Abstract:
A piston-type rail retarder includes a shock absorber having a cylinder defining a cavity between a pair of end walls. A piston rod extends axially through one of the end walls, and a piston is mounted on the piston rod between the end walls so as to divide the cavity into first and second working chambers containing a damping fluid. A first array of passages extends through the piston and communicates between the first and second chambers, and a sprung valve assembly is provided for obturating the first array of passages in the event of the velocity of the piston rod relative to the cylinder exceeding a predetermined value. The first valve mechanism includes a valve plate and a spring for biasing the valve plate into an open position. An array of restricted orifices extend through the valve plate and communicate directly with the corresponding array of first passages when the valve plate is closed. The restricted orifices limit the flow of fluid through the first array of passages so as to set up a pressure differential between the first and second working chambers and to apply a predetermined damping force. The dimensions of the restricted orifices are determined in accordance with the desired force-displacement profile of the shock absorber.

Description:
BACKGROUND OF THE INVENTION 
     THIS invention relates to a shock absorber, and in particular to a shock absorber for use as a wagon retarder of the type in which movement of a wagon wheel along the rail is regarded by the retarder acting against a peripheral flange of the wheel. 
     In the applicant&#39;s corresponding U.S. Pat. No. 5,730,260, which is incorporated herein by reference in its entirety, a rail retarder-type shock absorber is disclosed which includes a cylinder defining a cavity between a pair of end walls. A piston rod extends axially through one of the end walls, and a piston is mounted on the piston rod between the end walls so as to divide the cavity into first and second working chambers containing a damping fluid. A first array of passages extends through the piston and communicates between the first and second chambers, and a sprung valve assembly is provided for obturating the first array of passages in the event of the velocity of the piston rod relative to the cylinder exceeding a predetermined value. A second array of passages communicates between the working chambers and provides a fluid flow path when the first array of passages is blocked by the sprung valve assembly. Restricted exhaust orifices communicate with the second array of passages for limiting the flow of fluid through the second array of passages so as to control the movement of the piston rod. The dimensions of the exhaust orifices are determined in accordance with the desired force-displacement profile of the shock absorber. 
     In a preferred embodiment, a so-called “relaxable orifice” version of the shock absorber is provided in the form of a moveable orifice plate which is spring biased against the rearmost face of the piston to define the restricted exhaust orifices and to allow them to open in response to an increase in pressure. 
     Whilst this shock absorber has been found to operate effectively, a number of manufacturing steps are associated with the formation of the first and second arrays of passages, as well as the formation and adjustment of the restricted exhaust orifices. 
     In addition, in the “relaxable orifice” version, it has been proved difficult to release sufficient fluid as the orifice plate opens to ensure that the counter-force exerted by the retarder is insufficient to cause bounce and possible derailment in the case of lighter rolling stock moving at relatively high speeds. 
     SUMMARY OF THE INVENTION 
     According to a first aspect of the invention there is provided a shock absorber comprising: 
     a) a cylinder defining a cavity between first and second end walls; 
     b) a piston rod extending axially through one of the end walls; 
     c) a piston assembly carried on the piston rod and mounted to slide axially within the cylinder so as, to divide the cavity into first and second working chambers arranged to contain damping fluid; 
     d) first passage means extending through the piston assembly and communicating between the first and second working chambers; 
     e) first valve means for at least partly closing the first passage means in the event of the axial velocity of the piston relative to the cylinder in a first axial direction exceeding a predetermined value, and 
     f) at least one permanently open restricted orifice being defined in the first valve means and communicating between the first and second working chambers via the first passage means on closure of the first valve means, 
     the restricted orifice being restricted relative to the first passage means so as to limit the flow of damping fluid between the first and second working chambers and to set up a pressure differential therebetween to apply a predetermined damping force, with the dimensions of the restricted orifice being determined in accordance with the desired force-displacement profile of the shock absorber. 
     Preferably, the valve means includes a valve plate and biasing means for biasing the valve plate into an open position, with the restricted orifice means extending through the valve plate and communicating directly with the first passage means when the valve plate is closed. 
