connected to the friction damper 12 by 5. Pivotable load arm 16
supports the operator 14 and the rod spring 10. Operator 14 simultaneously applies and controls the rod spring and damper with their respective forces derived from a common load-bearing force provided by arm 16. According to the embodiment shown in FIGS. 1 and 2,
Suspension systems are used on fixed single-axle diesel cars. In this embodiment, the arm 16 is connected at its inner end to the rail car body 20 by a low friction plastic push pivot 18. Arm 1
The outer end of 6 supports one end of the railcar axle 22 by a bearing adapter 23 (FIG. 2) and a bearing keeper 25 located below. An elastomeric suspension pad 24 (FIG. 1) is interposed between adapter 23 and arm 16 and transmits cushion loads between adapter 23 and arm 16. A rod spring, friction damper, and operator associated with the same independently movable load arm are operatively associated with the body 20 and the other end (not shown) of the axle 22 and between the body 20 and the axle 22. A load-bearing support is provided. Referring to FIGS. 1 and 2, operator 1
4 forms an upper contact surface 26 and a lower contact surface 28 which have a tapered outer shape and diverge relative to each other in the direction of the damper 12 (see FIGS. 1 and 2). The upper contact surface 26 is supported against a lower rod spring load support surface in load-transferring relationship therewith, to prevent movement during application of a load, as explained below. The lower contact surface 28 is supported against and movable with respect to an underlying inclined support surface 30 formed by the outer end of the load arm. Upper contact surface 2
6 is parallel to and opposite the upper load application surface 32, which surface 32 is formed by the vehicle body in the example and is supported conformally against the upper rod spring-loaded support surface. As explained below, the rod spring between surfaces 26 and 32 is laterally compressed during loading. In the embodiment, the friction damper is constituted by two relatively movable parallel friction surfaces 34, 38. The friction surface 34 is shown in FIGS.
As shown, it is formed by a conventional friction pad attached by a friction shoe 36 pivoted to the larger end of the operator 14 by a pin or pivot 15. The friction surface 38 is formed by a member projecting downwardly from the vehicle body. Friction surfaces 34, 38 are arranged with respect to operator 14 so that they are engaged with each other in an unloaded condition (see FIG. 2). Friction surfaces 34, 38
Friction surfaces 34, 38 create a variable friction damping force when pressed together by normal force applied by operator 14 during movement of arm 16 with respect to body 20. In an embodiment, the friction surface 38 is sloped with respect to the trajectory of movement of the arm 16 such that the friction damping force produced during downward movement of the load arm exceeds the friction damping force produced during upward movement of the load arm, as described below. . As shown schematically in FIG.
can simultaneously provide the vertical component of the force applied to the rod spring and damper by the load arm. Referring to Figure 3, the load supporting force is the contact surface 2
8 is applied by the operator 14 via the planes 28, 30 by the load arm 16 along an inclination vector 40 perpendicular to the plane of 8. Operator 14 applies a first component of the load arm force to the rod spring through surface 26 along vector 42 which appears at the lower spring load bearing surface essentially as a load bearing compressive force only. Operator 14 similarly applies a second component of the load arm force along vertical vector 44 to damper 12 via pivot 15. The force of the load arm occurs as a normal force pressing the friction surfaces 34 and 38 together. The distance that the friction surfaces 34, 38 must be moved relative to each other to obtain the desired amount of friction damping is sufficiently small (preferably negligible) that the operator 14 is in static contact with the friction surfaces. The friction surfaces 34, 38 are arranged so that they do not move laterally with respect to the rod spring during actuation to apply a force over a variable range. As the material comprising the friction surfaces 34 or 38 wears, the operator will, of course, assume a relatively close clearance position relative to the friction surface; however, this type of operator movement is most practical. This occurs gradually so that in some cases the operator has no effect of being in contact with the rod spring. Therefore, the friction damper is sized to self-regulate its wear over its service life. In order for the suspension to maintain the loading conditions in FIG.
