Note: Descriptions are shown in the official language in which they were submitted.
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RAIL ROAD FREIGHT CAR WITH DAMPED SUSPENSION
Field of the Invention
This invention relates to the field of rail road freight cars.
Background of the Invention
This invention can be used with the invention described in my US Patent No.
6,659,016 entitled Rail Road Freight Car with Resilient Suspension, filed
August 1,
2001.
Auto rack rail road cars are used to transport automobiles. Typically, auto-
rack
rail road cars are loaded in the "circus loading" manner, by driving vehicles
into the cars
from one end, and securing them in places with chocks, chains or straps. When
the trip is
completed, the chocks are removed, and the cars are driven out.
Automobiles are a high value, relatively fragile type of lading. Damage due to
dynamic loading in the railcar may tend to arise principally in two ways.
First, there are
the longitudinal input loads transmitted through the draft gear due to train
line action or
shunting. Second, there are vertical, rocking and transverse dynamic responses
of the rail
road car to track perturbations as transmitted through the rail car
suspension. It would be
desirable to improve ride quality to lessen the chance of damage occurring.
In the context of longitudinal train line action, damage most often occurs
from
two sources (a) slack run-in and run out; (b) humping or flat switching. Rail
road car
draft gear have been designed against slack run-out and slack run-in during
train
operation, and also against the impact as cars are coupled together.
Historically, common
types of draft gear, such as that complying with, for example, AAR
specification M-901-
G, have been rated to withstand an impact at 5 m.p.h. (8 km/h) at a coupler
force of
500,000 Lbs. (roughly 2.2 x 106 N). Typically, these draft gear have a travel
of 2 3/4 to 3
1A inches in buff before reaching the 500,000 Lbs. load, and before "going
solid". The
term "going solid" refers to the point at which the draft gear exhibits a
steep increase in
resistance to further displacement. If the impact is large enough to make the
draft gear
"go solid" then the force transmitted, and the corresponding acceleration
imposed on the
lading, increases sharply. While this may be acceptable for ores, coal or
grain, it is
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undesirably severe for more sensitive lading, such as automobiles or auto
parts, rolls of
paper, fresh fruit and vegetables and other high value consumer goods such as
household
appliances or electronic equipment. Consequently, from the relatively early
days of the
automobile industry there has been a history of development of longer travel
draft gear to
provide lading protection for relatively high value, low density lading, in
particular
automobiles and auto parts, but also farm machinery, or tractors, or highway
trailers.
Historically, the need for slack was related, at least in part, to the
difficulty of
using a steam locomotive to "lift" (that is, move from a standing start) a
long string of rail
road cars with journal bearings, particularly in cold weather. For practical
purposes,
presently available diesel-electric locomotives are capable of lifting a unit
train of one
type of cars having little or no slack. Given the availability of locomotives
that develop
continuous high torque from a standing start, it is possible to re-examine the
issue of
slack action from basic principles. By eliminating, or reducing, the
accumulation of
slack, the use of short travel buff gear may tend to reduce the relative
longitudinal motion
between adjacent rail road cars, and may tend to reduce the associated
velocity
differentials and accelerations between cars. The use of short travel, or
ultra-short travel,
buff gear also has the advantage of eliminating the need for relatively
expensive, and
relatively complicated EOCC units, and the fittings required to accommodate
them.
In terms of dynamic response through the trucks, there are a number of loading
conditions to consider. First, there is a direct vertical response in the
"vertical bounce"
condition. This may typically arise when there is a track perturbation in both
rails at the
same point, such as at a level crossing or at a bridge or tunnel entrance
where there may
be a relatively sharp discontinuity in track stiffness. A second "rocking"
loading
condition occurs when there are alternating track perturbations, typically
such as used
formerly to occur with staggered spacing of 39 ft rails. This phenomenon is
less frequent
given the widespread use of continuously welded rails, and the generally lower
speeds,
and hence lower dynamic forces, used for the remaining non-welded track. A
third
loading condition arises from elevational changes between the tracks, such as
when
entering curves in which case a truck may have a tendency to warp. A fourth
loading
condition arises from truck "hunting", typically at higher speeds, where the
truck
oscillates transversely between the rails. During hunting, the trucks tend
most often to
deform in a parallelogram manner. Fifth, lateral perturbations in the rails
sometimes arise
where the rails widen or narrow slightly, or one rail is more worn than
another, and so on.
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There are both geometric and historic factors to consider related to these
loading
conditions. One historic factor is the near universal usage of the three-piece
style of
freight car truck in North America. While other types of truck are known, the
three piece
truck is overwhelmingly dominant in freight service in North America. The
three piece
truck relies on a primary suspension in the form of a set of springs trapped
in a "basket"
between the truck bolster and the side frames. For wheel load equalisation, a
three piece
truck uses one set of springs, and the side frames pivot about the truck
bolster ends in a
manner like a walking beam. The 1980 Car & Locomotive Cyclopedia, states at
page
669 that the three piece truck offers "interchangeability, structural
reliability and low first
cost but does so at the price of mediocre ride quality and high cost in terms
of car and
track maintenance". It would be desirable to retain many or all of these
advantages while
providing improved ride quality.
In terms of rail road car truck suspension loading regimes, the first
consideration
is the natural frequency of the vertical bounce response. The static
deflection from light
car (empty) to maximum laded gross weight (full) of a rail car at the coupler
tends to be
typically about 2 inches. In addition, rail road car suspensions have a
dynamic range in
operation, including a reserve travel allowance.
In typical historical use, springs were chosen to suit the deflection under
load of a
full coal car, or a full gain car, or fully loaded general purpose flat car.
In each case, the
design lading tended to be very heavy relative to the rail car weight. For
example, the
live load for a 286,000 lbs. car may be of the order of five times the weight
of the dead
sprung load (i.e., the weight of the car, including truck bolsters but less
side frames, axles
and wheels). Further, in these instances, the lading may not be particularly
sensitive to
abusive handling. That is, neither coal nor grain tends to be badly damaged by
poor ride
quality. As a result, these cars tend to have very stiff suspensions, with a
dominant
natural frequency in vertical bounce mode of about 2 Hz. when loaded, and
about 4 to 6
Hz. when empty. Historically, much effort has been devoted to making freight
cars light
for at least two reasons. First, the weight to be back hauled empty is kept
low, reducing
the fuel cost of the backhaul. Second, as the ratio of lading to car weight
increases, a
higher proportion of hauling effort goes into hauling lading, rather than
hauling the
railcar.
By contrast, an autorack car, or other type of car for carrying relatively
high
value, low density lading such as auto parts, electronic consumer goods, or
white goods
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more generally, has the opposite loading profile. A two unit articulated
autorack car may
have a light car (i.e., empty) weight of 165,000 lbs., and a lading weight
when fully
loaded of only 35 ¨ 40,000 lbs., per car body unit. That is, not only may the
weight of
the lading be less than the sprung weight of the rail road car unit, it may be
less than 40
% of the car weight. The lading typically has a high, or very high, ratio of
value to
weight. Unlike coal or grain, automobiles are relatively fragile, and hence
more sensitive
to a gentle (or a not so gentle) ride. As a relatively fragile, high value,
high revenue form
of lading, it may be desirable to obtain superior ride quality to that
suitable for coal or
grain.
One way to improve ride quality is to increase the dead sprung weight of the
rail
road car body. Another way to improve ride quality is to decrease the spring
rate.
Decreasing the spring rate involves further considerations. Historically the
deck height
of a flat car tended to be very closely related to the height of the upper
flange of the
center sill. This height was itself established by the height of the cap of
the draft pocket.
The size of the draft pocket was standardised on the basis of the coupler
chosen, and the
allowable heights for the coupler knuckle. The deck height usually worked out
to about
41 inches above top of rail. For some time auto rack cars were designed to a
19 ft height
limit. To maximise the internal loading space, it has been considered
desirable to lower
the main deck as far as possible, particularly in tri-level cars. Since the
lading is
relatively light, the rail car trucks have tended to be light as well, such as
70 Ton trucks,
as opposed to 100, 110 or 125 Ton trucks for coal, ore, or grain cars at
263,000, 286,000
or 315,000 lbs. gross weight on rail. Since the American Association of
Railroads (AAR)
specifies a minimum clearance of 5" above the wheels, the combination of low
deck
height, deck clearance, and minimum wheel height set an effective upper limit
on the
spring travel, and reserve spring travel range available. If softer springs
are used, the
remaining room for spring travel below the decks may well not be sufficient to
provide
the desired reserve height. In consequence, the present inventor proposes,
contrary to
lowering the main deck, that the main deck be higher than 42 inches to allow
for more
spring travel.
As noted above, many previous auto rack cars have been built to a 19 ft
height.
Another major trend in recent years has been the advent of "double stack"
intermodal
container cars capable of carrying two shipping containers stacked one above
the other in
a well or to other freight cars falling within the 20 ft 2 in. height limit of
AAR plate H.
Many main lines have track clearance profiles that can accommodate double
stack cars.
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Consequently, it is now possible to use auto rack cars built to the higher
profile of the
double stack intermodal container cars.
While decreasing the primary vertical bounce natural frequency appears to be
advantageous for auto rack rail road cars generally, including single car unit
auto rack rail
road cars, articulated auto rack cars may also benefit not only from adding
ballast, but
from adding ballast preferentially to the end units near the coupler end
trucks. As
explained more fully in the description below, the interior trucks of
articulated cars tend
to be more heavily burdened than the end trucks, primarily because the
interior trucks
share loads from two adjacent car units, while the coupler end trucks only
carry loads
from one end of one car unit. It would be advantageous to even out this
loading so that
the trucks have roughly similar vertical bounce frequencies.
Three piece trucks currently in use tend to use friction dampers, sometimes
assisted by hydraulic dampers such as can be mounted, for example, in the
spring set.
Friction damping has most typically been provided by using spring loaded
blocks, or
snubbers, mounted with the spring set, with the friction surface bearing
against a mating
friction surface of the columns of the side frames, or, if the snubber is
mounted to the
side frame, then the friction surface is mounted on the face of the truck
bolster. There are
a number of ways to do this. In some instances, as shown at p. 847 of the 1984
Car &
Locomotive Cyclopedia lateral springs are housed in the end of the truck
bolster, the
lateral springs pushing horizontally outward on steel shoes that bear on the
vertical faces
of the side columns of the side frames. This provides roughly constant
friction (subject to
the wear of the friction faces), without regard to the degree of compression
of the main
springs of the suspension.
In another approach, as shown at p. 715 of the 1997 Car & Locomotive
Cyclopedia, one of the forward springs in the main spring group, and one of
the rearward
springs in the main spring group bear upon the underside, or short side, of a
wedge. One
of the long sides, typically an hypotenuse of a wedge, engages a notch, or
seat, formed
near the outboard end of the truck bolster, and the third side has the
friction face that
abuts, and bears against, the friction face of the side column (either front
or rear, as the
case may be), of the side frame. The action of this pair of wedges then
provides damping
of the various truck motions. In this type of truck the friction force varies
directly with
the compression of the springs, and increases and decreases as the truck
flexes. In the
vertical bounce condition, both friction surfaces work in the same direction.
In the
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warping direction (when one wheel rises or falls relative to the other wheel
on the same
side, thus causing the side frame to pivot about the truck bolster) the
friction wedges
work in opposite directions against the restoring force of the springs.
The "hunting" phenomenon has been noted above. Hunting generally occurs on
tangent (i.e., straight) track as railcar speed increases. It is desirable for
the hunting
threshold to occur at a speed that is above the operating speed range of the
rail car.
During hunting the side frames tend to want to rotate about a vertical axis,
to a non-
perpendicular angular orientation relative to the truck bolster sometimes
called
"parallelogramming" or lozenging. This will tend to cause angular deflection
of the
spring group, and will tend to generate a squeezing force on opposite diagonal
sides of
the wedges, causing them to tend to bear against the side frame columns. This
diagonal
action will tend to generate a restoring moment working against the angular
deflection.
The moment arm of this restoring force is proportional to half the width of
the wedge,
since half of the friction plate lies to either side of the centreline of the
side frame. This
tends to be a relatively weak moment connection, and the wedge, even if wider
than
normal, tends to be positioned over a single spring in the spring group.
Typically, for a truck of fixed wheelbase length, there is a trade-off between
wheel load equalisation and resistance to hunting. Where a car is used for
carrying high
density commodities at low speeds, there may tend to be a higher emphasis on
maintaining wheel load equalisation. Where a car is light, and operates at
high speed
there will be a greater emphasis on avoiding hunting. In general, the
parallelogram
deformation of the truck in hunting is deterred by making the truck laterally
more stiff.
One approach to discouraging hunting is to use a transom, typically in the
form of a
channel running from between the side frames below the spring baskets. Another
approach is to use a frame brace.
One way to address the hunting issue is to employ a truck having a longer
wheelbase, or one whose length is proportionately great relative to its width.
For
example, at present two axle truck wheelbases may range from about 5' ¨ 3" to
6' ¨ 0".
However, the standard North America track gauge is 4' ¨ 8 1/2", giving a
wheelbase to
track width ratio possibly as small as 1.12. At 6' ¨ 0" the ratio is roughly
1.27. It would
be preferable to employ a wheelbase having a longer aspect ratio relative to
the track
gauge. As described herein, one aspect of the present invention employs a
truck with a
longer wheelbase, preferably about 80 or 86 inches, giving a ratio of 1.42 or
1.52. This
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increase in wheelbase length may tend also to be benign in terms of wheel
loading
equalisation.