     Preferably, the first valve means is located operatively upstream of the first passage means within the first working chamber. 
     The restricted orifice means typically has a diameter of 0-3 mm, and preferably of around 2 mm. 
     In one form of the invention, the shock absorber includes second passage means communicating between the first and second working chambers, second valve means for at least partly obturating the second passage means, and second biasing means for biasing the second valve means into a closed position, the second valve means being arranged to open in response to a predetermined pressure differential existing between the first and second working chambers. 
     Typically, the second valve means includes a second valve plate defining a pressure accumulator well portion which is responsive to pressure exerted by fluid via the second passage means, the second valve plate being arranged to open in response to a predetermined pressure differential existing between the first and second working chambers. 
     Advantageously, the first passage means comprises a plurality of equi-spaced cylindrical first passages extending axially through the piston assembly, the at least one permanently open restricted orifice comprises a plurality of corresponding equi-spaced cylindrical restricted orifices extending axially through the first valve means, and alignment means are provided for aligning the first passages with the restricted orifices. 
     According to a further aspect of the invention there is provided a shock absorber comprising: 
     a) a cylinder defining a cavity between first and second end walls; 
     b) a piston rod extending axially through one of the end walls; 
     c) a piston assembly carried on the piston rod and mounted to slide axially within the cylinder so as to divide the cavity into first and second working chambers arranged to contain damping fluid; 
     d) first passage means extending through the piston assembly and communicating between the first and second working chambers; 
     e) second passage means extending through the piston assembly and communicating in use between the first and second working chambers; 
     f) first valve means for at least partly closing the first passage means in the event of the axial velocity of the piston relative to the cylinder in a first axial direction exceeding a predetermined value; and 
     g) second valve means for at least partly obturating the second passage means, 
     the second valve means including a valve plate defining a pressure accumulator well which is responsive to pressure exerted by fluid via the second passage means, and biasing means for biasing the valve plate into a closed position, the second valve means being arranged to open in response to a predetermined pressure differential existing between the first and second working chambers. 
     Advantageously, at least one restricted orifice is defined in the first valve means and communicates between the first and second working chambers via the first passage means on closure of the first valve means, the restricted orifice being restricted relative to the first passage means so as to limit the flow of damping fluid between the first and second working chambers and set up a pressure differential therebetween to apply a damping force. 
     The second valve means may be in the form of a true valve, in that it is capable of completely obturating the second passage means. 
     Alternatively, the second valve means may alternatively be in the form of an orifice plate which only partly obturates the second passage means so as to define a permanently open restricted exhaust orifice. 
     Typically, the pressure accumulator well includes an annular stepped recess, and the piston assembly includes an operatively rearmost face having a complementally stepped projection which locates within the recess in a sliding fit. 
     The invention extends to a rail retarder including a shock absorber of the type described, as well as to a piston assembly comprising integers a) to f) above. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 shows a partly cross-sectional side view of a rail retarder of the invention fitted to a rail; 
     FIG. 2 shows an exploded partly cross-sectional side view of a first embodiment of a piston assembly of the invention; 
     FIGS. 3A &amp; 3B show assembled partly cross-sectional side views of the piston assembly of FIG. 2 in the respective open and closed positions; 
     FIG. 4A shows a top plan view of the piston assembly of FIGS. 3A and 3B; 
     FIG. 4B shows a top plan view of a piston head forming part of the piston assembly of FIGS. 3A and 3B; 
     FIG. 4C shows an underplan view of the piston head of FIG. 4B; 
     FIG. 5A shows a force-stroke graph illustrating the performance of the rail retarder of FIG. 1; 
     FIG. 5B shows a force-stroke graph illustrating the performance of the rail retarder of FIG. 6; 
     FIG. 6 shows a partly cross-sectional side view of a second embodiment of a rail retarder of the invention; 
     FIG. 7A shows a top plan view of a piston head forming part of the rail retarder of FIG. 6; 
     FIG. 7B shows an underplan view of the piston head of the rail retarder of FIG. 6; 
     FIGS. 8A to  8 C show partly cross-sectional side views of a piston assembly of FIG. 6 in three different modes of operation; 
     FIG. 9 shows an assembled partly cross-sectional side view of a piston assembly forming part of a third embodiment of a rail retarder of the invention in a first idle mode of operation; 
     FIG. 10 shows the piston assembly of FIG. 9 in a second intermediate closed mode of operation; 
     FIG. 11 show the piston assembly of FIG. 9 in a third open mode of operation; 
     FIG. 12 shows a force-stroke graph illustrating the performance of the rail retarder fitted with the piston assembly of FIGS. 9 to  11 ; and 
     FIG. 13 shows two comparative end load-speed graphs illustrating the comparative performances of the rail retarder forming the third embodiment of the invention and a prior art fixed orifice rail retarder. 