can move or move parallel to 8. To this end, a layer of low friction material 46 is interposed between surfaces 28 and 30. As a result, when the load arm pivots with respect to the body, the position of the operator 14 with respect to the rod spring is not changed in a plane coincident with surface 26, so that the lower load-bearing surface in opposing contact with that surface 26 of the rod spring remains static. is in contact with. In an embodiment, the friction surface 38 is inclined with respect to the locus of movement of the load arm such that the frictional damping force obtained during downward movement of the load arm exceeds the frictional dampening force obtained during upward movement. In particular, friction surface 3
8 is inclined at an acute angle with respect to the trajectory of the arcuate movement of the load arm indicated by the arrow in FIG. As a result, as load arm 16 moves upwardly from the position of FIG. 2 toward the position of FIG. 1, the force applied along vector 44 of FIG. It occurs as a normal force that gradually decreases in a direction that is tilted away from the trajectory of motion. Conversely, as the load arm moves downwardly from the position of FIG. 1 toward the position of FIG. 2, the friction surface 38 moves closer to the locus of motion of the load arm and hence of the operator, thus increasing the available normal force. The suspension system of the present invention is shown under two loading conditions in each of FIGS. 1 and 2. In FIG. 1, the suspension system is considered to be at its heaviest load, and in FIG. 2, it is considered to be at a light load. In Figures 1 and 2, the rod spring is deflected to approximately 40-45% and 15-20% of its free upright height, respectively, and in both cases the unloaded sidewall of the rod spring is entirely Matches the circular outline.
Under impact loads, the rod spring of FIGS. 1-2 can deflect as much as 55-60% of its free upright height, although it cannot undergo relatively large deformations without damage. Comparing and explaining FIGS. 1 and 2, the area of each load-bearing surface of the rod spring increases as the applied load-bearing compressive force increases.
This increase in load bearing surface area is achieved as portions of the spring side and end surfaces are pressed into contact with the illustrated opposed surfaces 26, 32, respectively. As a result, the shape factor of the rod spring increases with progressively increasing loads. The term "shape factor" is determined as the ratio of the area of the spring loaded bearing surface to the area of the free expanding unloaded surface in response to an applied load.
As the shape factor increases, increasing compressive loads are required to obtain a given deflection. That is, the spring becomes relatively stronger, or its resistance to compression increases proportionately as the shape factor increases. By forming the rod spring in a cross-sectional shape with a variable load shape factor, it is possible to obtain a rod spring with a non-linear or variable load deflection curve of the type shown in FIG. 17th
Illustration shows 22.9 cm (9 inch) long axis, 16.6 cm (6.55 inch)
Figure 2 shows a load vs. deflection curve for an example rod spring having a minor axis of 16.6 cm (6.55 inches) and a length of 16.6 cm (6.55 inches). The present invention is of particular interest in that the oval rod spring cross section provides this type of deflection curve with a very favorable suspension frequency and reduced weight, as will be shown below. As explained above, the opposing surfaces of surfaces 26, 32 remain parallel to and in contact with the upper and lower rod spring load bearing surfaces, respectively. In the embodiment of FIGS. 1 and 2, the rod spring is substantially perpendicular to the longitudinal axis of the body when surfaces 26, 32 are subjected to opposing vertical forces exerted by the vehicle body and load arm, respectively. The surfaces 26, 32 are arranged such that they are laterally compressed between the surfaces 26 and 32 without rotational movement along a fixed position vector that coincides with the longitudinal axis. Of course, the direction of the applied load is not axial, the rod spring is unloaded and does not act as an end load column, and the rod spring that is compressively loaded does not act as a shear spring or a combined shear spring. It is understood that it does not act as a compression spring. As a result, these contact surfaces undergo no relative rotational, moving or sliding movement during the application of a load, thus causing the rod spring to move in its direction when it is compressed in the manner shown and described in the text. Torsional forces, shear forces or couples rotating parallel to and about the longitudinal axis are eliminated or minimized, thereby eliminating or eliminating surface wear and destructive shear stresses on the elastomeric material constituting the rod spring. It can be reduced considerably. The rod spring of the present invention is shown in more detail in FIGS. 12-16. Rod springs are made of a solid body of elastomeric material, preferably natural rubber. Upper and lower load bearing surfaces (number 48, 5