In a typical spring seat and spring group arrangement, the side frame window
may
typically be of the order of 21 inches in height from the spring seat base to
the underside
of the overarching compression member, and the width of the side frame window
between the wear plates on the side frame columns is typically about 18",
giving a side
frame window that is taller than wide in the ratio of about 7:6. Similarly,
the bottom
spring seat has a base that is typically about 18 inches long to correspond to
the width of
the side frame window, and about 16 inches wide in the transverse direction,
that is being
longer than wide. It may be advantageous to make the side frame windows wider,
and
the spring seat correspondingly longer to accommodate larger diameter long
travel
springs with a softer spring rate. At the same time, lengthening the wheel
base of the
truck may also be advantageous since it is thought that a longer wheelbase may
ameliorate truck hunting performance, as noted above. Such a design change is
counter-
intuitive since it may generally be desired to keep truck size small, and
widening the
unsupported window span may not have been considered desirable heretofore.
Another way to raise the hunting threshold is to increase the parallelogram
stiffness between the bolster and the side frames. It is possible, as
described herein, to
employ pairs of wedges, of comparable size to those previously used, the two
wedges
being placed side by side and each individually supported by a different
spring, or being
the outer two wedges in a three deep spring group, to give a larger moment arm
to the
restoring force and to the damping associated with that force.
The use of multiple variable friction force dampers in which the wedges are
mounted over members of the spring group, is shown in US Patent 3,714,905 of
Barber,
issued February 6, 1973. The damper arrangement shown by Barber is not
apparently
presently available in the market, and does not seem ever to have been made
available
commercially.
Notably, the damper wedges shown in Barber appear to have relatively sharply
angled wedges, with an included angle between the friction face (i.e., the
face bearing
against the side frame column) and the sliding face (i.e., the angled face
seated in the
damper pocket formed in the bolster, typically the hypotenuse) of roughly 35
degrees.
The angle of the third, or opposite, horizontal side face, namely the face
that seats on top
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of the vertically oriented spring, is the complementary angle, in this
example, being about
55 degrees. It should be noted that as the angle of the wedge becomes more
acute, (i.e.,
decreasing from about 35 degrees) the wedge may have an undesirable tendency
to jam in
the pocket, rather than slide.
Barber, above, shows a spring group of variously sized coils with four
relatively
small corner coils loading the four relatively sharp angled dampers. From the
relative
sizes of the springs illustrated, it appears that Barber was contemplating a
spring group of
relatively traditional capacity ¨ a load of about 80,000 lbs., at a "solid"
condition of 3
1/16 inches of travel, for example, and an overall spring rate for the group
of about
25,000 lbs/inch, to give 2 inches of overall rail car static deflection for
about 200,000 lbs
live load.
Apparently keeping roughly the same relative amount of damping overall as for
a
single damper, Barber appears to employ individual B331 coils (k = 538 lb/in,
(+/-))
under each friction damper, rather than a B432 coil (k = 1030 lb/in, (+/-)) as
might
typically have been used under a single damper for a spring group of the same
capacity.
As such, it appears that Barber contemplated that springs accounting for
somewhat less
than 15 % of the overall spring group stiffness would underlie the dampers.
These spring stiffnesses might typically be suitable for a rail road car
carrying
iron ore, gain or coal, where the lading is not overly fragile, and the design
ratio of live
load to dead sprung load is typically greater than 3: 1. It might not be
advantageous for a
rail road car for transporting automobiles, auto parts, consumer electronics
or other white
goods of relatively low density and high value where the design ratio of live
load to dead
sprung load may be well less than 2: 1, and quite possibly lying in the range
of 0.4: 1 to
1: 1.
It has been noted that the frictional force produced by friction damper wedges
differs depending on whether the damper is being loaded, or unloaded. In the
terminology employed, the damper is being "loaded" when the bolster is moving
downward in the sideframe window, since the spring force is increasing, and
hence the
load, or force on the damper is increasing. Similarly, the damper is being
"unloaded"
when the bolster is moving upward toward the top of the sideframe window,
since the
force in the springs, and hence the load in the wedges, is decreasing.
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The equations can be written as
While loading: Fd = Itic Fs ( Cot (4)) - l's)
1
While unloading: Fd = cFs ( Cot (4)) + )
1
Where: Fd = friction force on the sideframe column
Fs = force in the spring
Fts = friction coefficient of the angled face on the bolster
pc is the coefficient of friction against the sideframe column
st is the included angle between the angled face on the bolster and
the friction face bearing against the column
For a given angle, a friction load factor, Cf can be determined as Cf =Fd / F.
This
load factor Cf will tend to be different depending on whether the bolster is
moving up or
down. A graph of upward and downward load factors as a function of wedge angle
is
shown in Figure 7 based on a ps of 0.2 and a it, of 0.4, values which are
thought to be
roughly representative of service conditions.
When the wheels encounter a perturbation in the rail, their reaction to the
perturbation will tend to transmit a force through the suspension into the
rail road car
body. The force transmitted will tend to be the sum of the spring force plus
the friction
force in the dampers. For a relatively gentle ride, it is desirable that the
damping force as
the wheels move up relative to the carbody not be excessive, and that the
damping be
stronger when the car body is moving upward relative to the wheels.
With a relatively sharply angled wedge, as typified by wedges in the 30 - 35
degree range such as appear to be shown by Barber, and as employed in wedges
known
to be commonly in use, the load factor may tend to be significantly higher
when the
bolster is moving downward relative to the side frame than when the bolster is
moving
upward. It may be desirable to lessen, or reverse this relationship, as may
tend to occur
for angles above about 40 to 45 degrees. (See Figure 7).
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In the past, spring groups have been arranged such that the spring loading
under
the dampers has been proportionately small. That is, the dampers have
typically been
seated on side spring coils, as shown in the AAR standard spring groupings
shown in the
1997 Car & Locomotive Cyclopedia at pages 743 ¨ 746, in which the side spring
coils,
inner and outer as may be, are often B321, B331, B421, B422, B432, or B433
springs as
compared to the main spring coils, such that the springs under the dampers
have lower
spring rates than the other coil combinations in the other positions in the
spring group.
As such, the dampers may be driven by less than 15 % of the total spring
stiffness of the
group generally.
In US Patent 5,046,431 of Wagner, issued September 10, 1991, the standard
inboard-and-outboard gib arrangement on the truck bolster was replaced by a
single
central gib mounted on the side frame column for engaging the shoulders of a
vertical
channel defined in the end of the truck bolster. In doing this, the damper was
split into
inboard and outboard portions, and, further, the inboard and outboard
portions, rather
than lying in a common transverse vertical plane, were angled in an outwardly
splayed
orientation.
Wagner's gib and damper arrangement may not necessarily be desirable in
obtaining a desired level of ride quality. In obtaining a soft ride it may be
desirable that
the truck be relatively soft not only in the vertical bounce direction, but
also in the
transverse direction, such that lateral track perturbations can be taken up in
the
suspension, rather than be transmitted to the car body, (and hence to the
lading), as may
tend undesirably to happen when the gibs bottom out (i.e., come into hard
abutting
contact with the side frame) at the limit of horizontal travel.
The present inventor has found it desirable that there be an allowance for
lateral
travel of the truck bolster relative to the wheels of the order of 1 to 1 -
1/2 inches to either
side of a neutral central position. Wagner does not appear to have been
concerned with
this issue. On the contrary, Wagner appears to show quite a tight gib
clearance, with
relatively little travel before solid contact. Furthermore, transverse
displacement of the
truck bolster relative to the side frame is typically resiliently resisted by
the horizontal
shear in the spring groups, and by the pendulum motion of the side frames
rocking on the
crowns of the bearing adapters, these two components being combined like
springs in
series. Wagner's canted dampers appear to make lateral translation of the
bolster stiffer,
rather than softer. This may not be advantageous for relatively fragile
lading. In the view
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of the present inventor, while it is advantageous to increase resistance to
the hunting
phenomenon, it may not be advantageous to do so at the expense of increasing
lateral
stiffness.
It is desirable that a relatively larger portion of the spring effort be used
to load
the dampers, with the employment of a larger damper wedge angle. As such, the
same
magnitude of damping force may tend to be achieved with a combination of
relatively
softer springs than previously used, with a larger included angle in the
wedges.
Alternatively, a greater damping force than before may be achieved with wedges
having a
relatively modest angle with springs of the same stiffness as before, the
included angle
being chosen in the 45 to 65 degree range. The opportunity to vary wedge angle
and
spring stiffness thus gives an opportunity to tune the amount of damping in
some
measure. In addition, it would be advantageous to use a larger included angle
in the
wedge, both for these reasons, and because wedges with a larger included angle
may tend
to be less prone to jamming and may result in more favourable dynamic
behaviour as
indicated by Figure 7.
In the damper groups themselves, it is thought that parallelogram deflection
of the
truck such that the truck bolster is not perpendicular to the side frame, as
during hunting,
may tend to cause the dampers to try to twist angularly in the damper seats.
In that
situation one corner of the damper may tend to be squeezed more tightly than
the other.
As a result, the tighter corner may try to retract relative to the less tight
corner, causing
the damper wedge to squirm and rotate somewhat in the pocket. This tendency to
twist
may also tend to reduce the squaring, or restoring force that tends to move
the truck back
into a condition in which the truck bolster is square relative to the side
frames.
Consequently, it may be desirable to discourage this twisting motion by
limiting
the freedom to twist, as, for example, by introducing a groove or ridge, or
keyway, or
channel feature to govern the operation of the spring in the damper pocket. It
may also
be advantageous to use a split wedge to discourage twisting, such that one
portion of the
wedge can move relative to the other, thus finding a different position in a
linear sense
without necessarily forcing the other portion to twist. Further still, it may
be
advantageous to employ a means for encouraging a laterally inboard portion of
the
damper, or damper group, to be biased to its most laterally inboard position,
and a
laterally outboard portion of the damper, or the damper group, to be biased to
its most
laterally outboard position. In that way, the moment arm of the restoring
force may tend
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to remain closer to its largest value. One way to do this, as described in the
description of
the invention, below, is to add a secondary angle to the wedge.
In the terminology herein, wedges have a primary angle 4), namely the included
angle between (a) the sloped damper pocket face mounted to the truck bolster,
and (b) the
side frame column face, as seen looking from the end of the bolster toward the
truck
center. This is the included angle described above. A secondary angle is
defined in the
plane of angle 4), namely a plane perpendicular to the vertical longitudinal
plane of the
(undeflected) side frame, tilted from the vertical at the primary angle. That
is, this plane
is parallel to the (undeflected) long axis of the truck bolster, and taken as
if sighting along
the back side (hypotenuse) of the damper.
The secondary angle 13 is defined as the lateral rake angle seen when looking
at
the damper parallel to the plane of angle 4). As the suspension works in
response to track
perturbations, the wedge forces acting on the secondary angle will tend to
urge the
damper either inboard or outboard according to the angle chosen. Inasmuch as
the
tapered region of the wedge may be quite thin in terms of vertical through-
thickness, it
may be desirable to step the sliding face of the wedge (and the co-operating
face of the
bolster seat) into two or more portions. This may be particularly so if the
angle of the
wedge is large.
Split wedges and two part wedges having a chevron, or chevron like, profile
when
seen in the view of the secondary angle can be used. Historically, split
wedges have been
deployed as a pair over a single spring, the split tending to permit the
wedges to seat
better, and to remain better seated, under twisting condition than might
otherwise be the
case. The chevron profile of a solid wedge may tend to have the same intent of
preventing rotation of the sliding face of the wedge relative to the bolster
in the plane of
the primary angle of the wedge. Split wedges and compound profile wedges can
be
employed in pairs as described herein.
In a further variation, where a single broad wedge is used, with a compound or
other profile, it may be desirable to seat the wedge on two or more springs in
an inboard-
and-outboard orientation to create a restoring moment such as might not tend
to be
achieved by a single spring alone. That is, even if a single large wedge is
used, the use of
two, spaced apart springs may tend to generate a restoring moment if the wedge
tries to
twist, since the deflection of one spring may then be greater that the other.
CA 02797026 2013-11-28
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When the dampers are placed in pairs, either immediately side-by-side or with
spacing between the pairs, the restoring moment for squaring the truck will
tend not only
to be due to the increase in compression to one set of springs due to the
extra tendency to
squeeze the dampers downward in the pocket, but due to the difference in
compression
between the springs that react to the extra squeezing of one diagonal set of
dampers and
the springs that act against the opposite diagonal pair that will tend to be
less tightly
squeezed.
The bolster is typically permitted to travel laterally to either side relative
to the
side frames, and for the side frames to have limited angular rotation about an
axis parallel
to the longitudinal axis of the rail car more generally. It is desirable that
after an initial
perturbation, the bolster return to a central, angularly squared position. An
increase in
the normal force at the friction face, as discussed, may tend to return the
side frames to a
"square" condition relative to the truck bolster. In sideways displacement,
return of the
truck to a centered position may tend to cease when the friction in the
dampers matches
the lateral restoring force in the spring groups. This tendency may be reduced
by the
tendency of the springs to return to a laterally centered position as the
truck works in the
vertical bounce and warp conditions. However, it may be desirable to enhance
this
restoring tendency. In the view of the present inventor it may be advantageous
to install
some, or all of the springs in the inner and outer rows of the spring group at
a slight
anhedral angle relative to each other, so that they form a symmetrical V.