    
    
     DESCRIPTION OF EMBODIMENTS 
     Referring first to FIG. 1, a wagon retarder  10  includes a pot housing  11  which is bolted via an integral flange  12  to a rail  13  in a marshalling yard. A cylinder  14  is mounted slidably within the pot housing  11 , and is formed with an upper closed end  16  having a mushroom head-shaped upper percussion surface  18  against which the outer flange  19  of a wagon wheel  20  abuts. 
     An opposite open end  21  of the cylinder is fitted with a gland nut  22  having a central aperture  23  through which a piston rod  24  extends. The lower end of the piston rod  24  is mounted fast against the base of the pot housing  11  by means of a retaining clip  25 . The retaining clip  25  holds the piston rod  24  in position via a cross bar  26  which extends through the lower end of the piston rod. 
     A piston head  28  and a valve disc or plate  30  is bolted to the opposite end of the piston rod so as to form a piston assembly  31  dividing the cylinder cavity into upper and lower working chambers  31 A and  31 B. 
     FIG. 2 shows clearly how the various components making up the piston assembly  31  are stacked onto a narrowed integral stem portion  32  at the upper end of the piston rod  24 . A circlip  34  locates within a lower circumferential indent  36  so as to hold a ring-shaped clack plate  38  in position on a lower projecting platform  40  of the piston  28 . The piston  28  is formed with a central stepped aperture  42  having a lower stepped portion  44  which is arranged to locate snugly over the complimentally stepped or shouldered portion  46  at the base of the piston rod stem  32 . A coil spring  48  locates on a spring seat  50  defined within the aperture  42 , and a tubular spacing bush  52  extends through the coil spring  48  and also locates against the spring seat  50 . The valve disc or plate  30  is formed with a central aperture  54  through which the bush  52  passes. An upper spring seat  56  locates the upper end of the coil spring  48 , and a clamping nut  58  holds the entire piston assembly  31  in position by being screwed down over a threaded portion  32 A of the stem  32 . It can clearly be seen from the detail of FIGS. 3A and 3B how the piston assembly is bolted in position onto the piston rod  24 . Referring also to FIGS. 4A to  4 C, the piston head  28  is formed with a first array of passages in the form of three equi-spaced outer arcuate passageways  60  around which raised platform portions  62  extend, each being formed with uppermost planar sealing faces  64 . The valve disc  30  is formed with a second array of passages in the form of three equi-spaced round cylindrical passageways  66  which extend completely through the valve disc. In this particular example, the passageways  66  have a diameter of 2 mm, in contrast to the passages  60 , which are at least 5 mm wide and 22 mm long. A locating or registering pin  68  extends from a lower planar sealing face  70  of the valve disc, and is arranged to locate within one of the passages  60 . 
     The piston head  28  is fitted with piston rings in the form of upper and lower brass rings  72 A and  72 B and a lower PTFE ring  74  so as to provide an effective seal between the two working chambers  31 A and  31 B. As is clear from FIG. 1, the gland nut  22  is fitted with an annular rubber sealing ring  76  providing a tight seal between the outer surface of the nut and the walls of the cylinder  14 . A glacier bush  78 , a dynamic oil seal  80  and a back-up washer  82  provide an effective dynamic seal between the piston rod  24  and the gland nut  22 . The dynamic oil seal  80  and washer  82  are held in position by means of a sprung clip  84 , with an outer rubber sealing flange  86  completing the dynamic sealing arrangement. A ball valve  88  is provided at the head of a breather passage  90  defined in the pot for limiting the ingress and allowing the escape of rain water and the like from the pot. The condition of the ball valve can be viewed via an inspection port  92 . 