0) are preferably convex and curved about the longitudinal axis of the body. Similarly, the ends 52, 54 are convex and merge at the sides and upper and lower load bearing surfaces of the body along curved corners as shown in FIGS. 14 and 15. A single upper boss 56 projects from the upper load bearing surface and two lower bosses 58 project from the lower load bearing surface. The bosses are of the type shown in FIG. 1 in order to prevent or substantially minimize the tendency of the bosses to move or rotate relative to the rod spring while under compressive loading and to facilitate positioning of the rod spring during assembly. It has a suitable profile for matching and engaging suitable recesses formed in surfaces 26,32. The curved shape of the load-bearing surface allows the adjacent side walls to be progressively compressed during compressive loading, resulting in lower stress in the elastomeric material during loading operations, thus significantly increasing the service life. be able to. Furthermore, the rounded load bearing surface allows the rod spring to remain stable during increasing deflection, thus allowing for a relatively large circular to circular ratio (major axis/minor axis). The ellipse ratio and length selected for the rod springs of FIGS. 13-16 are in accordance with the desired suspension frequency, operating load range, weight, and service life. Generally, the higher the ellipse ratio utilized, the lower the resulting suspension frequency for a given load. The use of harder elastomeric materials will increase the frequency of the suspension; however, hysteresis effects will occur due to the higher hardness of the elastomeric materials.
Rod springs installed in suspension systems therefore generate high heat after repeated load cycles. As a result, the use of high hardness elastomeric materials significantly shortens the service life of rod springs. By increasing the length of the rod spring, a higher load bearing capacity is obtained, but the spring frequency is increased somewhat by increasing the resistance against the unloaded end face. In most practical embodiments, the rod spring preferably has an ellipse ratio of about 2.5:1 or less;
A range of 1.1:1 to about 2.2:1 is most preferred. The length of the rod spring is at least equal to the length of the minor axis;
Its hardness is about 40°~70° International Rubber Hardness (IRHD)
is within the range of The six rod springs are constructed from 50° hardness natural rubber according to the following specific example, but not limitation.

[Table] As understood from the above examples, the ellipticity ratio is an important factor in the weight of the obtained rod spring, and a rod spring with a relatively high ellipticity ratio is lighter in weight than a rod spring with a relatively low ellipticity ratio. Conversely, as the ellipticity ratio approaches 1:1, which corresponds to a circular cross section, the weight of the rod spring to obtain a given frequency and load condition increases considerably. 4, 5 and 23, the suspension system of the present invention is shown in use on an articulated railcar truck. Figures 4, 5 and 23
The suspension shown is generally equivalent to the suspension shown in FIGS. 1 and 2, which suspension has two identical suspension units 59, each having an independent movable load. arms, rod springs and associated friction dampers, but differ in that each load arm acts between a respective truck side frame and axle. Identical parts of the suspension systems of FIGS. 4, 5 and 23 are designated by the same numbers with the suffix "A". The illustrated articulated diesel railcar truck has two spaced-apart configurations whose corresponding ends are connected to each other by a transverse member 61 having a generally V-shape with a concave center in the side profile as shown in FIG. It has parallel lateral frames 60. The illustrated truck supports a vehicle body bolster 62 that mounts a center sill 63 connected to a suitable vehicle body. Side supports 64, each attached by two side frames, have vertical load-bearing supports for the ends of bolster 62. The horizontal load-bearing support for the bolster 62 and rotating vehicle body-track connection comprises an elastomeric spring ring 65 and a diagonal member 6 attached by the bolster 62.
9. Provided by a track center post 67 attached by 9. Diagonal member 69 transmits lateral and brake loads from transverse member 61 to column 67.