Summary of the Invention
In an aspect of the invention there is a rail road freight car having at least
one rail
car unit. The rail road freight car is supported by three piece rail car
trucks for rolling
motion along rail road tracks. At least a first rail car truck of the three
piece rail car
trucks has a rigid truck bolster and a pair of first and second side frame
assemblies. The
truck bolster has first and second ends. The first rail car truck has a
suspension including
first and second spring groups mounted between the first and second ends of
the truck
bolster and the first and second side frames respectively. The rail car truck
suspension
has a natural vertical bounce frequency of less than 4.0 Hz. when the rail
road freight car
is unloaded. A first set of friction dampers is mounted between the truck
bolster and the
first side frame assembly. A second set of friction dampers is mounted between
the truck
CA 02797026 2013-11-28
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bolster and the second side frame assembly. The first set of friction dampers
includes at
least a first friction damper and a second friction damper. The first and
second friction
dampers are independently biased, and the second friction damper is mounted
more
laterally outboard than the first friction damper.
In an additional feature of that aspect of the invention, the set of dampers
includes
at least third and fourth friction dampers. The third and fourth friction
dampers are
independently biased, and the third friction damper is mounted more laterally
outboard
than the fourth friction damper. The third friction damper is longitudinally
spaced
relative to the first friction damper, and the fourth friction damper is
longitudinally
spaced relative to the second friction damper. In another additional feature,
the
suspension is at rest on a straight track, a transverse vertical plane bisects
the truck
bolster to define a plane of symmetry, and the first, second, third and fourth
friction
dampers are arranged in a symmetrical formation relative to the transverse
vertical plane.
In yet another additional feature, a longitudinal vertical plane intersects
the side
frame and the first, third, second and fourth dampers are symmetrically
arranged in a
symmetrical formation relative to the longitudinal vertical plane. In still
another
additional feature, the four dampers are arranged in a formation that is both
longitudinally
and transversely symmetrical.
In a further additional feature, the first damper has a first friction face.
The
second damper has a second friction face. The first friction face lies in a
first plane. The
second friction face lies in a second plane, and the first and second planes
have mutually
parallel normal vectors. In yet a further additional feature, the first damper
has a first
friction face. The second damper has a second friction face, and the first and
second
friction faces are coplanar. In still a further additional feature, the first
and second
dampers sit side-by-side. In another additional feature, the first and second
dampers are
transversely spaced from each other. In still another additional feature, the
first and
second dampers are separated by a land, and a spring of one of the spring
groups acts
against the land.
In yet another additional feature, the natural vertical bounce frequency is
less than
3 Hz. when the rail road car is unladen. In still yet another additional
feature, the first
and second spring groups each have a vertical bounce spring rate, and the
vertical bounce
spring rate is less than 20,000 lbs per inch, per spring group. In an
additional feature, the
CA 02797026 2013-11-28
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first and second spring groups each have a vertical bounce spring rate, and
the vertical
bounce spring rate is less than 12,000 lbs per inch, per spring group.
In another additional feature, the side frames have wear plates facing the
bolster
ends. The sets of friction dampers are mounted in pockets defined in the ends
of the
bolsters, and the friction dampers have friction faces bearing against the
wear plates of
the side frames. In yet another additional feature, the first and second
friction dampers
bear on a common wear plate. In still another additional feature, the wear
plate presents
an uninterrupted surface to the first and second dampers.
In a further additional feature, the first and second dampers each include an
angled wedge seated in one of the pockets of the bolster. In yet a further
additional
feature, the angled wedge has a first surface slidingly engaged in a first
pocket of the
pockets of the bolster. The first surface is inclined at a primary angle
defined between
the first surface and the friction face thereof. The primary angle is greater
than 35
degrees. The pocket has a mating inclined surface. In still a further feature,
the wedge
first surface has a secondary angle cross-wise to the first angle.
In still yet a further additional feature, the first damper is biased
laterally inboard
and the second damper is biased laterally outboard. In another additional
feature, a
friction discouraging material is applied to enhance sliding of the first
damper relative to
the bolster pocket. In yet another additional feature, at least one of the
dampers is a split
damper. In still another additional feature, the split damper is laterally
asymmetrically
biased in a direction chosen from (a) laterally inboard; and (b) laterally
outboard. In still
yet another additional feature, the angled surface is stepped. In a further
additional
feature, the wedge has an inclined chevron cross-section. The chevron has
asymmetric
wings. In another additional feature, the wedge has an inclined chevron cross-
section,
one wing of the chevron lying at a steeper angle than the other. In yet
another additional
feature, the wedge has a pair of first and second inclined flanks. One of the
flanks is
steeper than the other.
In still another additional feature, the first spring group has an overall
vertical
bounce spring rate, 1(1. A portion of the spring group provides biasing for
the dampers.
The portion has a summed vertical spring rate, 1c2, that is at least 20 % of
the overall
vertical bounce spring rate. In still yet another additional feature, the
ratio of k2 k1 is at
least as great as 1/4. In another additional feature, the ratio of k2 k1 is at
least as great as
CA 02797026 2013-11-28
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-
1/3. In yet another additional feature, the ratio of k2 k1 is at least as
great as 4/9. In
another additional feature, the spring groups include coil springs and first
and second
dampers seat on coils having an outer diameter of greater than 7 1/2 inches.
In still
another additional feature, the spring groups include coil springs and the
first and second
dampers seat on coils having an outer diameter of greater than 5 inches. In
yet another
additional feature, the side frames are mounted to a wheelset, and the truck
bolster has at
least one inch of lateral travel to either side relative to the wheelset. In a
further
additional feature, the first side frame is swingingly mounted on wheel
bearings, and the
first side frame, by itself, has a transverse swinging natural frequency of
less than 1.4 Hz.
In still a further additional feature, the side frames are mounted on a
wheelset. The first
truck has a natural frequency for lateral displacement of the truck bolster
relative to the
wheelset; and the natural frequency for lateral displacement is less than 1.0
Hz.
In yet a further additional feature, the first truck has an AAR rating of at
least "70
Ton". In still a further additional feature, the first truck has a capacity
chosen from the
set of rail road freight car truck capacities consisting of (a) 70 Ton; (b) 70
Ton Special;
(c) 100 Ton; (d) 110 Ton; and (e) 125 Ton. In another additional feature, the
first truck
has a wheelset having wheels of greater than 33 inches in diameter.
In a further additional feature, each of the side frames has a pair of side
frame
columns. The side frame columns have bearing surfaces for engaging the
friction
dampers. A bolster window is defined therebetween. The bolster window has a
height
and a width. The width is measured between the friction faces and the width is
greater
than the depth. In another additional feature, the width is at least 8/7 as
large as the
depth. In still another additional feature, the width is at least 24 inches.
In yet another
additional feature, the first and second side frames each respectively have a
spring seat
for receiving, respectively. The first and second spring groups, and the
spring seat has a
transverse width of greater than 15 inches. In still yet another additional
feature, the
spring seat has a length of at least 24 inches.
In a further additional feature, at least one rail car unit has ballasting
supported by
the first truck. In yet a further additional feature, the rail road car is an
articulated rail
road car. In still a further additional feature, the rail car unit is an end
car unit of the
articulated rail road car. The end car unit has a coupler end and an
articulated connector
end, and the first rail car truck supports the coupler end of the end car
unit. In another
additional feature, the end car unit, when empty, has a weight distribution
asymmetrically
CA 02797026 2013-11-28
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biased toward the first truck. In yet another additional feature, the end car
unit has
ballasting distributed asymmetrically heavily toward the coupler end thereof.
In still
another additional feature, the rail car unit has a deck carried above the
first truck upon
which lading can be carried. In a further additional feature, the deck is
surmounted by a
housing for protection the lading. In yet a further additional feature, the
housing has
doors giving access to the deck. In still a further additional feature, the
deck is a circus-
loading deck upon which wheeled vehicles can be conducted. In another
additional
feature, the rail road freight car is an auto-rack rail road car.
In yet another additional feature, the rail car unit has at least a first
coupler end,
and a coupler mounted thereat. The coupler has less than 25/32" of slack. In
still another
additional feature, the rail car unit has at least a first end, and a coupler
mounted thereat.
The couplers are chosen from the set of coupler families consisting of (a) AAR
Type F
couplers; (b) AAR Type H couplers; and (c) AAR Type CS couplers. In still yet
another
additional feature, the rail car unit has at least a first coupler end, draft
gear mounted
thereat and a coupler mounted to the draft gear. The draft gear has a
deflection of less
than 2 'A inches at 500,000 lbs buff load. In a further additional feature,
the rail car unit
has at least a first coupler end, draft gear mounted thereat and a coupler
mounted to the
draft gear. The draft gear has a deflection of less than 1 inch at 700,000 lbs
buff load.
In another aspect of the invention there is a railroad three piece freight car
truck.
The truck has a rigid truck bolster and a pair of first and second side frame
assemblies.
The truck bolster has first and second ends. A resilient suspension includes
first and
second spring groups mounted between the first and second ends of the truck
bolster and
the first and second side frames respectively. The resilient suspension of the
first of the
trucks has a vertical bounce spring rate of less than 20,000 lbs per spring
group. A first
set of friction dampers is mounted between the truck bolster and the first
side frame
assembly. A second set of friction dampers is mounted between the truck
bolster and the
second side frame assembly. The set of friction dampers includes at least a
first friction
damper and a second friction damper. The first and second friction dampers are
independently biased, and the first friction damper is mounted more laterally
outboard
than the second friction damper.
In another aspect of the invention there is a rail road freight car three
piece truck.
The truck has a bolster, a pair of first and second side frames, a pair of
first and second
spring groups, and a wheelset. The sideframes are mounted to the wheelset, and
the
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bolster is mounted transversely relative to the side frames. The spring groups
are
mounted in the sideframes. The bolster has first and second ends resiliently
supported by
the first and second spring groups. Each of the spring groups has a vertical
spring rate of
less than 20,000 lbs/in. A first set of friction dampers is mounted to act
between the first
end of the truck bolster and the first side frame, and a second set of
friction dampers is
mounted to act between the second end of the bolster and the second side
frame. Each of
the sets of friction dampers include four dampers arranged in a four cornered
layout.
In another aspect of the invention there is a railroad freight car three piece
truck
for rolling along rail road tracks. The tracks have a gauge width. The three
piece truck
has a bolster, a pair of first and second side frames, a pair of first and
second spring
groups, and a pair of first and second axles each having wheels mounted at
opposite ends
thereof. The wheelset has a longitudinal wheelbase and a transverse track
width
corresponding to the gauge width. The wheelbase is at least 1.3 times as great
as the
gauge width. The side frames are mounted to the wheelset, and the bolster is
mounted
transversely relative to the side frames. The spring groups are mounted in the
sideframes.
The bolster has first and second ends resiliently supported by the first and
second spring
groups. A first set of friction dampers is mounted to act between the first
end of the truck
bolster and the first side frame, and a second set of friction dampers is
mounted to act
between the second end of the bolster and the second side frame. Each of the
sets of
friction dampers include four individually sprung dampers arranged in a four
cornered
layout.
In another aspect of the invention there is a railroad freight car three piece
truck.
The truck has a bolster, a pair of first and second side frames, a pair of
first and second
spring groups, and a wheelset. The side frames are mounted to the wheelset,
and the
bolster is mounted transversely relative to the side frames. The spring groups
are
mounted in the side frames. The bolster has first and second ends resiliently
supported
by the first and second spring groups. A first set of friction dampers is
mounted to act
between the first end of the truck bolster and the first side frame, and a
second set of
friction dampers is mounted to act between the second end of the bolster and
the second
side frame. Each of the sets of friction dampers include four individually
sprung dampers
arranged in a four cornered layout. The four dampers include a first
transversely inboard
damper and a first transversely outboard damper, seated in respective first
and second
damper pockets. The first transversely inboard damper is biased to a
transversely inboard
CA 02797026 2013-11-28
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position in the first damper pocket, and the first transversely outboard
damper is biased to
a transversely outboard position in the second damper pocket.
In another aspect of the invention there is a railroad freight car three piece
truck.
The truck has a bolster, a pair of first and second side frames, a pair of
first and second
spring groups, and a wheelset. The side frames are mounted to the wheelset,
and the
bolster is mounted transversely relative to the side frames. The spring groups
are
mounted in the side frames. The bolster has first and second ends resiliently
supported
by the first and second spring groups. A first set of friction dampers is
mounted to act
between the first end of the truck bolster and the first side frame, and a
second set of
friction dampers is mounted to act between the second end of the bolster and
the second
side frame. Each of the sets of friction dampers include four individually
sprung dampers
arranged in a four cornered layout. The dampers are wedge shaped. The wedge
shapes
have a primary angle of greater than 35 degrees.
In another aspect of the invention there is a railroad freight car three piece
truck.
The truck has a bolster, a pair of first and second side frames, a pair of
first and second
spring groups, and a wheelset. The side frames are mounted to the wheelset,
and the
bolster is mounted transversely relative to the side frames. The spring groups
are
mounted in the side frames. The bolster has first and second ends resiliently
supported
by the first and second spring groups. A first set of friction dampers is
mounted to act
between the first end of the truck bolster and the first side frame. A second
set of friction
dampers is mounted to act between the second end of the bolster and the second
side
frame. Each of the sets of friction dampers include four dampers arranged in a
four
cornered layout. Each of the bolster ends has a set of damper pockets for
receiving the
first and second sets of the dampers. The dampers are damper wedges having a
spring
loaded base portion, a friction face for engaging a friction wear plate, and a
sliding face
for engaging the damper pockets. The wedges have a primary wedge angle between
the
friction face and the sliding face of greater than 35 degrees.
In another aspect of the invention there is a railroad freight car three piece
truck.