     Both working chambers  31 A and  31 B are charged with hydraulic oil up to a level  94 , with a pocket of nitrogen gas  96  being located above the oil level and being arranged to compress in response to the initial application of force via the wagon wheel  20 . The nitrogen pocket also acts as a pneumatic spring to return the sliding cylinder  14  to its original extended position after it has undergone a downward compression stroke. 
     In response to the application of force via the flange  19  of the wagon wheel against the percussion surface  18 , the cylinder  14  commences a downward stroke. The nitrogen gas is compressed during a first part of the stroke. Thereafter, oil starts flowing from the upper chamber  31 A to the lower chamber  31 B via the passageways  60  in the direction of arrows  98 . The velocity of the oil travelling through the passageways  60 , and in particular at zones  100  just beneath the valve disc  30  will increase as the cylinder accelerates downwards. The increase in velocity of the fluid at the zone  100  will create a reduced pressure zone which will tend to cause the valve disc  30  to close against the biasing force of the spring  48  to the FIG. 3B position after a predetermined velocity has been reached. In this position, the lower sealing face  70  of the valve disc abuts sealingly against the upper sealing face  64  of the platform portions  62 . As a result, the only possible fluid flow paths that now exist between the first and second chambers  31 A and  31 B are via the restricted passageways  66  and the arcuate passageways  60 , as is shown by arrow  101 . The restricted passageways  66  will create a pressure drop in respect of the fluid flowing through these passageways, resulting in an increase pressure differential existing between the first and second chambers, thereby creating a counter-force which effectively retards the downward movement of the piston head and “extracts” the kinetic energy of the, wagon via the wagon wheel  20 , so as to slow it down to a predetermined speed. 
     Once the wagon wheel has passed over the retarder, a forward stroke is initiated by the expansion of the pocket of nitrogen gas  96 . During the return stroke, the spring  48  naturally biases the valve plate into its open FIG. 3A position. In addition, the clack plate  38  is drawn upwardly by the return flow to partly block off the passageways  60 , thereby slowing and controlling the return stroke of the cylinder so as to prevent overshoot. 
     Referring now to FIG. 5A, a typical force-stroke graph is shown for a wagon retarder of the type described in which each restricted passageway has a diameter of 2 mm. In a particular test that was run, the nitrogen gas pressure was set to 700 kPa, and there was an ambient temperature of 20° C., with an oil mass of 178 g. The incoming speed of the wagon was 2.604 ms −1 , and the outgoing speed was 2.488 ms −1 , with the load on each wheel being approximately 45 kN. The resultant energy extracted by the retarder, which is represented by the hatched area  102  under the curve  102 A, was found to be 1242 J. The average retarding force was 15.7 kN, with the maximum force being 35.7 kN. The duration of the power or retarding stroke was found to be 123 ms, and the duration of the return stroke was 149 ms, with the cylinder reaching a maximum downward velocity of 1.285 ms −1 . 
     In FIGS. 6 to  8 C, a second embodiment of a rail retarder of the invention  106  is shown in which identical components are indicated with identical numerals. As can more clearly be seen in FIGS. 7A and 7B, a third array of round cylindrical passages  108  having a diameter of 5 mm extend entirely through the piston  110 . A secondary valve assembly  112  comprising a valve disc  113  and a stack of six plate washers  114  making up a plate washer pack  116  is positioned slidably on a base portion  118  of the stem  120 A of the piston rod  24 . The valve plate  113  is biased by the pack of plate springs  116  to close off the downstream openings of the passageways  108 . 