For the operation of articulated railcars, connector 7
1 is attached by a transverse member 61 adjacent to the concave central portion and associated with the illustrated truck and a second vehicle body (not shown) to prevent rotation of the truck frame with respect to each other when traversing a curved track. It extends between the track (not shown). Sill end 73 mounts a drawbar or coupler 55 at pivot 57 (FIGS. 4 and 5). The drawbar 55 acts between an adjacent end of the illustrated vehicle body and an adjacent end of a second vehicle body (not shown) and is capable of transmitting impact and traction or push and pull forces therebetween. can. When coupled, the two trucks operate as an autopilot dual axle truck. Its autopilot characteristics are further explained below. Side supports and spring rings suitable for use in the illustrated rail car are disclosed in U.S. Pat. No. 3,961,582, which disclosure is hereby incorporated by reference. In the illustrated embodiment, the track pivot axes through the spring ring 65 are longitudinally offset with respect to the drawbar pivot 57 so that the tracks are longitudinally offset with respect to each other as the railcar traverses the tortuous track by toggling action. have a tendency to move to In that embodiment, the connector 71 is capable of allowing relative track movement and, if desired, supplementary relative movement of the tracks about a longitudinal axis as disclosed in U.S. Pat. No. 3,961,582. Can tolerate rotational movement. In other applications, the drawbar pivot axis is placed in perpendicular alignment with the track pivot axis and the connector 71 is longitudinally rigid and can still accommodate relative track rotational movement if desired. An alternative embodiment of the truck of Figure 4 is shown in Figures 25-27, in which identical parts are designated by identical numbers followed by the suffix "AA". . The truck shown in FIGS. 25 to 27 includes a side frame 60AA and members 61AA, 69.
It is generally identical to the track of FIG. 4, except that AA is of a different shape. Furthermore, the 25th
The tracks in Figures to Figure 27 are connected to connector 71AA.
pivoting latch 73 and central sill 63 to prevent rotational movement of the truck when not engaged
It has a member 74 fixed to AA. As shown in FIG. 26, latch 73 is positioned in the locus of motion of the telescoping element forming connector 71AA so that it is held in a retracted position deflected from member 74. The spring (not shown) is the connector 71
Latch 73 is pressed into engagement with member 74 when AA is disengaged. 6, 7 and 24, the suspension system of the present invention is shown in use on a dual axle track. Figures 6, 7 and 24
The suspension shown is generally similar to that shown in FIGS. 1 and 2, the suspension having four suspension units 66, each having an independently movable load arm; They are made of rod springs and associated friction dampers, but differ in that each load arm acts between a respective truck side frame and axle. Figures 6 and 7
Identical parts of the suspension system of FIGS. and 24 are designated by identical numbers with the suffix "B". The illustrated dual axle truck has two spaced apart sections which are generally T-shaped in side view and generally similar to the side frame 60AA of the truck of FIG. 26, as shown in FIG. parallel lateral frame 7
5. The outer ends of each side frame are of reduced thickness and supported by rod springs disposed beneath each, and the intermediate portions are of increased thickness. Two load arms 16B are pivoted to this intermediate section at two adjacent pivots 18B as shown (FIG. 7). In one embodiment, as shown in FIG. 24, a transverse member 77 having a generally V-shaped concave central portion is bolted to the side frames adjacent the intermediate portion. The members 77 are sufficiently torsionally flexible to allow the lateral frames 75 to swing relative to each other about the longitudinal axes of the members 77 in respective vertical planes. The side supports 81, each mounted by two side frames, have vertical load-bearing supports for the ends of the body bolsters 83.
Bolsters 83 and horizontal load-bearing supports for rotating vehicle body-track connections are bolted to elastomeric spring rings 85 and bolsters 83 attached by members 77, as shown in FIG. It is formed by the track center column 87. The bolted connections between the side frames and the members 77 and between the struts 87 and the bolsters 83 are preferably formed with self-locking nuts or upsetting with high vibration resistance.
Lateral supports and spring rings suitable for use with the illustrated track are disclosed in U.S. Pat. No. 3,961,584, the disclosure of which is incorporated herein by reference. Referring to FIG. 28, the load arms of the dual axle truck of FIG. 7 are mounted by a common pivot 19 as shown in FIG. It is shown with . Pivot 93 is side frame 75BB
The pivot 93 is fixed relative to the frame, and the pivot 93 is made movable relative to the frame by mounting the pivot 93 in a vertical slot formed in the frame. See slot 174, described below with reference to FIG. 8 and 9, the suspension system of the present invention is shown for use on a single axle rubber tire vehicle. The suspension system of Figures 8 and 9 is generally similar to the suspension system described and illustrated above, and identical parts are designated by the same number followed by the suffix "C".