The truck has a bolster, a pair of first and second side frames, a pair of
first and second
spring groups, and a wheelset. The side frames are mounted to the wheelset,
and the
bolster is mounted transversely relative to the side frames. The spring groups
are
mounted in the side frames. The bolster has first and second ends resiliently
supported
by the first and second spring groups. A first set of friction dampers are
mounted to act
CA 02797026 2013-11-28
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between the first end of the truck bolster and the first side frame, and a
second set of
friction dampers are mounted to act between the second end of the bolster and
the second
side frame. Each of the sets of friction dampers include four dampers arranged
in a four
cornered layout. The dampers are wedge shaped. The wedge shapes have a primary
angle of greater than 35 degrees.
In another aspect of the invention there is a rail road freight car three
piece truck.
The truck has a bolster, a pair of first and second side frames, a pair of
first and second
spring groups, and a wheelset. The side frames are mounted to the wheelset,
and the
bolster is mounted transversely relative to the side frames. The spring groups
are
mounted in the side frames. The bolster has first and second ends resiliently
supported
by the first and second spring groups. Each of the spring groups have an
overall vertical
spring rate. Each of the spring groups include a plurality of springs. A first
set of friction
dampers are mounted to act between the first end of the truck bolster and the
first side
frame, and a second set of friction dampers are mounted to act between the
second end of
the bolster and the second side frame. Each of the sets of friction dampers
are sprung on
damper loading members of plurality of springs of the spring groups. The
damper
loading member of each of the spring groups account for at least 25 % of the
vertical
spring rate of each of the spring groups.
In another aspect of the invention there is a railroad freight car. The
freight car
has a rail road car body carried on rail road car trucks for rolling motion
along rail road
car tracks. The rail road car body has a deck for carrying lading, and side
sills running
alongside the deck. The body has a first end and a second end. The body has a
coupler
mounted to at least the first end of the body. A main bolster is mounted to
the body
adjacent to the first end of the body longitudinally inboard of the coupler.
The main
bolster extends transversely between the side sills. The main bolster has
first and second
arms extending laterally outboard from a centreplate. The centreplate is
mounted to a
first of the rail road car trucks on a vertical axis defining a truck center.
The first rail
road car truck has wheels spaced laterally outboard a half track gauge width
distance
from the truck center. The arms of the main bolster has a wheel clearance
portion
extending laterally away from the truck center over a range of distance
bracketing the
half track gauge width distance. The wheel clearance portion of the main
bolster lying at
least 7 inches higher than the first height.
CA 02797026 2013-11-28
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In another aspect of the invention there is a railroad freight car. The
freight car
has a rail road car body carried on rail road car trucks for rolling motion
along rail road
car tracks. The rail road car body has a deck for carrying lading, and side
sills running
alongside the deck. The body has a first end and a second end. The body has a
coupler
mounted to at least the first end of the body. A main bolster is mounted to
the body
adjacent to the first end of the body longitudinally inboard of the coupler.
The main
bolster has first and second arms extending laterally outboard from a center
plate. The
first rail road car truck has wheels for running along the rail road track.
The arms of the
main bolster has a wheel clearance relief defined therein. The arms of the
bolster has a
first depth of section at the clearance relief and a second depth of section
laterally
outboard of the clearance relief. The second depth of section are greater than
the first
depth of section.
In another aspect of the invention there is a three piece rail road car truck.
The
truck has a first rail car truck having a truck bolster and a pair of first
and second side
frame assemblies. The truck bolster has first and second ends. First and
second spring
groups are mounted between the first and second ends of the truck bolster and
the first
and second side frames respectively. A first set of friction dampers are
mounted between
the truck bolster and the first side frame assembly. A second set of friction
dampers are
mounted between the truck bolster and the second side frame assembly. The
first set of
friction dampers include at least a first friction damper. The first spring
group has at least
a first spring and a second spring. The first friction damper is sprung on the
first and
second springs. The first spring is mounted laterally outboard relative to the
first spring.
Brief Description of the Drawings
Figure la shows a side view of a single unit auto rack rail road car;
Figure lb shows a cross-sectional view of the auto-rack rail road car of
Figure la in
a bi-level configuration, one half section of Figure lb being taken through
the main bolster and the other half taken looking at the cross-tie outboard of
the main bolster;
Figure lc shows a half sectioned partial end view of the rail road car of
Figure la
illustrating the wheel clearance below the main deck, half of the section
being taken through the main bolster, the other half section being taken
outboard of the truck with the main bolster removed for clarity;
CA 02797026 2013-11-28
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Figure id shows a partially sectioned side view of the rail road car of Figure
lc
illustrating the relationship of the truck, the bolster and the wheel
clearance,
below the main deck;
Figure 2a shows a side view of a two unit articulated auto rack rail road car;
Figure 2b shows a side view of an alternate auto rack rail road car to that of
Figure
2a, having a cantilevered articulation;
Figure 3a shows a side view of a three unit auto rack rail road car;
Figure 3b shows a side view of an alternate three unit auto rack rail road car
to the
articulated rail road unit car of Figure 3a, having cantilevered
articulations;
Figure 3c shows an isometric view of an end unit of the three unit auto rack
rail road
car of Figure 3b;
Figure 4a is a partial side sectional view of the draft pocket of the coupler
end of any
of the rail road cars of Figures la, 2a, 2b, 3a, or 3b taken on '4a ¨ 4a' as
indicated in Figure la; and
Figure 4b shows a top view of the draft gear at the coupler end of Figure 4a
taken on
`4b ¨ 4b' of Figure 4a;
Figure 5a shows an isometric view of a three piece truck for the auto rack
rail road
cars of Figures la, 2a, 2b, 3a or 3h;
Figure 5b shows a side view of the three piece truck of Figure 5a;
Figure 5c shows a top view of half of the three piece truck of Figure 5b;
Figure 5d shows a partial section of the three piece truck of Figure 5b taken
on '5d ¨
5d';
Figure 5e shows a partial isometric view of the truck bolster of the three
piece truck
of Figure 5a showing friction damper seats;
Figure 5f shows a force schematic for dampers in the side frame of the truck
of
Figure 5a.
Figure 6a shows a side view of an alternate three piece truck to that of
Figure 5a;
Figure 6b shows a top view of half of the three piece truck of Figure 6a; and
Figure 6c shows a partial section of the three piece truck of Figure 6a taken
on `6c ¨
6c'.
Figure 7 shows a graph of Friction Factor for sliding dampers as a function of
Wedge Angle, in upward and downward motion as an aid to explanation of
the dampers of the truck of Figure 5a;
CA 02797026 2013-11-28
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-
Figure 8a shows an alternate version of the bolster of Figure 5e, with a
double sized
damper pocket for seating a large single wedge having a welded insert;
Figure 8b shows an alternate optional dual wedge for a truck bolster like that
of
Figure 8a;
Figure 8c shows an alternate bolster, similar to that of Figure 5a, having a
pair of
spaced apart wedge pockets, and pocket inserts with both primary and
secondary wedge angles;
Figure 8d shows an alternate bolster, similar to that of Figure 8c, and split
wedges;
Figure 9 shows an optional non-metallic wear surface arrangement for dampers
such
as used in the bolster of Figure 8b;
Figure 10a shows a bolster similar to that of Figure 8c, having a wedge pocket
having primary and secondary angles and a split wedge arrangement for use
therewith;
Figure 10b shows an alternate stepped single wedge for the bolster of Figure
10a;
Figure 10c is a view looking along a plane on the primary angle of the split
wedge of
Figure 10a relative to the bolster pocket;
Figure 10d is a view looking along a plane on the primary angle of the stepped
wedge of Figure 10b relative to the bolster pocket;
Figure 11a shows an alternate bolster and wedge arrangement to that of Figure
8b,
having secondary wedge angles;
Figure llb shows an alternate, split wedge arrangement for the bolster of
Figure
11a;
Figure 11c shows a cross-section of a stepped damper wedge for use with a
bolster
as shown in Figure 11a;
Figure lid shows an alternate stepped damper to that of Figure 11c;
Figure 12a is a section of Figure 5b showing a replaceable side frame wear
plate;
Figure 12b is a sectional view on of the side frame of Figure 12a with the
near end
of the side frame sectioned and the nearer wear plate removed to show the
location of the wear plate of Figure 12a;
Figure 12c shows a compound bolster pocket for the bolster of Figure 12a;
Figure 12d shows a side view detail of the bolster pocket of Figure 12c, as
installed,
relative to the main springs and the wear plate;
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Figure 12e shows an isometric view detail of a split wedge version and a
single
wedge version of wedges for use in the compound bolster pocket of Figure
12c;
Figure 12f shows an alternate, stepped steeper angle profile for the primary
angle of
the wedge of the bolster pocket of Figure 12d;
Figure 12g shows a welded insert having a profile for mating engagement with
the corresponding face of the bolster pocket of Figure 12d;
Figure 13a shows an alternate spring arrangement to that of Figure 12a;
Figure 13b shows mutually inclined springs on section `13b ¨ 13b' of Figure
13a;
Figure 14a shows an exploded isometric view of an alternate bolster and side
frame
assembly to that of Figure 5a, in which horizontally acting springs drive
constant force dampers;
Figure 14b shows a side-by-side double damper arrangement similar to that of
Figure 14a;
Figure 15 shows an isometric view of an alternate spring seat basket for the
truck of
Figure 5a, having a spring insertion access feature.
Figure 16a shows an isometric view of an alternate railroad car truck to that
of
Figure 5a.
Figure 16b shows a side view of the three piece truck of Figure 16a.
Figure 16c shows a top view of the three piece truck of Figure 16a.
Figure 16d shows an end view of the three piece truck of Figure 16a.
Figure 16e shows a schematic of a spring layout for the truck of Figure 16a.
DETAILED DESCRIPTION OF THE INVENTION
The description that follows, and the embodiments described therein, are
provided
by way of illustration of an example, or examples, of particular embodiments
of the
principles of the present invention. These examples are provided for the
purposes of
explanation, and not of limitation, of those principles and of the invention.
In the
description, like parts are marked throughout the specification and the
drawings with the
same respective reference numerals. The drawings are not necessarily to scale
and in
some instances proportions may have been exaggerated in order more clearly to
depict
certain features of the invention.
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In terms of general orientation and directional nomenclature, for each of the
rail
road cars described herein, the longitudinal direction is defined as being
coincident with
the rolling direction of the car, or car unit, when located on tangent (that
is, straight)
track. In the case of a car having a center sill, whether a through center
sill or stub sill,
the longitudinal direction is parallel to the center sill, and parallel to the
side sills, if any.
Unless otherwise noted, vertical, or upward and downward, are terms that use
top of rail,
TOR, as a datum. The term lateral, or laterally outboard, refers to a distance
or
orientation relative to the longitudinal centerline of the railroad car, or
car unit, indicated
as CL - Rail Car. The term "longitudinally inboard", or "longitudinally
outboard" is a
distance taken relative to a mid-span lateral section of the car, or car unit.
Pitching
motion is angular motion of a rail car unit about a horizontal axis
perpendicular to the
longitudinal direction. Yawing is angular motion about a vertical axis. Roll
is angular
motion about the longitudinal axis.
Reference is made in this description to rail car trucks and in particular to
three
piece rail road freight car trucks. Several AAR standard truck sizes are
listed at page 711
in the 1997 Car & Locomotive Cyclopedia. As indicated, for a single unit rail
car having
two trucks, a "40 Ton" truck rating corresponds to a maximum gross car weight
on rail
(GWR) of 142,000 lbs. Similarly, "50 Ton" corresponds to 177,000 lbs, "70 Ton"
corresponds to 220,000 lbs, "100 Ton" corresponds to 263,000 lbs, and "125
Ton"
corresponds to 315,000 lbs. In each case the load limit per truck is then half
the
maximum gross car weight on rail. Two other types of truck are the "110 Ton"
truck for
286,000 LbsGWR and the "70 Ton Special" low profile truck sometimes used for
auto
rack cars. Given that the rail road car trucks described herein tend to have
both
longitudinal and transverse axes of symmetry, a description of one half of an
assembly
may generally also be intended to describe the other half as well, allowing
for differences
between right hand and left hand parts.
Figures la, 2a, 2b, 3a, and 3b, show different types of rail road freight cars
in the
nature of auto rack rail road cars, all sharing a number of similar features.
Figure la
(side view) shows a single unit autorack rail road car, indicated generally as
20. It has a
rail car body 22 supported for rolling motion in the longitudinal direction
(i.e., along the
rails) upon a pair of three-piece rail road freight car trucks 23 and 24
mounted at main
bolsters at either of the first and second ends 26, 28 of rail car body 22.
Body 22 has a
housing structure 30, including a pair of left and right hand sidewall
structures 32, 34 and
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_
an over-spanning canopy, or roof 36 that co-operate to define an enclosed
lading space.
Body 22 has staging in the nature of a main deck 38 running the length of the
car between
first and second ends 26, 28 upon which wheeled vehicles, such as automobiles
can be
conducted by circus-loading. Body 22 can have staging in either a bi-level
configuration,
as shown in Figure lb, in which a second, or upper deck 40 is mounted above
main deck
38 to permit two layers of vehicles to be carried; or a tri-level
configuration with a mid-
level deck, similar to deck 40, and a top deck, also similar to deck 40, are
mounted above
each other, and above main deck 38 to permit three layers of vehicles to be
carried. The
staging, whether hi-level or tri-level, is mounted to the sidewall structures
32, 34. Each
of the decks defines a roadway, trackway, or pathway, by which wheeled
vehicles such as
automobiles can be conducted between the ends of rail road car 20.
A through center sill 50 extends between ends 26, 28. A set of cross-bearers
52
extend to either side of center sill 50, terminating at side sills 56, 58 that
run the length of
car 20 parallel to outer sill 50. Main deck 38 is supported above cross-
bearers 52 and
between side sills 56, 58. Sidewall structures 32, 34 each include an array of
vertical
support members, in the nature of posts 60, that extend between side sills 56,
58, and top
chords 62, 64. A corrugated sheet roof 66 extends between top chords 62 and 64
above
deck 38 and such other decks as employed. Radial arm doors 68, 70 enclose the
end
openings of the car, and are movable to a closed position to inhibit access to
the interior
of car 20, and to an open position to give access to the interior. Each of the
decks has
bridge plate fittings (not shown) to permit bridge plates to be positioned
between car 20
and an adjacent car when doors 68 or 70 are opened to permit circus loading of
the decks.
Both ends of car 20 have couplers and draft gear for connecting to adjacent
rail road cars.
Two ¨ Unit Articulated Auto Rack Car
Similarly, Figure 2a shows a two unit articulated auto rack rail road car,
indicated
generally as 80. It has a first rail car unit body 82, and a second rail car
unit body 85,
both supported for rolling motion in the longitudinal direction (i.e., along
the rails) upon
rail car trucks 84, 86 and 88. Rail car trucks 84 and 88 are mounted at main
bolsters at
respective coupler ends of the first and second rail car unit bodies 83 and
84. Truck 86 is
mounted beneath articulated connector 90 by which bodies 83 and 84 are joined
together.
Each of bodies 83 and 84 has a housing structure 92, 93, including a pair of
left and right
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hand sidewall structures 94, 96 (or 95, 97) and a canopy, or roof 98 (or 99)
that define an
enclosed lading space. A bellows structure 100 links bodies 82 and 83 to
discourage
entry by vandals or thieves.
Each of bodies 82, 83 has staging in the nature of a main deck similar to deck
38
running the length of the car unit between first and second ends 104, 106
(105, 107) upon
which wheeled vehicles, such as automobiles can be conducted. Each of bodies
82, 83
can have staging in either a bi-level configuration, as shown in Figure lb, or
a tri-level
configuration. Other than brake fittings, and other minor fittings, car unit
bodies 82 and
83 are substantially the same, differing in that car body 82 has a pair of
female side-
bearing arms adjacent to articulated connector 90, and car body 83 has a co-
operating
pair of male side bearing arms adjacent to articulated connector 90.
Each of car unit bodies 82 and 83 has a through center sill 110 that extends
between the first and second ends 104, 106 (105, 107). A set of cross-bearers
112, 114
extend to either side of center sill 110, terminating at side sills 116, 118.
Main deck 102
(or 103) is supported above cross-bearers 112, 114 and between side sills 116,
118.
Sidewall structures 94, 96 and 95, 97 each include an array of vertical
support members,
in the nature of posts 120, that extend between side sills 116, 118, and top
chords 126,
128. A corrugated sheet roof 130 extends between top chords 126 and 128 above
deck
102 and such other decks as may be employed.
Radial arm doors 132, 134 enclose the coupler end openings of car bodies 82
and
83 of rail road car 80, and are movable to respective closed positions to
inhibit access to
the interior of rail road car 80, and to respective open positions to give
access to the
interior thereof. Each of the decks has bridge plate fittings (upper deck
fittings not
shown) to permit bridge plates to be positioned between car 80 and an adjacent
auto rack
rail road car when doors 132 or 134 are opened to permit circus loading of the
decks.
For the purposes of this description, the cross-section of Figure lb can be
considered typical also of the general structure of the other railcar unit
bodies described
below, whether 82, 85, 202, 204, 142, 144, 146, 222, 224 or 226. It should be
noted that
Figure lb shows a stepped section in which the right hand portion shows the
main bolster
75 and the left hand section shows a section looking at the cross-tie 77
outboard of the
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main bolster. The sections of Figures lb and lc are typical of the sections of
the end
units described herein at their coupler end trucks, such as trucks 232, 148,
84, 88, 210,
206. The upward recess in the main bolster 75 provides vertical clearance for
the side
frames (typically 7" or more). That is, the clearance 'X' in Figure lc is
about 7 inches in
one embodiment between the side frames and the bolster for an unladen car at
rest.
As may be noted, the web of main bolster 75 has a web rebate 79 and a bottom
flange that has an inner horizontal portion 69, an upwardly stepped horizontal
portion 71
and an outboard portion 73 that deepens to a depth corresponding to the depth
of the
bottom flange of side sill 58. Horizontal portion 69 is carried at a height
corresponding
generally to the height of the bottom flange of side sill 58, and portion 71
is stepped
upwardly relative to the height of the bottom flange of side sill 58 to
provide greater
vertical clearance for the side frame of truck 23 or 24 as the case may be.
Three or More Unit Articulated Auto Rack Car
Figure 3a shows a three unit articulated autorack rail road car, generally as
140.
It has a first end rail car unit body 142, a second end rail car unit body
144, and an
intermediate rail car unit body 146 between rail car unit bodies 142 and 144.
Rail car
unit bodies 142, 144 and 146 are supported for rolling motion in the
longitudinal
direction (i.e., along the rails) upon rail car trucks 148, 150, 152, and 154.
Rail car trucks
148 and 150 are "coupler end" trucks mounted at main bolsters at respective
coupler ends
of the first and second rail car bodies 142 and 144. Trucks 152 and 154 are
"interior" or
"intermediate" trucks mounted beneath respective articulated connectors 156
and 158 by
which bodies 142 and 144 are joined to body 146. For the purposes of this
description,
body 142 is the same as body 82, and body 144 is the same as body 83. Rail car
body
146 has a male end 159 for mating with the female end 160 of body 142, and a
female
end 162 for mating with the male end 164 of rail car body 144.
Body 146 has a housing structure 166 like that of Figure lb, that includes a
pair
of left and right hand sidewall structures 168 and a canopy, or roof 170 that
co-operate to
define an enclosed lading space. Bellows structures 172 and 174 link bodies
142, 146
and 144, 146 respectively to discourage entry by vandals or thieves.
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Body 146 has staging in the nature of a main deck 176, similar to deck 38,
running the length of the car unit between first and second ends 178, 180
defining a
roadway upon which wheeled vehicles, such as automobiles can be conducted.
Body 146
can have staging in either a bi-level configuration or a tri-level
configuration, to co-
operate with the staging of bodies 142 and 144.
Other than brake fittings, and other ancillary features, car bodies 142 and
144 are
substantially the same, differing to the extent that car body 142 has a pair
of female side-
bearing arms adjacent to articulated connector 156, and car body 144 has a co-
operating
pair of male side bearing arms adjacent to articulated connector 158.
Other articulated auto-rack cars of greater length can be assembled by using a
pair
of end units, such as male and female end units 82 and 83, and any number of
intermediate units, such as intermediate unit 146, as may be suitable. In that
sense, rail
road car 140 is representative of multi-unit articulated rail road cars
generally.
Alternate Configurations
Alternate configurations of multi-unit rail road cars are shown in Figures 2b
and
3b. In Figure 2b, a two unit articulated auto-rack rail road car is indicated
generally as
200. It has first and second rail car unit bodies 202, 204 supported for
rolling motion in
the longitudinal direction by three rail road car trucks, 206, 208 and 210
respectively.
Rail car unit bodies 202 and 204 are joined together at an articulated
connector 212. In
this instance, while rail car bodies 202 and 204 share the same basic
structural features of
rail car body 22, in terms of a through center sill, cross-bearers, side
sills, walls and
canopy, and vehicles decks, rail car body 202 is a "two-truck" body, and rail
car body
204 is a single truck body. That is, rail car body 202 has main bolsters at
both its first,
coupler end, and at its second, articulated connector end, the main bolsters
being mounted
over trucks 206 and 208 respectively. By contrast, rail car body 204 has only
a single
main bolster, at its coupler end, mounted over truck 210. Articulated
connector 212 is
mounted to the end of the respective center sills of rail car bodies 202 and
204,
longitudinally outboard of rail car truck 208. The use of a cantilevered
articulation in this
manner, in which the pivot center of the articulated connector is offset from
the nearest
truck center, is described more fully in my co-pending U.S. patent application
09 /
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- 30 -
614,815 for a Rail Road Car with Cantilevered Articulation filed July 12,
2000,
incorporated herein by reference, and may tend to permit a longer car body for
a given
articulated rail road car truck center distance as therein described.
Figure 3b shows a three-unit articulated rail road car 220 having first end
unit
222, second end unit 224, and intermediate unit 226, with cantilevered
articulated
connectors 228 and 230. End units 222 and 224 are single truck units of the
same
construction as car body 204. Intermediate unit 226 is a two truck unit having
similar
construction to car body 202, but having articulated connectors at both ends,
rather than
having a coupler end. Figure 3c shows an isometric view of end unit 224 (or
222).
Analogous five pack articulated rail road cars having cantilevered
articulations can also
be produced. Many alternate configurations of multi-unit articulated rail road
cars
employing cantilevered articulations can be assembled by re-arranging, or
adding to, the
units illustrated.
In each of the foregoing descriptions, each of rail road cars 20, 80, 140, 200
and
220 has a pair of first and second coupler ends by which the rail road car can
be
releasably coupled to other rail road cars, whether those coupler ends are
part of the same
rail car body, or parts of different rail car bodies of a multi-unit rail road
car joined by
articulated connections, draw-bars, or a combination of articulated
connections and draw-
bars.
Figures 4a and 4b show the draft gear at a first coupler end 300 of rail road
car
20, coupler end 300 being representative of either of the coupler ends and
draft gear
arrangement of rail road car 20, and of rail road cars 80, 140, 200 and 220
more
generally. Coupler pocket 302 houses a coupler indicated as 304. It is mounted
to a
coupler yoke 308, joined together by a pin 310. Yoke 308 houses a coupler
follower 312,
a draft gear 314 held in place by a shim (or shims, as required) 316, a wedge
318 and a
filler block 320. Fore and aft draft gear stops 322, 324 are welded inside
coupler pocket
302 to retain draft gear 314, and to transfer the longitudinal buff and draft
loads through
draft gear 314 and on to coupler 304. In the preferred embodiment, coupler 304
is an
AAR Type F7ODE coupler, used in conjunction with an AAR Y45AE coupler yoke and
an AAR Y47 pin. In the preferred embodiment, draft gear 314 is a Mini-BuffGear
such as
manufactured Miner Enterprises Inc., or by the Keystone Railway Equipment
Company,
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- 31 -
of 3420 Simpson Ferry Road, Camp Hill, Pa. As taken together, this draft gear
and
coupler assembly yields a reduced slack, or low slack, short travel, coupling
as compared
to an AAR Type E coupler with standard draft gear or hydraulic EOCC device. As
such
it may tend to reduce overall train slack. In addition to mounting the Mini-
BuffGear
directly to the draft pocket, that is, coupler pocket 302, and hence to the
structure of the
rail car body of rail road car 20, (or of the other rail road cars noted
above) the
construction described and illustrated is free of other long travel draft
gear, sliding sills
and EOCC devices, and the fittings associated with them.
Mini-BuffGear has between 5/8 and 3/4 of an inch displacement travel in buff
at a
compressive force greater than 700,000 Lbs. Other types of draft gear can be
used to
give an official rating travel of less than 2 1/2 inches under M-901-G, or if
not rated, then
a travel of less than 2.5 inches under 500,000 Lbs. buff load. For example,
while Mini-
BuffGear is preferred, other draft gear is available having a travel of less
than 1 3/4 inches
at 400,000 Lbs., one known type has about 1.6 inches of travel at 400,000
Lbs., buff load.
It is even more advantageous for the travel to be less than 1.5 inches at
700,000 Lbs. buff
load and, as in the embodiment of Figures 4a and 4b, preferred that the travel
be at least
as small as 1" inches or less at 700,000 Lbs. buff load.
Similarly, while the AAR Type F7ODE coupler is preferred, other types of
coupler having less than the 25/32" (that is, less than about 3/4") nominal
slack of an
AAR Type E coupler generally or the 20/32" slack of an AAR E5OARE coupler can
be
used. In particular, in alternative embodiments with appropriate housing
changes where
required, AAR Type F79DE and Type F73BE (members of the Type F Family), with
or
without top or bottom shelves; AAR Type CS; or AAR Type H couplers can be used
to
obtain reduced slack relative to AAR Type E couplers.
In each of the autorack rail car embodiments described above, each of the car
units has a weight, that weight being carried by the rail car trucks with
which the car is
equipped. In each of the embodiments of articulated rail cars described above
there is a
number of rail car units joined at a number of articulated connectors, and
carried for
rolling motion along railcar tracks by a number of railcar trucks. In each
case the number
of articulated car units is one more than the number of articulations, and one
less than the
number of trucks. In the event that in alternate embodiements some of the cars
units are
CA 02797026 2013-11-28
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_
joined by draw bars the number of articulated connections will be reduced by
one for
each draw bar added, and the number of trucks will increase by one for each
draw bar
added. Typically articulated rail road cars have only articulated connections
between the
car units. All cars described have releasable couplers mounted at their
opposite ends.
In each embodiment described above, where at least two car units are joined by
an
articulated connector, there are end trucks (e.g. 150, 232) inset from the
coupler ends of
the end car units, and intermediate trucks (e.g. 154, 234) that are mounted
closer to, or
directly under, one or other of the articulated connectors (e.g. 156, 230). In
a car having
cantilevered articulations, such as shown in Figure 2b or 3b, the articulated
connector is
mounted at a longitudinal offset distance (the cantilever arm CA) from the
truck center.
In each case, each of the car units has an empty weight, and also a design
full weight.
The full weight is usually limited by the truck capacity, whether 70 ton, 100
ton, 110 ton
(286,000 lbs.) or 125 ton. In some instances, with low density lading, the
volume of the
lading is such that the truck loading capacity cannot be reached without
exceeding the
volumetric capacity of the car body.
The dead sprung weight of a rail car unit is generally taken as the body
weight of
the car unit, including any ballast, as described below, plus that portion of
the weight of
the truck borne by the springs, (most typically taken as being the weight of
the truck
bolsters). The unsprung weight of the trucks is, primarily, the weight of the
side frames,
the axles and the wheels, plus ancillary items such as the brakes, springs,
and axle
bearings and bearing adapters. The unsprung weight of a three piece truck may
generally
be about 8800 lbs. The live load is the weight of the lading. The sum of (a)
the live
load; (b) the dead sprung load; and (c) the unsprung weight of the trucks is
the gross
railcar weight on rail.
In each of the embodiments described above, each of the rail car units has a
weight and a weight distribution of the dead sprung weight of the carbody
which
determines the dead sprung load carried by each truck. In each of the
embodiments
described above, the sum of the sprung weights of all of the car bodies of an
articulated
car is designated as Wo. (The sprung mass, Mo, is the sprung weight Wo divided
by the
gravitational constant, g. In each case where a weight is given herein, it is
understood
that conversion to mass can be readily made in this way, particularly as when
calculating
natural frequencies). For a single unit, symmetrical rail road car, such as
car 20, the
weight on both trucks is equal. In all of the articulated auto rack rail
road car
CA 02797026 2013-11-28
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embodiments described above, the distributed sprung weight on any end truck,
is at least
2/3, and no more than 4/3 of the nearest adjacent interior truck, such as an
interior truck
next closest to the nearest articulated connector. It is advantageous that the
dead sprung
weight be in the range of 4/5 to 6/5 of the dead sprung weight carried by the
interior
truck, and it is preferred that the dead sprung weight be in the range of 90 %
to 110 % of
the interior truck. It is also desirable that the dead sprung weight on any
truck, Wu's, fall
in the range of 90 % to 110 % of the value obtained by dividing Wo by the
total number
of trucks of the rail road car. Similarly, it is desirable that the dead
sprung weight plus
the live load carried by each of the trucks be roughly similar such that the
overall truck
loading is about the same. In any case, for the embodiments described above,
the design
live load for one truck, such as an end truck, can be taken as being at least
60 % of the
design live load of the next adjacent truck, such as an internal truck. In
terms of overall
dead and live loads, in each of the embodiments described the overall sprung
load of the
end truck is at least 70 % of the nearest adjacent internal truck,
advantageously 80 % or
more, and preferably 90 % of the nearest adjacent internal truck.
Inasmuch as the car weight would generally be more or less evenly distributed
on
a lineal foot basis, and as such the interior trucks would otherwise reach
their load
capacities before the coupler end trucks, weight equalisation may be achieved
in the
embodiments described above by adding ballast to the end car units. That is,
the dead
sprung weight distribution of the end car units is biased toward the coupler
end, and
hence toward the coupler end truck (e.g. 84, 88, 206, 210, 150, 232). For
example, in the
embodiments described above, a first ballast member is provided in the nature
of a main
deck plate 350 of unusual thickness T that forms part of main deck 38 of the
rail car unit.
Plate 350 extends across the width of the end car unit, and from the
longitudinally
outboard end of the deck a distance LB. In the embodiment of Figures 3b and 3c
for
example, the intermediate or interior truck 234 may be a 70 ton truck near its
sprung load
limit of about 101,200 lbs., on the basis of its share of loads from rail car
units 222 and
226 (or, symmetrically 224 and 226 as the case may be), while, without
ballast, end
trucks 232 would be at a significantly smaller sprung load, even when rail car
220 is fully
loaded. In this case, thickness T can be 1 1/2 inches, the width can be 112
inches, and the
length LB can be 312 inches, giving a weight of roughly 15,220 lbs., centered
on the
truck center of end truck 232. This gives a dead load of end car unit 222 of
roughly
77,000 lbs., a dead sprung load on end truck 232 of about 54,000 lbs., and a
total sprung
load on truck 232 can be about 84,000 lbs. By comparison, center car unit 226
has a dead
sprung load of about 60,000 lbs., with a dead sprung load on interior truck
234 of about
CA 02797026 2013-11-28
-34-
55,000 lbs., and yielding a total sprung load on interior truck 234 of 101,000
lbs when car
220 is fully loaded. In this instance as much as a further 17,000 lbs. (+/-)
of additional
ballast can be added before exceeding the "70 Ton" gross weight on rail limit
for the
coupler end truck, 232. Ballast can also be added by increasing the weight of
the lower
flange or webs of the center sill, also advantageously reducing the center of
gravity of the
car.
Similar weight distributions can be made for other capacities of truck whether
100
Ton, 110 Ton or 125 Ton. Although any of these sizes of trucks can be used, it
is
preferable to use a truck with a larger wheel diameter. That is, while 33 inch
wheels (or
even 28" wheels in a 70 Ton Special") can be used, wheels larger than 33
inches in
diameter are preferred such as 36 inch or 38 inch wheels.
Figures 5a, 5b, Sc, 5d and 5e all relate to a three piece truck 400 for use
with the
rail road cars of Figure la, 2a, 2b, 3a or 3b. Figures lc and id show the
relationship of
this truck to the deck level of these rail road cars. Truck 400 has three
major elements,
those elements being a truck bolster 402, symmetrical about the truck
longitudinal
centreline, and a pair of first and second side frames, indicated as 404. Only
one side
frame is shown in Figure 5c given the symmetry of truck 400. Three piece truck
400 has
a resilient suspension (a primary suspension) provided by a spring groups 405
trapped
between each of the distal (i.e., transversely outboard) ends of truck bolster
402 and side
frames 404.
Truck bolster 402 is a rigid, fabricated beam having a first end for engaging
one
side frame assembly and a second end for engaging the other side frame
assembly (both
ends being indicated as 406). A center plate or center bowl 408 is located at
the truck
center. An upper flange 410 extends between the two ends 406, being narrow at
a central
waist and flaring to a wider transversely outboard termination at ends 406.
Truck bolster
402 also has a lower flange 412 and two fabricated webs 414 extending between
upper
flange 410 and lower flange 412 to form an irregular, closed section box beam.
Additional webs 415 are mounted between the distal portions of upper flange
410 and
414 where bolster 402 engages one of the spring groups 405. The transversely
distal
region of truck bolster 402 also has friction damper seats 416, 418 for
accommodating
friction damper wedges as described further below.
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_
Side frame 404 is a casting having bearing seats 419 into which bearing
adapters
420, bearings 421, and a pair of axles 422 mount. Each of axles 422 has a pair
of first
and second wheels 423, 425 mounted to it in a spaced apart position
corresponding to the
width of the track gauge of the track upon which the rail car is to operate.
Side frame 404
also has a compression member, or upper beam member 424, a tension member, or
lower
beam member 426, and vertical side columns 428 and 430, each lying to one side
of a
vertical transverse plane 425 bisecting truck 400 at the longitudinal station
of the truck
center. A generally rectangular opening is defined by the co-operation of the
upper and
lower beam members 424, 426 and vertical columns 428, 430, into which the
distal end
of truck bolster 402 can be introduced. The distal end of truck bolster 402
can then move
up and down relative to the side frame within this opening. Lower beam member
426 has
a bottom or lower spring seat 432 upon which spring group 405 can seat.
Similarly, an
upper spring seat 434 is provided by the underside of the distal portion of
bolster 402 to
engages the upper end of spring group 405. As such, vertical movement of truck
bolster
402 will tend to compress or release the springs in spring group 405.
In the embodiment of Figure 5a, spring group 405 has two rows of springs 436,
a
transversely inboard row and a transversely outboard row, each row having four
large (8
inch +/-) diameter coil springs giving vertical bounce spring rate constant,
k, for group
405 of less than 10,000 lbs / inch. This spring rate constant can be in the
range of 6000 to
10,000 lbs / in., and is advantageously in the range of 7000 to 9500 lbs / in,
giving an
overall vertical bounce spring rate for the truck of double these values,
preferably in the
range of 14000 to 18,500 lbs / in for the truck. The spring array can include
nested coils
of outer springs, inner springs, and inner-inner springs depending on the
overall spring
rate desired for the group, and the apportionment of that stiffness. The
number of
springs, the number of inner and outer coils, and the spring rate of the
various springs can
be varied. The spring rates of the coils of the spring group add to give the
spring rate
constant of the group, typically being suited for the loading for which the
truck is
designed.
Each side frame assembly also has four friction damper wedges arranged in
first
and second pairs of transversely inboard and transversely outboard wedges 440,
441, 442
and 443 that engage the sockets, or seats 416, 418 in a four-cornered
arrangement. The
corner springs in spring group 405 bear upon a friction damper wedge 440, 441,
442 or
443. Each of vertical columns 428, 430 has a friction wear plate 450 having
transversely
inboard and transversely outboard regions against which the friction faces of
wedges 440,
CA 02797026 2013-11-28
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441, 442 and 443 can bear, respectively. Bolster gibs 451 and 453 lie inboard
and
outboard of wear plate 450 respectively. Gibs 451 and 453 act to limit the
lateral travel
of bolster 402 relative to side frame 404. The deadweight compression of the
springs
under the dampers will tend to yield a reaction force working on the bottom
face of the
wedge, trying to drive the wedge upward along the inclined face of the seat in
the bolster,
thus urging, or biasing, the friction face against the opposing portion of the
friction face
of the side frame column. In one embodiment, the springs chosen can have an
undeflected length of 15 inches, and a dead weight deflection of about 3
inches.
As seen in the top view of Figure Sc, and in the schematic sketch of Figure 5f
the
side-by-side friction dampers have a relatively wide averaged moment arm L to
resist
angular deflection of the side frame relative to the truck bolster in the
parallelogram
mode. This moment arm is significantly greater than the effective moment arm
of a
single wedge located on the spring group (and side frame) centre line.
Further, the use of
independent springs under each of the wedges means that whichever wedge is
jammed in
tightly, there is always a dedicated spring under that specific wedge to
resist the
deflection. In contrast to older designs, the overall damping face width is
greater because
it is sized to be driven by relatively larger diameter (e.g., 8 in +/-)
springs, as compared to
the smaller diameter of, for example, AAR B 432 out or B 331 side springs, or
smaller.
Further, in having two elements side-by-side the effective width of the damper
is
doubled, and the effective moment arm over which the diagonally opposite
dampers work
to resist parallelogram deformation of the truck in hunting and curving
greater than it
would have been for a single damper.
In the illustration of Figure 5e, the damper seats are shown as being
segregated by
a partition 452. If a longitudinal vertical plane 454 is drawn through truck
400 through
the center of partition 452, it can be seen that the inboard dampers lie to
one side of plane
454, and the outboard dampers lie to the outboard side of plane 454. In
hunting then, the
normal force from the damper working against the hunting will tend to act in a
couple in
which the force on the friction bearing surface of the inboard pad will always
be fully
inboard of plane 454 on one end, and fully outboard on the other diagonal
friction face.
For the purposes of conceptual visualisation, the normal force on the friction
face of any
of the dampers can be idealised as an evenly distributed pressure field whose
effect can
be approximated by a point load whose magnitude is equal to the integrated
value of the
pressure field over its area, and that acts at the centroid of the pressure
field. The center
of this distributed force, acting on the inboard friction face of wedge 440
against column
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428 can be thought of as a point load offset transversely relative to the
diagonally
outboard friction face of wedge 443 against column 430 by a distance that is
notionally
twice dimension L' shown in the conceptual sketch of Figure 5f. In the
example, this
distance is about one full diameter of the large spring coils in the spring
set. It is a
significantly greater effective moment arm distance than found in typical
friction damper
wedge arrangements. The restoring moment in such a case would be,
conceptually, MR =
[(F1 + F3) - (F2 + FLOW. As indicated by the formulae on the conceptual sketch
of Figure
5f, the difference between the inboard and outboard forces on each side of the
bolster is
proportional to the angle of deflection c of the truck bolster relative to the
side frame, and
since the normal forces due to static deflection xo may tend to cancel out, MR
=
4k,Tan(c)Tan(0)L, where 0 is the primary angle of the damper, and k, is the
vertical
spring constant of the coil upon which the damper sits and is biased.
Further, in typical friction damper wedges, the enclosed angle of the wedge
tends
to be somewhat less than 35 degrees measured from the vertical face to the
sloped face
against the bolster. As the wedge angle decreases toward 30 degrees, the
tendency of the
wedge to jam in place increases. Conventionally the wedge is driven by a
single spring
in a large group. The portion of the vertical spring force acting on the
damper wedges
can be less than 15 % of the group total. In the embodiment of Figure 5b, it
is 50 % of
the group total (i.e., 4 of 8 equal springs). The wedge angle of wedges 440,
442 is
significantly greater than 35 degrees. With reference again to Figure 7, the
use of more
springs, or more precisely a greater portion of the overall spring stiffness,
under the
dampers, permits the enclosed angle of the wedge to be over 35 degrees, and
advantageously larger, in the range of between roughly 37 to 40 or 45 degrees
to roughly
60 or 65 degrees.
Where a softer suspension is used employing a relatively small number of large
diameter springs, such as in a 2 x 4, 3 x 3, or 3 x 5 group as described in
the detailed
description of the invention herein, dampers may be mounted over each of four
corner
positions. In that case, the portion of spring force acting under the damper
wedges may
be in the 25 ¨ 50 % range for springs of equal stiffness. If the coils or coil
groups are not
of equal stiffness, the portion of spring force acting under the dampers may
be in the
range of perhaps 20 % to 70 %. The coil groups can be of unequal stiffness if
inner coils
are used in some springs and not in others, or if springs of differing spring
constant are
used.
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,
The size of the spring group embodiment of Figure 5b yields a side frame
window
opening having a width between the vertical columns of side frame 404 of
roughly 33
inches. This is relatively large compared to existing spring groups, being
more than 25 %
greater in width. Truck 400 has a correspondingly greater wheelbase length,
indicated as
WB. WB is advantageously greater than 73 inches, or, taken as a ratio to the
track gauge
width, is advantageously greater than 1.30 time the track gauge width. It is
preferably
greater than 80 inches, or more than 1.4 times the gauge width, and in one
embodiment is
greater than 1.5 times the track gauge width, being as great, or greater than,
about 86
inches. Similarly, the side frame window is advantageously wider than tall,
the
measurement across the wear plate faces of the side frame columns being
advantageously
greater than 24", possibly in the ratio of greater than 8:7 of width to
height, and possibly
in the range of 28" or 32" or more, giving ratios of greater than 4:3 and
greater than 3:2.
The spring seat may have lengthened dimensions to correspond to the width of
the side
frame window, and a transverse width of 15 'A - 17" or more.
In Figures 6a, 6b and 6c, there is an alternate embodiment of soft spring
rate, long
wheelbase three piece truck, identified as 460. Truck 460 employs constant
force inboard
and outboard, fore and aft pairs of friction dampers 466 mounted in the distal
ends of
truck bolster 468. In this arrangement, springs 470 are mounted horizontally
in pockets
in the distal ends of truck bolster 468 and urge, or bias, each of the
friction dampers 466
against the corresponding friction surfaces of the vertical columns of the
side frames.
The spring force on friction damper wedges 440, 441, 442 and 443 varies as a
function of the vertical displacement of truck bolster 402, since they are
driven by the
vertical springs of spring group 405. By contrast, the deflection of springs
470 does not
depend on vertical compression of the main spring group 472, but rather is a
function of
an initial pre-load. Although the arrangement of Figure 6a, 6b and 6c still
provides
inboard and outboard dampers and independent springing of the dampers, the
embodiment of Figures 5b is preferred.
In the embodiments of Figures la, lb, 2a, 2b, 3a and 3b, the ratio of the dead
sprung weight, WD, of the rail car unit (being the weight of the car body plus
the weight
of the truck bolster) without lading to the live load, WL, namely the maximum
weight of
lading, be at least 1:1. It is advantageous that this ratio WD : WL lie in the
range of 1:1
to 10:3. In one embodiment of rail car of Figures la, lb, 2a, 2b, 3a and 3b
the ratio can
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be about 1.2 : 1 It is more advantageous for the ratio to be at least 1.5 : 1,
and preferable
that the ratio be greater than 2 : 1.
Figures 8a and 8b
Figures 8a and 8b show a partial isometric view of a truck bolster 480 that is
generally similar to truck bolster 400 of Figure 5d, except insofar as bolster
pocket 482
does not have a central partition like web 452, but rather has a continuous
bay extending
across the width of the underlying spring group, such as spring group 436. A
single wide
damper wedge is indicated as 484. Damper 484 is of a width to be supported by,
and to
be acted upon, by two springs 486, 488 of the underlying spring group. In the
event that
bolster 400 may tend to deflect to a non-perpendicular orientation relative to
the
associated side frame, as in the parallelogamming phenomenon, one side of
wedge 484
will tend to be squeezed more tightly than the other, giving wedge 484 a
tendency to
twist in the pocket about an axis of rotation perpendicular to the angled face
(i.e., the
hypotenuse face) of the wedge. This twisting tendency may also tend to cause
differential compression in springs 486, 488, yielding a restoring moment both
to the
twisting of wedge 484 and to the non-square displacement of truck bolster 480
relative to
the truck side frame. As there may tend to be a similar moment generated at
the opposite
spring pair at the opposite side column of the side frame, this may tend to
enhance the
self-squaring tendency of the truck more generally.
Also included in Figure 8b is an alternate pair of damper wedges 490, 492.
This
dual wedge configuration can similarly seat in bolster pocket 482, and, in
this case, each
wedge 490, 492 sits over a separate spring. Wedges 490, 492 are vertically
slidable
relative to each other along the primary angle of the face of bolster pocket
482. When the
truck move to an out of square condition, differential displacement of wedges
490, 492
may tend to result in differential compression of their associated springs,
e.g., 486, 488
resulting in a restoring moment as above.
The sliding motion described above may tend to cause wear on the moving
surfaces, namely (a) the side frame columns, and (b) the angled surfaces of
the bolster
pockets. To alleviate, or ameliorate, this situation, consumable wear plates
494 can be
mounted in bolster pocket 482 (with appropriate dimensional adjustments) as in
Figure
8b. Wear plates 494 can be smooth steel plates, possibly of a hardened, wear
resistant
alloy, or can be made from a non-metallic, or partially non-metallic,
relatively low
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,
friction wear resistant surface. Other plates for engaging the friction
surfaces of the
dampers can be mounted to the side frame columns, and indicated by item 496 in
Figure
14a.
For the purposes of this example, it has been assumed that the spring group is
two
coils wide, and that the pocket is, correspondingly, also two coils wide. The
spring group
could be more than two coils wide. The bolster pocket is assumed to have the
same
width as the spring group, but could be less wide. For two coils where in some
embodiments the group may be more than two coils wide. A symmetrical
arrangement of
the dampers relative to the side frame and the spring group is desirable, but
an
asymmetric arrangement could be made. In the embodiments of Figures 5a, 8a and
16a,
the dampers are in four cornered arrangements that are symmetrical both about
the center
axis of the truck bolster and about a longitudinal vertical plane of the side
frame.
Similarly, the wedges themselves can be made from a relatively common
material, such as a mild steel, and the given consumable wear face members in
the nature
of shoes, or wear members. Such an arrangement is shown in Figure 9 in which a
damper wedge is shown generically as 500. The replaceable, consumable wear
members
are indicated as 502, 504. The wedges and wear members have mating male and
female
mechanical interlink features, such as the cross-shaped relief 503 formed in
the primary
angled and vertical faces of wedge 500 for mating with the corresponding
raised cross
shaped features 505 of wear members 502, 504. Sliding wear member 502 is
preferably
made of a non-metallic, low friction material.
Although Figure 9 shows a consumable insert in the nature of a wear plate, the
entire bolster pocket can be made as a replaceable part, as in Figure 8a. This
bolster
pocket can be made of a high precision casting, or can be a sintered powder
metal
assembly having desired physical properties. The part so formed is then welded
into
place in the end of the bolster, as at 506 indicated in Figure 8a.
The underside of the wedges described herein, wedge 500 being typical in this
regard, has a seat, or socket 507, for engaging the top end of the spring
coil, whichever
spring it may be, spring 562 being shown as typically representative. Socket
507 serves
to discourage the top end of the spring from wandering away from the intended
generally
central position under the wedge. A bottom seat, or boss for discouraging
lateral
wandering of the bottom end of the spring is shown in Figure 14a as item 508.
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Thus far only primary angles have been discussed. Figure 8c shows an isometric
view of an end portion of a truck bolster 510, generally similar to bolster
400. As with all
of the truck bolsters shown and discussed herein, bolster 510 is symmetrical
about the
longitudinal vertical plane of the bolster (i.e., cross-wise relative to the
truck generally)
and symmetrical about the vertical mid-span section of the bolster (i.e., the
longitudinal
plane of symmetry of the truck generally, coinciding with the rail car
longitudinal center
line). Bolster 510 has a pair of spaced apart bolster pockets 512, 514 for
receiving
damper wedges 516, 518. Pocket 512 is laterally inboard of pocket 514 relative
to the
side frame of the truck more generally. Consumable wear plate inserts 520, 522
are
mounted in pockets 512, 514 along the angled wedge face.
As can be seen, wedges 516, 518 have a primary angle, a as measured between
vertical sliding face 524, (or 526, as may be) and the angled vertex 528 of
outboard face
530. For the embodiments discussed herein, primary angle a will tend to be
greater than
40 degrees, and may typically lie in the range of 45 ¨ 65 degrees, possibly
about 55 - 60
degrees. This angle will be common to the slope of all points on the sliding
hypotenuse
face of wedge 516 (or 518) when taken in any plane parallel to the plane of
outboard end
face 530. This same angle a is matched by the facing surface of the bolster
pocket, be it
512 or 514, and it defines the angle upon which displacement of wedge 516, (or
518) is
intended to move relative to that surface.
A secondary angle D gives the inboard, (or outboard), rake of the hypotenuse
surface of wedge 516 (or 518). The true rake angle can be seen by sighting
along plane
of the hypotenuse face and measuring the angle between the hypotenuse face and
the
planar outboard face 530. The rake angle is the complement of the angle so
measured.
The rake angle may tend to be greater than 5 degrees, may lie in the range of
10 to 20
degrees, and is preferably about 15 degrees. A modest angle is desirable.
When the truck suspension works in response to track perturbations, the damper
wedges may tend to work in their pockets. The rake angles yield a component of
force
tending to bias the outboard face 530 of outboard wedge 518 outboard against
the
opposing outboard face of bolster pocket 514. Similarly, the inboard face of
wedge 516
will tend to be biased toward the inboard planar face of inboard bolster
pocket 512.
These inboard and outboard faces of the bolster pockets are preferably lined
with a low
friction surface pad, indicated generally as 532. The left hand and right hand
biases of
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,
the wedges may tend to keep them apart to yield the full moment arm distance
intended,
and, by keeping them against the planar facing walls, may tend to discourage
twisting of
the dampers in the respective pockets.
Bolster 510 includes a middle land 534 between pockets 512, 514, against which
another spring 536 may work, such as might be found in a spring group that is
three (or
more) coils wide. However, whether two, three, or more coils wide, and whether
employing a central land or no central land, bolster pockets can have both
primary and
secondary angles as illustrated in the example embodiment of Figure 8c, with
or without
(though preferably with) wear inserts.
In the case where a central land, such as land 534 separates two damper
pockets,
the opposing wear plates of the side frame columns need not be monolithic.
That is, two
wear plate regions could be provided, one opposite each of the inboard and
outboard
dampers, presenting planar surfaces against which those dampers can bear.
Advantageously, the normal vectors of those regions are parallel, and most
conveniently
those surfaces are co-planar and perpendicular to the long axis of the side
frame, and
present a clear, un-interrupted surface to the friction faces of the dampers.
The examples of Figures 8a, 8b and 8c are arranged in order of incremental
increases in complexity. The Example of Figure 8d again provides a further
incremental
increase in complexity. Figure 8d shows a bolster 540 that is similar to
bolster 510
except insofar as bolster pockets 542, 544 each accommodate a pair of split
wedges 546,
548. Pockets 542, 544 each have a pair of bearing surfaces 550, 552 that are
inclined at
both a primary angle and a secondary angle, the secondary angles of surfaces
550 and
552 being of opposite hand to yield the damper separating forces discussed
above.
Surfaces 550 and 552 are also provided with linings in the nature of
relatively low
friction wear plates 554, 556. Each of pockets 542 and 544 accommodates a pair
of split
wedges 558, 560. Each pair of split wedges seats over a single spring 562.
Another
spring 564 bears against central land 566.
The example of Figure 10a shows a combination of a bolster 570 and biased
split
wedges 572, 574. Bolster 570 is the same as bolster 540 except insofar as
bolster pockets
576, 578 are stepped pockets in which the steps, e.g., items 580, 582, have
the same
primary angle, and the same secondary angle, and are both biased in the same
direction,
unlike the symmetrical sliding faces of the split wedges in Figure 8d, which
are left and
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right handed. Thus the outboard pair of split wedges 584 has a first member
586 and a
second member 588 each having primary angle a and secondary angle 13, and are
of the
same hand such that in use both the first and second members will tend to be
biased in the
outboard direction (i.e. toward the distal end of bolster 570). Similarly, the
inboard pair
of split wedges 590 has a first member 592 and a second member 594 each having
primary angle a, and secondary angle fi, except that the sense of secondary
angle (3 is in
the opposite direction such that members 592 and 592 will tend in use to be
driven in the
inboard direction (i.e., toward the truck center).
As shown in the partial sectional view of Figure 10c, a replaceable monolithic
stepped wear insert 596 is welded in the bolster pocket 580 (or 582 if
opposite hand, as
the case may be). Insert 596 has the same primary and secondary angles a and
fl as the
split wedges it is to accommodate, namely 586, 588 (or, opposite hand, 592,
594). When
installed, and working, the more outboard of the wedges, 588 (or, opposite
hand, the
more inboard of the wedges 592) has a vertical and longitudinally planar
outboard face
600 that bears against a similarly planar outboard face 602 (or, opposite
hand, inboard
face 604) These faces are preferably prepared in a manner that yields a
relatively low
friction sliding interface between them. In that regard, a low friction pad
may be
mounted to either surface, preferably the outboard surface of pocket 580. The
hypotenuse face 606 of member 588 bears against the opposing outboard land 610
of
insert 596. The overall width of outboard member 588 is greater than that of
outboard
land 610, such that the inboard planar face of member 588 acts as an abutment
face to
fend inboard member 586 off of the surface of the step 612 in insert 596.