     In FIG. 8A, the piston assembly is shown in a first idle mode of operation which is effectively identical to that of FIG.  3 A. In FIG. 8B, a second mode of operation essentially identical to that of FIG. 3B is illustrated, in which the restricted passageways  66  communicate directly with the passageways  60 . In FIG. 8C, a third mode of operation is indicated in which the increase in pressure through the third array of passages  108  causes the secondary valve plate  113  to open against the bias of the pack plate springs  116 . Opening of the valve plate or disc  114  will typically occur when the downward force applied to the shock absorber exceeds around 20 kN so as to allow damping fluid to be dumped more rapidly into the working chamber  31 B in the direction of arrow  122 . Naturally the force and rate at which the valve plate  114  opens can be adjusted from less than 20 kN to 35 kN or more by varying the number as well as the resilient properties of the plate springs and the pack of plate springs. 
     In FIG. 5B, a typical force-stroke graph is shown for a wagon retarder of FIG.  6 . The test conditions were substantially identical to those which gave rise to the FIG. 5A graph. The outgoing speed of the wagon was 2.564 ms −1 , with the resultant energy extracted by the retarder being represented by the hatched area  124  beneath the curve  124 A being 1257 J. The average retarding force exerted by the retarder was found to be 15.57 kN, with the maximum force being 36.7 kN. The duration of the power or retarding stroke was found to be 124 ms, and the duration of the return stroke was 258 ms, with the cylinder reaching a maximum downward velocity of 1.308 ms −1 . The efficiency of the retarder of FIG. 6 was found to be 58.7%, in contrast to the 44.1% efficiency figure of the retarder of FIG.  5 A. It can clearly be seen from the graphs that, whilst the maximum retarding force of the retarder of FIG. 6 was considerably less at 26.7 kN, the total energy extracted by this retarder was marginally greater than the energy extracted by the retarder of FIG.  5 A. This was as a result of the curve being “fatter”, in that the energy extraction was spread more evenly over the entire downward stroke of the retarder. One favourable consequence of this is that wheel-lift is reduced, in that the maximum counter-force exerted by the retarder is lower, and does not rise to as sharp a peak as is the case in FIG.  5 A. The second embodiment of the retarder is particularly well suited to the high speed pull-out of wagons at the lower end of the yard where a gentler retarding force is required, especially in respect of lighter wagons, so as to ensure that derailment arising from wheel-lift does not occur. 
     In FIG. 9, those components of the piston assembly which are substantially identical to the piston assembly of FIG. 6 are identified with the same numerals suffixed by an “A”. A piston assembly  130  includes a secondary valve assembly  132  comprising a valve disc  134  and a stack of six plate washers  114 A making up a plate washer pack  116 A which is positioned slidably on a base portion  118 A of a stem  120 A of piston rod  24 A. The valve plate  134  is biased by the pack of plate springs  116 A to close off downstream openings of three equi-spaced passageways, one of which is indicated at  108 A. The passageways  108 A extend through a piston body  110 A which fits over the stem  120 A of the piston rod  24 A, and is held in position by means of a clamping nut  58 A via a tubular spacer bush  52 A. A coil spring  48 A locates on a spring seat  50 A, and biases the valve disc or plate  30 A into an open position. The piston assembly divides the cylinder cavity into upper and lower working chambers  31 A and  31 B. 
     The piston head  110 A is formed with three equi-spaced outer arcuate passageways  60 A around which a raised platform portions  62 A extend, each being formed with uppermost planar sealing faces  64 A. The valve disc  30 A is formed with a second array of passages in the form of three equi-spaced round cylindrical passageways  66 A which extend completely through the valve disc. A locating and registering pin  68 A extends from a lower planar sealing face  70 A of the valve disc, and is arranged to locate within one of the passages  60 A. 