The only difference is that the suspension has a separate member 71 forming the inclined support surface 30C. In the embodiments of FIGS. 8 and 9, the vehicle:
2 terminating in parallel spaced apart portions 78, 80 in overlapping relationship with an eye or hitch 72 and an axle 82 used to couple a conventional trailer hitch;
It is constituted by a trailer converter vehicle having two branched frame members 74,76. The bolster 83 has portions 78 and 80 as shown in FIG.
Installed by. Axle 82 mounts member 71 and connects member 7 by means of a parallelogram linkage.
4.76 is supported. As shown in FIGS. 8 and 9, this linkage includes portions 78,
Two coangular upper control rods 84, 86 converge from a pivot mount on 80 to a pivot mount near the center of axle 82 and similarly mounted but extending parallel to portions 78, 80 and extending parallel to portions 78, 80.
It is formed from two co-angular lower control rods 88,90 located below 8,80. 10 and 11, a tandem axle rubber tire vehicle suspension system is shown. This suspension system has the following features except that a rod spring is disposed between opposing and relatively movable parts of two common pivoted load arms.
It is generally similar to the suspension system shown in FIGS. 1-9. Two sets of load arms (only one set shown)
supports the ends of tandem axles 94,96. A single rod spring constitutes a single load bearing spring element for each load arm. Only one set of load arms and their associated rod spring dampers and other suspension elements are shown and described herein, as well as the other set of load arms and their respective associated suspension elements. I would like you to understand that. In particular, with reference to the suspension of FIGS. 10 and 11, the suspension includes an outer or leading load arm 98.
and an inner or trailing load arm 100. The arms are commonly pivoted in the middle by a common central pushing 102. Pushing 102
is attached by a saddle 104 that projects downward from a vehicle frame 106 and is fixed. The upper portion of the saddle 104 includes a torque rod 10 associated with the axles 94, 96, as is conventional in suspension systems.
8 to upright axle bowl brackets 110, respectively. The outer ends of the load arms terminate in respective rubber pushings 112, 114 which are connected to the axle by axle brackets 116, 118 as shown. The suspension system of FIGS. 10 and 11 has a rod spring 120 of elliptical cross-section that is generally similar to the rod springs described and illustrated above. As shown, the upper load bearing surface of the rod spring is supported against a surface formed by overlying portion 122 of arm 100. The lower rod spring load bearing surface is a movable damper bracket 1.
24 are supported on opposing surfaces 123 formed by. The lower surface of the damper bracket is sloped and conformably slidably engages the sloped surface formed by arm 98 and is generally connected to operator 14 as described and illustrated above with respect to FIGS. 1 and 2. is movable relative to the arm during compressive loading of the rod spring in a similar manner. The thicker end of the damper bracket supports a damper pad 126 made of a suitable friction shoe material. This pad is supported on one side of a damper element 128 having a tubular open end attached by the distal end of arm portion 122 as shown in FIG. One side of the damper element has pads 126 in sliding relationship such that arms 98 and 100 each pivot relative to each other.
forming a substantially planar surface 130 (see FIG. 12) that engages oppositely. The other side of the damper element has two additional damper pads 13 which are generally V-shaped in cross-section (see FIG. 10) and are attached by arms 100 as shown in FIG.
6 and 138, respectively. The damper bracket 124 is connected to the operator 1 in FIGS. 1 and 2.
4, pressing pads 126, 136 and 138 into frictional engagement with surfaces 130, 132 and 134, respectively, when subjected to the loads produced by arms 98 and 100. An alternative embodiment of the suspension system of FIG. 11 is shown in FIG. 29, with like parts designated by like numbers followed by a suffix "D". The suspension system of FIG. 11 has two load arms 140, 142 whose outer ends are connected to axle brackets 116D and 118.