In similar manner inboard wedge member 586 has a hypotenuse face 614 that
bears against the inboard land portion 616 of insert 596. The total width of
bolster pocket
580 is greater than the combined width of wedge members, such that a gap is
provided
between the inboard (non-contacting) face of member 586 and the inboard planar
face of
pocket 580. The same relationship, but of opposite hand, exists between pocket
582 and
members 592, 594.
In an optional embodiment, a low friction pad, or surfacing, can be used at
the
interface of members 586, 588 (or 592, 594) to facilitate sliding motion of
the one
relative to the other.
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In this arrangement, working of the wedges, i.e., members 586, 588 against the
face of insert 596 will tend to cause both members to move in one direction,
namely to
their most outboard position. Similarly, members 592 and 594 will work to
their most
inboard positions. This may tend to maintain the wedge members in an untwisted
orientation, and may also tend to maintain the moment arm of the restoring
moment at its
largest value, both being desirable results.
When a twisting moment of the bolster relative to the side frames is
experienced,
as in parallelogram deformation, all four sets of wedges will tend to work
against it. That
is, the diagonally opposite pairs of wedges in the outboard pocket of one side
of the
bolster and on the inboard pocket on the other side will be compressed, and
the opposite
side will be, relatively, relieved, such that a differential force will exist.
The differential
force will work on a moment arm roughly equal to the distance between the
centers of the
inboard and outboard pockets, or slightly more given the gap arrangement.
In the further alternative arrangement of Figures 10b and 10d, a single,
stepped
wedge 620 is used in place of the pair of split wedges e.g., members 586, 588.
A
corresponding wedge of opposite hand is used in the other bolster pocket.
In the further alternative embodiment of Figures 11a, a truck bolster 630 has
welded bolster pocket inserts 632 and 634 of opposite hands welded into
accommodations in its distal end. In this instance, each bolster pocket has an
inboard
portion 636 and an outboard portion 638. Inboard and outboard portions 636 and
638
share the same primary angle a, but have secondary angles p that are of
opposite hand.
Respective inboard and outboard wedges are indicated as 640 and 642, and each
seats
over a vertically oriented spring 644, 646. In this case bolster 630 is
similar to bolster
480 of Figure 8a, to the extent that the bolster pocket is continuous ¨ there
is no land
separating the inner and outer portions of the bolster pocket. Bolster 630 is
also similar
to bolster 510 of Figure 8c, except that rather than the bolster pockets of
opposite hand
being separated, they are merged without an intervening land.
In the further alternative of Figure 11b, split wedge pairs 648, 650 (inboard)
and
652, 654 (outboard) are employed in place of the single inboard and outboard
wedges
640 and 642.
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In some instances the primary angle of the wedge may be steep enough that the
thickness of section over the spring might not be overly great. In such a
circumstance the
wedge may be stepped in cross section to yield the desired thickness of
section as show in
the details of Figures 11c and 11d.
Figure 12a shows the placement of a low friction bearing pad for bolster 480
of
Figure 8a. it will be appreciated that such a pad can be used at the interface
between the
friction damper wedges of any of the embodiments discussed herein. In Figure
12a, the
truck bolster is identified as item 660 and the side frame is identified as
item 662. Side
frame 662 is symmetrical about the truck centerline, indicated as 664. Side
frame 662
has side frame columns 668 that locate between the inner and outer gibs 670,
672 of truck
bolster 660. The spring group is indicated generally as 674, and has eight
relatively large
diameter springs arranged in two rows, being an inboard row and an outboard
row. Each
row has four springs in it. The four central springs 676, 677, 678, 679 seat
directly under
the bolster end 680. The end springs of each row, 681, 682, 683, 684 seat
under
respective friction damper wedges 685, 686, 687, 688. Consumable wear plates
689, 690
are mounted to the wide, facing flanges 691, 692 of the side frame columns,
688. As
shown in Figure 12b, plates 689, 690 are mounted centrally relative to the
side frames,
beneath the juncture of the side frame arch 692 with the side frame columns.
The lower
longitudinal member of the side frame, bearing the spring seat, is indicated
as 694.
Referring now to Figure 12c and 12e, bolster 660 has a pair of left and right
hand,
welded-in bolster pocket assemblies 700, 702, each having a cast steel,
replaceable,
welded-in wedge pocket insert 704. Insert 704 has an inboard-biased portion
706, and an
outboard-biased portion 708. Inboard end spring 682 (or 681) bears against an
inboard-
biased split wedge pair 710 having members 712, 714, and outboard end spring
684 (or
683) bears against an outboard- biased split wedge pair 716 having members
718, 720.
As suggested by the names, the outboard-biased wedges will tend to seat in an
outboard
position as the suspension works, and the inboard-biased wedges will tend to
seat in an
inboard position.
Each insert portion 706, 708 is split into a first part and a second part for
engaging, respectively, the first and second members of a commonly biased
split wedge
pair. Considering pair 710, inboard leading member 712 has an inboard planar
face 724,
that, in use, is intended slidingly to contact the opposed vertically planar
face of the
bolster pocket. Leading member 712 has a bearing face 726 having primary angle
a and
CA 02797026 2013-11-28
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,
secondary angle D. Trailing member 714 has a bearing face 728 also having
primary
angle a and secondary angle D, and, in addition, has a transition, or step,
face 730 that has
a primary angle a and a tertiary angle cp.
Insert 704 has a corresponding an array of bearing surfaces having a primary
angle a, and a secondary angle 11, with transition surfaces having tertiary
angle q for
mating engagement with the corresponding surfaces of the inboard and outboard
split
wedge members. As can be seen, a section taken through the bearing surface
resembles a
chevron with two unequal wings in which the face of the secondary angle 13 is
relatively
broad and shallow and the face associated with tertiary angle q is relatively
narrow and
steep.
In Figure 12e, it can be seen that the sloped portions of split wedge members
718,
720 extend only partially far enough to overlie a coil spring 726. In
consequence, wedge
members 718 and 720 each have a base portion 728, 730 having a fore-and-aft
dimension
greater than the diameter of spring 726, and a width greater than half the
diameter of
spring 726. Each of base portions 728, 730 has a downwardly proud, roughly
semi-
circular boss 732 for seating in the top of the coil of spring 726. The
upwardly angled
portion 734, 736 of each wedge member 718, 720 is extends upwardly of base
portion
728, 730 to engage the matingly angled portions of insert 704.
In a further alternate embodiment, the split wedges can be replaced with
stepped
wedges 740 of similar compound profile, as shown In Figure 12f. In the event
that the
primary wedge angle is relatively steep (i.e., greater than about 45 degrees
when
measured from the horizontal, or less than about 45 degrees when measured from
the
vertical). Figure 12g shows a welded in insert 742 having a profile for mating
engagement with the corresponding wedge faces.
Figures 13a and 13b illustrate a further alternate embodiment, being generally
similar to the 2 x 4 spring layout of the embodiment of Figure 8a. However, in
this
example, while the damper arrangement is as in Figure 8a, the central four
springs 744,
745, 746, and 747 are installed in inboard and outboard pairs in spring seats
in which the
springs do not act on a vertical line (assuming no lateral translation of
bolster 748 relative
to side frame 750), but rather are splayed to act on a dihedral angle from the
vertical,
this splayed inclination tending to urge bolster 748 to a centered neutral
position of lateral
CA 02797026 2013-11-28
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translation relative to side frame 750. The angle of splay is relatively
modest, being in
the range of 0 to 10 degrees from the vertical, and may be about 5 degrees.
Figures 14a and 14b illustrate a bolster, side frame and damper arrangement in
which dampers 760, 761 are independently sprung on horizontally acting springs
762,
763 housed in side-by-side pockets 764, 765 in the distal end of bolster 770.
Although
only two dampers are shown, it will be understood that a pair of dampers faces
toward
each of the opposed side frame columns. Dampers 760, 761 each include a block
768
and a consumable wear member 772, the block and wear member having male and
female indexing features 774 to maintaining their relative position. An
arrangement of
this nature permits the damper force to be independent of the compression of
the springs
in the main spring group. A removable grub screw fitting 778 is provided in
the spring
housing to permit the spring to be pre-loaded and held in place during
installation.
Figure 15 shows a bottom spring seat 780 for a side frame 782. Bottom Spring
seat 780 has a base portion 784 upon which to rest the springs of a spring
group, such as
those described above, and includes an upstanding peripheral retaining wall,
786.
Retaining wall 786 has an opening, or gate 788 to permit springs to slide into
place from
the outside. The last spring slid in during installation, or the first spring
out during
removal, seats in a depression, or relief, or seat, 790, and is thereby
discouraged from
moving out through gate 788 while in operation.
Figures 16a, 16b and 16c show a preferred truck 800, having a bolster 802, a
side
frame 804, a spring group 806, and a damper arrangement 808. The spring group
has a 5
x 3 arrangement, with the dampers being in a spaced arrangement generally as
shown in
Figure 8c, and having a primary damper angle that may tend to be somewhat
sharper
given the smaller proportion of the total spring group that works under the
dampers (i.e.,
4 / 15 as opposed to 4 / 9 in Figure 8c).
The embodiments described have natural vertical bounce frequencies that are
less
than the 4 ¨ 6 Hz. range of freight cars more generally. In addition, a
softening of the
suspension to 3.0 hz would be an improvement, yet the embodiments described
herein,
whether for individual trucks or for overall car response can employ
suspensions giving
less than 3.0 Hz in the unladen vertical bounce mode. That is, the fully laden
natural
vertical bounce frequency for one embodiment of rail cars of Figures la, lb,
2a, 2b, 3a
and 3b is 1.5 Hz or less, with the unladen vertical bounce natural frequency
being less
CA 02797026 2013-11-28
- 48 -
than 2.0 Hz, and advantageously less than 1.8 Hz. It is preferred that the
natural vertical
bounce frequency be in the range of 1.0 Hz to 1.5 Hz. The ratio of the unladen
natural
frequency to the fully laden natural frequency is less than 1.4 : 1.0,
advantageously less
than 1.3 : 1.0, and even more advantageously, less than 1.25 : 1Ø
In the embodiments described above, it is preferred that the spring group be
installed without the requirement for pre-compression of the springs. However,
where a
higher ratio of dead sprung weight to live load is desired, additional ballast
can be added
up to the limit of the truck capacity with appropriate pre-compression of the
springs. It is
advantageous for the spring rate of the spring groups be in the range of 6,400
to 10,000
lbs/in per side frame group, or 12,000 to 20,000 lbs/in per truck in vertical
bounce.
In the embodiments of Figures 5a, 8a, and 16a, the gibs are shown mounted to
the
bolster inboard and outboard of the wear plates on the side frame columns. In
the
embodiments shown herein, the clearance between the gibs and the side plates
is
desirably sufficient to permit a motion allowance of at least 3/4" of lateral
travel of the
truck bolster relative to the wheels to either side of neutral, advantageously
permits
greater than 1 inch of travel to either side of neutral, and more preferably
permits travel
in the range of about 1 or 1 ¨ 1/8" to about 1 ¨ 5/8 or 1 ¨ 9/16" inches to
either side of
neutral, and in one embodiment against either the inboard or outboard stop.
In a related feature, in the embodiments of Figures 5a, 8a and 16a, the side
frame
is mounted on bearing adapters such that the side frame can swing transversely
relative to
the wheels. While the rocker geometry may vary, the side frames shown, by
themselves,
have a natural frequency when swinging of less than about 1.4 Hz, and
preferably less
than 1 Hz, and advantageously about 0.6 to 0.9 Hz. Advantageously, when
combined
with the lateral spring stiffness of a spring group in shear, the overall
lateral natural
frequency of the truck suspension, for an unladen car, may tend to be less
than 1 Hz for
small deflections, and preferably less than 0.9 Hz.
The most preferred embodiments of this invention combine a four cornered
damper arrangement with spring groups having a relatively low vertical spring
rate, and a
relatively soft response to lateral perturbations. This may tend to give
enhanced
resistance to hunting, and relatively low vertical and transverse force
transmissibility
through the suspension such as may give better overall ride quality for high
value low
CA 02797026 2013-11-28
- 49 -
density lading, such as automobiles, consumer electronic goods, or other
household
appliances, and for fresh fruit and vegetables.
While the most preferred embodiments combine these features, they need not all
be present at one time, and various optional combinations can be made. As
such, the
features of the embodiments of the various figures may be mixed and matched,
without
departing from the spirit or scope of the invention. For the purpose of
avoiding redundant
description, it will be understood that the various damper configurations can
be used with
spring groups of a 2 X 4, 3 X 3, 3:2:3, 3 X 5 or other arrangement. Similarly,
although
the discussion involves trucks for rail road cars for carrying low density
lading, it applies
to trucks for carrying relatively fragile high density lading such as rolls of
paper, for
example, where ride quality is an important consideration. Further, while the
improved
ride quality features of the damper and spring sets are most preferably
combined with a
low slack, short travel, set of draft gear, for use in a "No Hump" car, these
features can be
used in cars having conventional slack and longer travel draft gear.
The principles of the present invention are not limited to auto rack rail road
cars,
but apply to freight cars, more generally, including cars for paper, auto
parts, household
appliances and electronics, shipping containers, and refrigerator cars for
fruit and
vegetables. More generally, they apply to three piece freight car trucks in
situations
where improved ride quality is desired, typically those involving the
transport of
relatively high value, low density manufactured goods.
Various embodiments of the invention have now been described in detail. Since
changes in and or additions to the above-described best mode may be made
without
departing from the nature, spirit or scope of the invention, the invention is
not to be limited
to those details, but only according to a purposive construction of the claims
as required by
law.