     In the first idle mode of operation the piston assembly operates in a manner which is essentially identical to that of FIGS. 3A and 8A. In FIG. 10, a second intermediate closed mode of operation essentially identical to that of FIGS. 3B and 8B is illustrated, in which the restricted passageways  66 A communicate directly with the passageways  60 A. In FIG. 11, a third open mode of operation is indicated in which the increase in pressure through the third array of passageways  108 A causes the secondary valve plate  134  to be opened against the bias of the pack of plate springs  116 A. Opening of the valve plate or disc  134  will typically occur when the downward force applied to the shock absorber exceeds around 30 kN so as to allow damping fluid to be dumped more rapidly into the working chamber  31 B in the direction of arrow  136 . The pressure at which the valve plate  34  opens can be adjusted from 5 MPa to 9 Mpa at a constant 501/min −1  flow rate or more by varying the number as well as the resilient properties of the plate springs and the pack of plate springs. 
     The valve plate  134  is provided with a central annular stepped recess or well  138 . When the valve plate is in the closed position, the lower face  138 A of the well and the upper face  140  of the valve plate serve to close off the passageways  108 A by mating with the complementally stepped planar faces  142  and  144  of the valve body. As the valve plate begins to open as a result of a pressure build-up within the passages  108 A, oil escapes through a composite opening  146  defined between adjacent faces  144  and  140  of the respective valve body and valve plate. The mating stepped arrangement between the piston and the valve plate contributes to the concentricity of the valve plate as it moves axially, which makes for more precise and consistent opening. 
     In prior art uniplanar valve plates, an immediate and marked pressure drop occurs as the valve plate opens against the counter-bias of the plate spring pack. This lends to cause the valve plate to be biased back to the closed or almost closed position, and then to cycle or hunt unstably between a closed and a partly open position. This is particularly prone to occur on initial opening of the valve plate, which tends to occur at a “cracking” force which is greater than the normal repetitive opening force. During the opening process, the annular well  138  acts as a pressure accumulator or sump to assist in the pressure build-up and to reduce a loss in pressure arising from the initial flow of fluid through the opening  146 . It has experimentally been found that the provision of the well  138  effectively reduces and stabilizes the end load or counter-force exerted by the piston assembly. The opening characteristics of the valve plate  134  can be varied by adjusting the width and/or depth of the well  138 . 
     In FIG. 13, graphs  150  and  152  show end load plotted against speed for a fixed orifice “non-relaxable” piston assembly in the case of the solid outline graph and a piston assembly of the invention in the case of the broken outline graph. The graphs are essentially identical for the FIG. 1 idle modes up to 2.604 m/s, as is shown at  154 . In the FIG. 10 mode of operation, which is shown at  156 , both of the graphs climb steeply, with the valve plate  134  beginning to open at just over 30 kN at an elbow  158  of the broken outline graph, with the graph then assuming a slightly upwardly slanting profile, as is shown at  160 , in which the end load increases very gradually up to around 32.5 kN at a speed of 3.623 ms −1 . In contrast, the solid outline graph  150  continues to climb steeply until an end load of just under 35 kN, and thereafter tapers off far less gradually, as is shown at  162 . It is thus clearly apparent that, whilst the mixed orifice version is able to absorb more energy, represented by the area between the respective curves, the relaxable orifice version of the present invention is more consistently responsive to an increase in end load, thereby effectively reducing the counter-force exerted by the rail retarder at high wagon speeds. This has the effect of reducing bounce, in particular with regard to lighter wagons travelling at higher speeds. In addition, as hunting or cycling of the valve plate is reduced, there is less resultant wear on the plate spring pack arising from rapid flexing thereof. 
     In FIG. 12, a typical force-stroke graph  164  is shown for a wagon retarder incorporating the piston assembly of FIGS. 1 to  3 . The nitrogen gas pressure was set to 700 kPa, at a temperature of 20° C. and an oil mass of 178 g. The incoming speed of the wagon was 2.688 ms −1 , and the outgoing speed was 2.604 ms −1 , with the load on each wheel being approximately 34 kN. The resultant energy extracted by the retarder, which is represented by the hatched area  164 A, was found to be 1273 J. The average retarding force was 15.9 kN, with the maximum force being 30.7 kN. The duration of the power or retarding stroke was found to be 110 ms, and the duration of the return stroke was found to be 241 ms, with the cylinder reaching a maximum downward velocity of 1.448 m/s.