D to axles 94D and 96D, respectively. The inner ends of arms 140, 142 are pivoted to respective ends of counterbalance beam 148.
It is pivoted at 44,146. This beam 14
8 is pivotally mounted in the middle of the arm pivot at position 150 by saddle 104D. Rod springs 150, 152 are supported by operators 154, 156, which are supported by arms 140, 142, respectively, as in the suspension system of FIG. Similarly, operators 154, 156
2,164 to press the friction shoes 158,160. A generally similar suspension system made of two load arms and a counterbalance beam is utilized in the trucks of FIGS. 6 and 7. A second alternative embodiment of the suspension system of FIG.
0 and which correspond to parts in FIG. 11 or FIG. 29 are designated by the same number followed by the suffix "DD". single load arm 1
70 is an axle bracket 116DD, 11 at one end.
8DD to axles 94DD and 96DD and supports rod springs 150DD and 152DD. A pin 172 that fits into a vertical slot 174 formed by saddle 104DD is attached to arm 170.
is supported midway along its length. As a result, the arm can slide vertically along the slot 174 relative to the saddle 104DD to distribute the load to the frame rail 176. Referring to FIGS. 18 and 19, FIGS.
Figures 7, 10 to 12, and 23 to 3
The suspension shown in Figure 0 is capable of self-steering when flexing by independent movement of the load arms in response to the action of centrifugal force, as will be explained. Although the suspension systems shown in FIGS. 4-7 and 23-29 can generally achieve autopilot when rounding curves, the suspension systems shown in FIGS. 18 and 1
FIG. 9 shows the automatic steering of the tandem axle suspension of FIGS. 10-12. The single axle suspension systems of FIGS. 1, 2, 8 and 9 can achieve autopilot when mounted at both ends of a vehicle. Further referring to FIGS. 18 and 19, the first
As the rubber-tired vehicle equipped with the suspension system of FIGS. 0 to 12 travels in the direction indicated by the arrow in FIG. There is a tendency to tilt the body of an elevated vehicle outward so that the As a result, the load-bearing forces imparted to the suspension elements associated with the outer wheels tend to increase, while at the same time the load-bearing forces imparted to the suspension elements associated with the inner wheels tend to decrease. Therefore, the first
As shown in Figure 9, the outer load arms tend to move away from the outer ends of the axles, while the inner load arms tend to attract the inner ends of the axles together.
As a result, the axle assumes its respective radial position with respect to the curve when the vehicle is traveling in a straight line until the centrifugal force load condition mentioned above is removed. Referring to FIG. 20, the suspension system of the present invention exhibits a substantially constant frequency over a range of applied loads. Although FIG. 20 shows the spring frequencies obtained under typical loading conditions for the tandem axle suspension shown in FIGS. 10-12, the remaining suspensions generally have similar frequency curves. can be given. Additionally, FIG. 20 shows frequency curves for two prior art tandem axle suspensions. As can be seen, the frequency curve obtained with the suspension of the present invention provides a substantially constant spring frequency between the load limits shown. 21 and 22 illustrate the operating characteristics of the tandem axle suspension of FIGS. 10-12, as well as the operating characteristics of the remaining suspensions described and illustrated in the text. Furthermore, FIGS. 21 and 22 show corresponding operating characteristics of a prior art moving beam suspension with leaf springs. Referring to FIG. 21, the suspension system of the present invention includes:
5.08cm (2 inches) high and 15cm long at 50m.ph
It exhibits relatively less tire elastic force (2 vs. 3) and lower tracking acceleration than prior art suspensions in response to an impact with a (6 inch) hump. 22nd
Referring to the figures, the suspension of the present invention exhibits less deflection and faster damping than prior art suspensions in response to similar impact conditions. While several preferred embodiments of the invention have been shown and described, it will be obvious that various modifications will occur to those skilled in the art. For example, the suspension of FIG. 11 is adapted to use two rod springs and two friction dampers by providing two cross arms, each arm connecting the rod spring and the operator to an arm pivot point. A lower part of the same outer diameter and a damper are attached to the lower part of the arm 98 for supporting it on one side of the arm 98, and an upper part of the same outer diameter is attached to the portion 122 for supporting the rod spring on the other side of the arm pivot point. Therefore, the invention is not limited to the specific embodiments described and illustrated herein, but the true scope and spirit of the invention should be determined with reference to the following claims.

[Brief explanation of the drawing]

1 is a side view of a fixed single-axle diesel vehicle with a suspension according to the invention in a loaded condition; FIG. 2 is a side view of the suspension of FIG. 1 in a lightly loaded condition; Generally equal side view,
3 is a force diagram showing the force transmission state occurring in the suspension system of FIG. 1, FIG. 4 is a top view of an articulated diesel railcar truck having the suspension system of FIG. 1, and FIG. Figure 6 is a cross-sectional view taken along line 5-5 of Figure 1.
7 is a cross-sectional view at 7-7 in FIG. 6, and FIG. 8 is a single vehicle with rubber tires having the suspension system shown in FIG. 1. 9 is a side view of the vehicle body shown in FIG. 8; FIG. 10 is a top view of a vehicle with tandem axles and rubber tires having the suspension system shown in FIG. 1; FIG. 11 is a line shown in FIG. 10. 12 is a sectional view taken along line 12-12 in FIG. 11, FIG. 13 is a perspective view of an elliptical cross-section rod spring according to the present invention, and FIG. 14 is a sectional view taken along line 12-12 in FIG. 15 is a side view of the rod spring in FIG. 13, FIG. 16 is an end view of the rod spring in FIG. 13, FIG. 17 is a graph of load versus deflection of the rod spring in FIG. 13, and FIG. The figures are Figures 10 to 1.
A schematic diagram showing centrifugal force load conditions when turning a curve with the suspension system shown in Figure 2, and Figure 19 is a schematic diagram showing the automatic steering of the suspension system shown in Figures 10 to 12 caused by the load conditions shown in Figure 18. FIG. 20 is a schematic diagram showing two prior art tandem axle suspensions and FIGS.
Graphs of static frequency versus load for the suspension system of Figure 12, Figure 21 for the prior art suspension system and Figures 10-
Graph of acceleration versus time with the suspension system of FIG. 12,
Figure 22 shows the prior art suspension system and Figures 10-1.
Graph of travel amount versus time with the suspension system in Figure 2, 2nd
3 is an end view of the articulated diesel railcar taken along line 23-23 in FIG. 4, and FIG. 24 is an end view taken along line 24-2 in FIG. 7.
Sectional view of the two-axle diesel railcar truck in No. 4, No. 25
The figure shows the fourth track showing two tracks connected to each other.
Top view showing another embodiment of the truck shown in Figure 26.
25 is a cross-sectional view taken along line 26--26 in FIG. 25, FIG. 27 is a cross-sectional view taken along line 27-27 in FIG. 26, and FIG. 28 is an alternative embodiment of the track shown in FIG. Side view generally similar to the truck of Figure 7;
29 is an alternative embodiment of the rubber-tired vehicle of FIG. 11, a side view generally similar to the vehicle of FIG. 11, and FIG. 30 is an alternative embodiment of the rubber-tired vehicle of FIG. 12 is a side view generally similar to the vehicle of FIG. 11; FIG. 10... Rod spring, 16... Load arm, 2
0...Body, 22...Axle, 34, 38...
friction surface.

Claims (1)

[Claims] 1. A movable load arm device for supporting an axle with respect to a body, a spring element, a device for forming a buffer friction surface connected to the body, a buffer element, an axle and a body. movably disposed on said load arm arrangement between and compressing said spring element in response to relative movement of said body and said load arm arrangement and simultaneously pushing said damping element to friction it against said damping friction surface; a force-resolving wedge device for supporting the spring element in a load-supporting relationship to the body so as to engage the damping element and supporting the damping element in a friction-generating relation to the damping friction surface. Characteristic vehicle suspension system. 2. The friction surface is configured such that the buffering frictional force generated while the load arm device moves in one direction exceeds the buffering frictional force generated while the load arm device moves in the opposite direction. The vehicle suspension system according to claim 1, wherein the vehicle suspension system is inclined with respect to a moving path of the load arm device. 3. The spring element consists of an elastomeric rod spring of oval cross-section, which rod spring has a longitudinal axis perpendicular to both axes of the ellipse and is only subjected to compressive loads transverse to said longitudinal axis. 2. The vehicle suspension system according to claim 1, wherein the vehicle suspension system is mounted so as to receive a. 4 support means for supporting and transmitting forces to said damping element and tapered operator means operatively associated with said support means for transmitting forces to said support means and said rod spring; one contact surface slidably attached to the movable load arm device and one load supporting surface of the rod spring in load supporting contact between the rod spring and the supporting means; The vehicle suspension system according to claim 3, characterized in that it has two contact surfaces with the other contact surface. 5. A ring inserted between the one contact surface and the support means to enable low-friction relative movement between the one contact surface and the support means during movement of the support arm device. 5. A vehicle suspension system according to claim 4, further comprising means. 6. The load arm device has a single load arm, the load arm being pivotally connected at one end and connected to the axle at the other end, the load arm having the force resolving wedge device adjacent to the other end. 2. A vehicle suspension system according to claim 1, wherein said vehicle suspension system is supported by said vehicle. 7. The vehicle suspension system according to claim 6, further comprising means adjacent to the pivot axis of the arm for allowing the one end to move with respect to the body. 8. The vehicle suspension system according to claim 1, wherein the buffer friction surface is inclined at an acute angle with respect to the curved movement path of the load arm device. 9. The load arm device has a parallelogram link pivoting at one end and connected to an axle at the other end;
2. A vehicle suspension system according to claim 1, wherein said link supports said spring element adjacent to its other end. 10. The load arm device according to claim 1, wherein the load arm device includes an axle support member and a parallelogram link pivotally connected between the body and the axle support member. Vehicle suspension system. 11 The load arm device includes a load arm assembly whose ends are pivotally connected to corresponding ends of two tandem axles, and a body for supporting the assembly intermediate the ends thereof. support means pivotally mounted, said assembly supporting two such spring elements at spaced apart positions on either side of said support means. Vehicle suspension system according to scope 1. 12 the load arm assembly has two arms and the support means has a counterbalance beam;
12. A vehicle suspension system as claimed in claim 11, characterized in that the balance beam is pivotally connected at both ends to both ends of the arm and to the body midway between the ends. 13. The load arm assembly has a single beam member, the support member being pivotally mounted by the body and having an intermediate length such that the beam member is movable perpendicularly to the body. 13. A vehicle suspension system as claimed in claim 12, further comprising means connected to said beam member at. 14 The force-resolving wedge device applies the load-supporting compressive force transmitted from the load arm device to the spring element as a first component, and applies the load-supporting compressive force transmitted from the load arm device as a compressive force as a second component perpendicular to the buffer friction surface to the buffer element. 9. The load arm arrangement according to claim 1, further comprising a tapered operator slidably connected to the load arm arrangement for applying a force. Vehicle suspension system. 15. The vehicle suspension system of claim 14, wherein the spring device includes an elastomeric spring element positioned to receive a compressive load. 16 The load arm arrangement has a counterbalance beam pivotally connected to the body intermediate its ends and two load arms, the load arms being adjacent to their respective first ends. and pivotably connected to opposite ends of the counterbalancing beam, respectively, and adjacent the second ends thereof to corresponding first ends of two tandem axles. A vehicle suspension system according to claim 1, characterized in that the two parts are connected to each other. 17 The load arm device has a load arm slidably connected to the body midway between the ends, the load arm being connected to corresponding first ends of two tandem axles at the ends of the tandem axles. 2. The vehicle suspension system according to claim 1, wherein the vehicle suspension system is pivotably connected adjacent to the section. 18 interconnecting the load arm to the body such that the distance between the first ends of the two tandem axles increases as the vertical load transmitted to the load arm by the body increases; 18. The vehicle suspension system according to claim 17, further comprising means for: