Note: Descriptions are shown in the official language in which they were submitted.
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RAIL ROAD CAR AND TRUCK THEREFOR
FIELD OF THE INVENTION
This invention relates generally to rail road freight cars and to trucks for
use with rail
road freight cars.
BACKGROUND OF THE INVENTION
Autorack rail road cars are used to transport automobiles. Typically, autorack
rail
road cars are loaded in the "circus loading" manner, by driving vehicles into
the cars
from one end, and securing them in place with chocks, chains or straps. When
the trip is
completed, the chocks are removed, and the cars are driven out. The
development of
autorack rail road cars can be traced back 80 or 90 years, when mass
production led to a
need to transport large numbers of automobiles from the factory to market.
Automobiles are a high value, relatively low density, relatively fragile type
of
lading. Damage to lading due to dynamic loading in the railcar may tend to
arise
principally in two ways. First, there are 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 railcar 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 1/4 to 3
1/4 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.
The subject of slack action is discussed at length in my co-pending US patent
application 09/920,437 filed August 1, 2001 and now issued as US P 6,659,016.
Since automobiles tend to be a relatively low density form of lading as
compared to
grain, ores, or coal, the volumetric capacity of the cars tends to be filled
up before the weight
reaches the maximum allowable weight for the trucks. This has led to efforts
to increase the
volumetric capacity of the cars. Over time, particularly in the period of 1945
¨ 1970,
autorack cars grew longer and taller. At present, an autorack car may be up to
about 90 feet
long and 20 ft ¨ 2 inches tall. Autorack cars may typically have a tall,
somewhat barn-like
housing. The housing has end doors that are intended to keep out thieves and
vandals.
The desire to increase the internal volume of the autorack car, and the
relatively light
weight of the lading, led to the development of a special 70 Ton rail road car
truck for use
with autorack cars. A 70 Ton "special" truck is shown in the 1997 Car and
Locomotive
Cyclopedia (Simmons-Boardman, Omaha, 1997) at page 726. The illustration
indicates
that the total loading of the spring groups, at solid, is indicated as 70,166
Lbs. per spring
group, giving a total of 140,334 Lbs. per truck and 280,668 Lbs. per single
unit autorack car.
The spring rate is indicated as 18,447 Lbs./in., per spring group or 36,894
Lbs./in for the
truck overall (there being one spring group per side frame, and two spring
groups per truck).
The truck shown in the 1997 Cyclopedia is a swing motion truck manufactured by
National
Castings Inc. In contrast to a regular 70 Ton truck that has, typically, 33
inch diameter
wheels, the 70 Ton special autorack truck has wheels that have a diameter of
only 28 inches.
This tends to allow for lower main deck wheel trackways, and hence greater
inside clearance
height. In part, the use of such a truck in an autorack car may reflect the
low density of the
lading. That is, a regular 70 Ton truck is designed to carry a gross weight on
rail of 110,000
Lbs, for a total full car weight of 220,000 Lbs. If the dead sprung weight of
a conventional
single unit autorack car is 75 - 85,000 Lbs., and the unsprung weight is about
15,000 Lbs,
that would leave about 120,000 Lbs., for lading. Assuming that a typical
passenger sedan
weighs about 2,500 Lbs., that would allow for about 48 automobiles before the
gross weight
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on rail would be exceeded. Even for larger, heavier vehicles, weighing perhaps
as much as
5,000 Lbs., this would still give some 24 light trucks, vans, or "sport
utility vehicles". But
the volumetric capacity of a single unit autorack rail road car may be about
12 ¨ 15 family
sedans and perhaps fewer light trucks, vans, or SUV's. Thus the autorack rail
road car truck
loading may often tend to be significantly less than 110,000 lbs.
In contrast to the philosophy underlying the design of the 70 Ton special 28
inch
truck, the present inventor believes that it is advantageous to use a truck
having wheels
larger than 33 inches in diameter for autorack rail road cars. Wheel life and
maintenance are
dependent on wheel loading, and, for the same loading history, inversely
dependent on
wheel diameter. A larger wheel may tend to have lower operating stresses for
the same
lading; may tend to have a greater wear allowance for braking; may tend to
undergo fewer
rotations than a wheel of smaller diameter for the same distance travelled,
and therefore may
tend to accumulate fewer cycles in terms of fatigue life; and may tend not to
get as hot
during braking. All of these factors may tend to increase wheel life and
reduce
maintenance. Further, a larger wheel diameter may be used in conjunction with
the use of
longer springs. The use of longer springs may permit the employment of springs
having a
softer spring rate, giving a gentler ride. In terms of fatigue life and wear,
this in turn may
tend to give a load history with reduced peak loads, and lower frequency of
those peak
loads. Attainment of any one of these advantages would be desirable.
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 and the dynamic response of the truck. 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. Rather than requiring independent suspension of each wheel, 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. It is a remarkably
simple and
durable layout. However, the dynamic performance of the truck flows from that
layout.
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 grain car, or fully loaded general purpose flat car.
In each case, the
design lading tended to be very heavy relative to the railcar 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.
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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
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.
Historically, autorack cars were made by building a rack structure on top of a
general purpose flat car. As such, the resultant car was sprung for the flat
car design
loads. When loaded with automobiles, this might yield a vertical bounce
natural
frequency in the range of 3 Hz. It would be preferable for the railcar
vertical bounce
natural frequency to be on the order of 1.4 Hz or less when loaded. Since this
natural
frequency varies as the square root of the quotient obtained by dividing the
spring rate of
the suspension by the overall sprung mass, it is desirable to reduce the
spring constant, to
increase the mass, or both.
One way to improve ride quality is to increase the dead sprung weight of the
rail
road car body. Deliberately increasing the mass of a freight car is counter
intuitive, since
many years of effort has gone into reducing the weight of railcars relative to
the weight of
the lading for the reasons noted above. One manufacturer, for example,
advertises a light
weight aluminium autorack car. However, given the high value and low density
of the
lading, adding weight may be reasonable to obtain a desired level of ride
quality.
Further, autorack railcars tend to be tall, long, and thin, with the upper
deck loads carried
at a relatively high location as measured from top of rail. A significant
addition of
weight at a low height relative to top of rail may also be beneficial in
reducing the height
of the center of gravity of the loaded car.
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
I
II
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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 autorack 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 railcar 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 autorack 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.
Consequently, it is now possible to use autorack 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 autorack rail road cars generally, including single car unit
autorack rail
road cars, articulated autorack 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
,
11
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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 1961
Car
Builder 's 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
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
railcar.
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
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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 may be 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 American 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, which may be about 80 to 86 inches, giving a ratio of 1.42
or 1.52.
This 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 or a larger number of softer coils of
smaller diameter. 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
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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 damper 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.
One determinant of overall ride quality is the dynamic response to lateral
perturbations. That is, when there is a lateral perturbation at track level,
the rigid steel
wheelsets of the truck may be pushed sideways relative to the car body.
Lateral
perturbations may arise for example from uneven track, or from passing over
switches or
from turnouts and other track geometry perturbations. When the train is moving
at speed,
the time duration of the input pulse due to the perturbation may be very
short.
The suspension system of the truck reacts to the lateral perturbation. It is
generally desirable for the force transmission to be relatively low. High
force
transmissibility, and corresponding high lateral acceleration, may tend not to
be
advantageous for the lading. This is particularly so if the lading includes
relatively
fragile goods, such as automobiles, electronic equipment, white goods, and
other
consumer products. In general, the lateral stiffness of the suspension
reflects the
combined displacement of (a) the sideframe between (i) the pedestal bearing
adapter and
(ii) the bottom spring seat (that is, the sideframes swing laterally as a
pendulum with the
pedestal bearing adapter being the top pivot point for the pendulum); and (b)
the lateral
deflection of the springs between (i) the lower spring seat in the sideframe
and (ii) the
upper spring mounting against the underside of the truck bolster, and (c) the
moment and
the associated angular displacement between the (i) spring seat in the
sideframe and (ii)
the upper spring mounting against the underside of the truck bolster.
In a conventional rail road car truck, the lateral stiffness of the spring
groups is
sometimes estimated as being approximately 1/2 of the vertical spring
stiffness. Thus the
choice of vertical spring stiffness may strongly affect the lateral stiffness
of the
suspension. The vertical stiffness of the spring groups may tend to yield a
vertical
I 1
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deflection at the releasable coupler from the light car (i.e., empty)
condition to the fully
laden condition of about 2 inches. For a conventional grain or coal car
subject to a
286,000 lbs., gross weight on rail limit, this may imply a dead sprung load of
some
50,000 lbs., and a live sprung load of some 220,000 lbs., yielding a spring
stiffness of 25
¨ 30,000 lbs./in., per spring group (there being, typically, two groups per
truck, and two
trucks per car). This may yield a lateral spring stiffness of 13 ¨ 16,000
lbs./in per spring
group. It should be noted that the numerical values given in this background
discussion
are approximations of ranges of values, and are provided for the purposes of
general
order-of-magnitude comparison, rather than as values of a specific truck.
The second component of stiffness relates to the lateral deflection of the
sideframe itself. In a conventional truck, the weight of the sprung load can
be idealized
as a point load applied at the center of the bottom spring seat. That load is
carried by the
sideframe to the pedestal seat mounted on the bearing adapter. The vertical
height
difference between these two points may be in the range of perhaps 12 to 18
inches,
depending on wheel size and sideframe geometry. For the general purposes of
this
description, for a truck having 36 inch wheels, 15 inches (+/-) might be taken
as a
roughly representative height.
The pedestal seat may typically have a flat surface that bears on an upwardly
crowned surface of the bearing adapter. The crown may typically have a radius
of
curvature of about 60 inches, with the center of curvature lying below the
surface (i.e.,
the surface is concave downward).
When a lateral shear force is imposed on the springs, there is a reaction
force in
the bottom spring seat that will tend to deflect the sideframe, somewhat like
a pendulum.
When the sideframe takes on an angular deflection in one direction, the line
of contact of
the flat surface of the pedestal seat with the crowned surface of the bearing
adapter will
tend to move along the arc of the crown in the opposite direction. That is, if
the bottom
spring seat moves outboard, the line of contact will tend to move inboard.
This motion is
resisted by a moment couple due to the sprung weight of the car on the bottom
spring
seat, acting on a moment arm between (a) the line of action of gravity at the
spring seat
and (b) the line of contact of the crown of the bearing adapter. For a 286,000
lbs. car the
apparent stiffness of the sideframe may be of the order of 18,000 ¨ 25,000
lbs./in,
measured at the bottom spring seat. That is, the lateral stiffness of the
sideframe (i.e., the
pendulum action by itself) can be greater than the (already relatively high)
lateral
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stiffness of the spring group in shear, and this apparent stiffness is
proportional to the
total sprung weight of the railcar (including lading). When taken as being
analogous to
two springs in series, the overall equivalent lateral spring stiffness may be
of the order of
8,000 lbs./in. to 10,000, per sideframe. A car designed for lesser weights may
have softer
apparent stiffness. This level of stiffness may not always yield as smooth a
ride as may
be desired.
There is another component of spring stiffness due to the unequal compression
of
the inside and outside portions of the spring group as the bottom spring seat
rotates
relative to the upper spring group mount under the bolster. This stiffness,
which is
additive to (that is, in parallel with) the stiffness of the sideframe, can be
significant, and
may be of the order of 3000 - 3500 lbs./in per spring group, depending on the
stiffness of
the springs and the layout of the group. Other second and third order effects
are
neglected for the purpose of this description. The total lateral stiffness for
one sideframe,
including the spring stiffness, the pendulum stiffness and the spring moment
stiffness, for
a S2HD 110 Ton truck may be about 9200 lbs/inch per side frame.
It has been observed that it may be preferable to have springs of a given
vertical
stiffness to give certain vertical ride characteristics, and a different
characteristic for
lateral perturbations. In particular, a softer lateral response may be desired
at high speed
(greater than about 50 m.p.h) and relatively low amplitude to address a truck
hunting
concern, while a different spring characteristic may be desirable to address a
low speed
(roughly 10 ¨ 25 m.p.h) roll characteristic, particularly since the overall
suspension
system may have a roll mode resonance lying in the low speed regime.
An alternate type of three piece truck is the "swing motion" truck. One
example
of a swing motion truck is shown at page 716 in the 1980 Car and Locomotive
Cyclopedia (1980, Simmons-Boardman, Omaha). This illustration, with captions
removed, is the basis of Figures la, lb and lc, herein, labelled "Prior Art".
Since the
truck has both lateral and longitudinal axes of symmetry, the artist has only
shown half
portions of the major components of the truck. The particular example
illustrated is a
swing motion truck produced by National Castings Inc., more commonly referred
to as
"NACO". Another example of a NACO Swing Motion truck is shown at page 726 of
the
1997 Car and Locomotive Cyclopedia (1997, Simmons-Boardroom, Omaha). An
earlier
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swing motion three piece truck is shown and described in US Patent 3,670,660
of Weber
et al., issued June 20, 1972.
In a swing motion truck, the sideframe is mounted as a "swing hanger" and acts
much like a pendulum. In contrast to the truck described above, the bearing
adapter has
an upwardly concave rocker bearing surface, having a radius of curvature of
perhaps 10
inches and a center of curvature lying above the bearing adapter. A pedestal
rocker seat
nests in the upwardly concave surface, and has itself an upwardly concave
surface that
engages the rocker bearing surface. The pedestal rocker seat has a radius of
curvature of
perhaps 5 inches, again with the center of curvature lying upwardly of the
rocker.
In this instance, the rocker seat is in dynamic rolling contact with the
surface of
the bearing adapter. The upper rocker assembly tends to act more like a hinge
than the
shallow crown of the bearing adapter described above. As such, the pendulum
may tend
to have a softer, perhaps much softer, response than the analogous
conventional
sideframe. Depending on the geometry of the rocker, this may yield a sideframe
resistance to lateral deflection in the order of 1/4 (or less) to about 'A of
what might
otherwise be typical. If combined in series with the spring group stiffness,
it can be seen
that the relative softness of the pendulum may tend to become the dominant
factor. To
some extent then, the lateral stiffness of the truck becomes less strongly
dependent on the
chosen vertical stiffness of the spring groups at least for small
displacements.
Furthermore, by providing a rocking lower spring seat, the swing motion truck
may tend
to reduce, or eliminate, the component of lateral stiffness that may tend to
arise because
of unequal compression of the inboard and outboard members of the spring
groups when
the sideframe has an angular displacement, thus further softening the lateral
response.
In the truck of US Patent 3,670,660 the rocking of the lower spring seat is
limited
to a range of about 3 degrees to either side of center, and a transom extends
between the
sideframes, forming a rigid, unsprung, lateral connecting member between the
rocker
plates of the two sideframes. In this context, "unsprung" refers to the
transom being
mounted to a portion of the truck that is not resiliently isolated from the
rails by the main
spring groups.
When the three degree condition is reached, the rockers "lock-up" against the
side
frames, and the dominant lateral displacement characteristic is that of the
main spring
groups in shear, as illustrated and described by Weber. The lateral, unsprung,
sideframe
CA 02860202 2016-12-20
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connecting member, namely the transom, has a stop that engages a downwardly
extending abutment on the bolster to limit lateral travel of the bolster
relative to the
sideframes. This use of a lateral connecting member is shown and described in
US Patent
3,461,814 of Weber, issued March 7, 1967. As noted in US Patent 3,670,660 the
use of a
spring plank had been known, and the use of an abutment at the level of the
spring plank
tended to permit the end of travel reaction to the truck bolster to be
transmitted from the
sideframes at a relatively low height, yielding a lower overturning moment on
the wheels
than if the end-of-travel force were transmitted through gibs on the truck
bolster from the
sidefiume columns at a relatively greater height. The use of a spring plank in
this way
was considered advantageous.
In Canadian Patent 2,090,031, (issued April 15, 1997 to Weber et al.,) noting
the
advent of lighter weight, low deck cars, Weber et al., replaced the transom
with a lateral
rod assembly to provide a rigid, unsprung connection member between the
platforms of
the rockers of the lower spring seats. As noted above, one type of car in
which relative
lightness and a low main deck has tended to be found is an Autorack car.
For the purposes of rapid estimation of truck lateral stiffness, the following
formula can be used:
ktruck = 2 x [ (ksideti-ameY1 + (kspring shear)1 r1
where
ksideframe = [kpendulum + kspring moment
kspring shear = The lateral spring constant for the spring group in
shear.
kpendulum = The force required to deflect the pendulum per unit of
deflection,
as measured at the center of the bottom spring seat.
kspring moment = The force required to deflect the bottom spring seat per unit
of
sideways deflection against the twisting moment caused by the
unequal compression of the inboard and outboard springs.
For the range of motion that may typically be of interest, and for small
angles of
deflection, kpendulum can be taken as being approximately constant at, for
example, the
value obtained for deflection of one degree. This may tend to be a
sufficiently accurate
approximation for the purposes of general calculation.
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In a pure pendulum, the lateral constant for small angles approximates k = W /
L,
where k is the lateral constant, W is the weight, and L is the pendulum
length. Further,
for the purpose of rapid comparison of the lateral swinging of the sideframes,
an
equivalent pendulum length for small angles of deflection can be defined as
Leg = W /
kpendulum= In this equation W represents the sprung weight borne by that
sideframe,
typically 1/4 of the total sprung weight for a symmetrical single unit
railcar. For a
conventional truck Leg may be of the order of about 3 or 4 inches. For a swing
motion
truck, Leg may be of the order of about 10 to 15 inches.
It is also possible to define the pendulum lateral stiffness (for small
angles) in
terms of the length of the pendulum, the radius of curvature of the rocker,
and the design
weight carried by the pendulum according to the formula:
kpendulum = (Flateral/Olateral) = (W/Lpendulum)[(Rcurvature/Lpendulum) + 1]
where:
kpendulum = the lateral stiffness of the pendulum
Fiaterat = the force per unit of lateral deflection
61ateral = a unit of lateral deflection
W = the weight borne by the pendulum
Lpendulum = the length of the pendulum, being the vertical distance from the
contact
surface of the bearing adapter to the bottom spring seat
Reurvature = the radius of curvature of the rocker surface
Following from this, if the pendulum stiffness is taken in series with the
lateral
spring stiffness, then the resultant overall lateral stiffness can be
obtained. Using this
number in the denominator, and the design weight in the numerator yields a
length,
effectively equivalent to a pendulum length if the entire lateral stiffness
came from an
equivalent pendulum according to Ltesultant = W / klateral total
For a conventional truck with a 60 inch radius of curvature rocker, and stiff
suspension, this length, Lresultant may be of the order of 6 ¨ 8 inches, or
thereabout.
So that the present invention may better be understood by comparison, in the
prior
art illustration of Figures la, lb, and lc, a NACO swing motion truck is
identified
generally as A20. Inasmuch as the truck is symmetrical about the truck center
both from
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side-to-side and lengthwise, the artist has shown only half of the bolster,
identified as
A22, and half of one of the sideframes, identified as A24.
In the customary manner, sideframe A24 has defined in it a generally
rectangular
window A26 that admits one of the ends of the bolster A28. The top boundary of
window A26 is defined by the sideframe arch, or compression member identified
as top
chord member A30, and the bottom of window A26 is defined by a tension member,
identified as bottom chord A32. The fore and aft vertical sides of window A26
are
defined by sideframe columns A34.
At the swept up ends of sideframe A24 there are sideframe pedestal fittings
A38
which each accommodate an upper rocker identified as a pedestal rocker seat
A40, that
engages the upper surface of a bearing adapter A42. Bearing adapter A42 itself
engages
a bearing mounted on one of the axles of the truck adjacent one of the wheels.
A rocker
seat A40 is located in each of the fore and aft pedestals, the rocker seats
being
longitudinally aligned such that the sideframe can swing transversely relative
to the
rolling direction of the truck A20 generally in what is referred to as a
"swing hanger"
arrangement.
The bottom chord of the sideframe includes pockets A44 in which a pair of fore
and aft lower rocker bearing seats A46 are mounted. The lower rocker seat A48
has a
pair of rounded, tapered ends or trunnions A50 that sit in the lower rocker
bearings A48,
and a medial platform A52. An array of four corner bosses A54 extend upwardly
from
platform A52.
An unsprung, lateral, rigid connecting member in the nature of a spring plank,
or
transom A60 extends cross-wise between the sideframes in a spaced apart,
underslung,
relationship below truck bolster A22. Transom A60 has an end portion that has
an array
of four apertures A62 that pick up on bosses A54. A grouping, or set of
springs A64
seats on the end of the transom, the corner springs of the set locating above
bosses A54.
The spring group, or set A64, is captured between the distal end of bolster
A22
and the end portion of transom A60. Spring set A64 is placed under compression
by the
weight of the railcar body and lading that bears upon bolster A22 from above.
In
consequence of this loading, the end portion of transom A60, and hence the
spring set,
are carried by platform A54. The reaction force in the springs has a load path
that is
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carried through the bottom rocker A70 (made up of trunnions A50 and lower
rocker
bearings A48) and into the sideframe A22 more generally.
Friction damping is provided by damping wedges A72 that seat in mating bolster
pockets A74. Bolster pockets A74 have inclined damper seats A76. The vertical
sliding
faces of the friction damper wedges then ride up an down on friction wear
plates A80
mounted to the inwardly facing surfaces of the sideframe columns.
The "swing motion" truck gets its name from the swinging motion of the
sideframe on the upper rockers when a lateral track perturbation is imposed on
the
wheels. The reaction of the sideframes is to swing, rather like pendula, on
the upper
rockers. When this occurs, the transom and the truck bolster tend to shift
sideways, with
the bottom spring seat platform rotating on the lower rocker.
The upper rockers are inserts, typically of a hardened material, whose
rocking, or
engaging, surface A82 has a radius of curvature of about 5 inches, with the
center of
curvature (when assembled) lying above the upper rockers (i.e., the surface is
upwardly
concave).
As noted above, one of the features of a swing motion truck is that while it
may
be quite stiff vertically, and while it may be resistant to parallelogram
deformation
because of the unsprung lateral connection member, it may at the same time
tend to be
laterally relatively soft.
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
of the vertically oriented spring, is the complementary angle, in this
example, being about
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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 railcar 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, grain 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.
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.
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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 - 'A
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
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.
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
CA 02860202 2016-12-20
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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
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
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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.
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.
SUMMARY OF THE INVENTION
In an aspect of the invention there is an autorack rail road car having a car
body
for the transport of automobiles, the car body being supported for rolling
motion along
rail road tracks by rail road car trucks. At least one of the trucks has
wheels whose
diameter is greater than 33 inches.
In a further feature of that aspect of the invention, at least one of the
trucks has
wheels that are at least 36 inches in diameter. In another feature of that
aspect of the
invention, the rail road car truck has wheels that are at least 38 inches in
diameter. In yet
a further feature of that aspect of the invention, at least one of the rail
road car trucks has
an overall vertical spring rate of less than 50,000 Lbs./in. In a further
feature, the overall
vertical spring rate of the truck is less than 40,000 Lbs./in. In a still
further feature, the
overall vertical spring rate is less than 30,000 Lbs./in. In a still further
feature, the overall
1
II
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vertical spring rate is less than 20,000 Lbs./in. In a still further feature,
the overall
vertical spring rate is in the range of 10,000 Lbs/in. to 20,000 Lbs./in.
In a still further feature, at least one of the trucks is a swing motion
truck. In an
additional feature, the truck includes a pair of first and second side frames
and a
transversely oriented truck bolster mounted between the side frames. The side
frames are
mounted to the wheelsets, and are able to swing laterally relative to the
wheels. The
effective equivalent length of the swinging side frames is greater than 10
inches.
In a still further feature, at least one of the trucks is free of unsprung
lateral cross-
members. In another feature of that feature of the invention, the truck is
free of a
transom.
In still another feature of that aspect of the invention, at least one of the
trucks has
friction dampers mounted in laterally spaced pairs, the dampers being biased
to exert a
squaring restorative moment couple on the truck bolster relative to the side
frames when
the truck bolster is deflected from square relative to the side frames. In
still another
feature of that aspect of the invention, at least one of the trucks has
springs mounted in
inboard and outboard pairs between the bolster and each of the side frames,
said inboard
and outboard pairs being oriented to provide a squaring restorative moment
couple to the
bolster relative to the side frames.
In still another feature of the invention, the railcar includes a railcar body
unit that
has a weight of at least 90,000 Lbs., in an unloaded condition. In a further
feature of the
invention, the railcar body unit has an unladen weight of at least 100,000
Lbs. In another
further feature the railcar body unit has an unladen weight of at least
120,000 Lbs. In
another further feature, the railcar body unit has an unladen weight of at
least 130,000
Lbs.
In another feature of that aspect of the invention, the rail road car body
unit
includes at least 15,000 Lbs., of ballast. In another feature, the rail road
car body unit
includes at least 25,000 Lbs., of ballast. In another feature of the
invention, the rail road
car body unit includes at least 40,000 Lbs., of ballast. In a further feature
of the
invention, the ballast weight is incorporated in a deck plate. In another
feature of the
invention the rail road car has a deck plate exceeding 3/8 inches in
thickness. In another
feature of the invention the rail road car body has a deck plate exceeding 1/2
inches in
1
11
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thickness. In another feature of the invention the rail road car body has a
deck plate
exceeding 3/4 inches in thickness. In another feature of the invention the
rail road car
body has a deck plate exceeding 1 inch in thickness. In another feature of the
invention
the rail road car body has a deck plate exceeding 1 1/4 inch in thickness.
In another feature of that aspect of the invention at least one of the railcar
trucks
has a wheelbase exceeding 73 inches in length. In another feature at least one
of the
trucks has a wheelbase that exceeds 1.3 times the gauge width of the rails. In
another
feature the wheelbase is in the range of 78 to 88 inches in length. In another
feature of
the invention the wheelbase is in the range of 1.3 to 1.6 times the track
gauge width.
In another feature of the invention, the rail road car is an articulated
railroad car.
In still another feature of the invention, the rail road car is an articulated
rail road car, and
one of the articulated connectors is cantilevered relative to the truck
closest thereto. In
another feature the articulated rail road car is a three pack rail road car.
In still another
feature the three pack rail road car has a middle unit connected between two
end units.
Each of the end units has a coupler end truck, and each of the end units has
an
asymmetric car body weight distribution in which most of the weight of the end
car body
is carried by the end truck. In a further feature, the end car body is
ballasted. In a still
further feature, the ballast of the end car body is has a distribution that is
biased toward
the end truck.
BRIEF DESCRIPTION OF THE DRAWINGS
Figure la shows a prior art exploded partial view illustration of a swing
motion
truck, much as shown at page 716 in the 1980 Car and Locomotive
Cyclopedia;
Figure lb shows a cross-sectional detail of an upper rocker assembly of the
truck
of Figure la;
Figure lc shows a cross-sectional detail of a lower rocker assembly of the
truck of
Figure la;
Figure 2a shows a side view of a single unit autorack rail road car;
Figure 2b shows a cross-sectional view of the autorack rail road car of Figure
2a in a
bi-level configuration, one half section of Figure 2b being taken through the
main bolster and the other half taken looking at the cross-tie outboard of the
main bolster;
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Figure 2c shows a half sectioned partial end view of the rail road car of
Figure 2a
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;
Figure 2d shows a partially sectioned side view of the rail road car of Figure
2c
illustrating the relationship of the truck, the bolster and the wheel
clearance,
below the main deck;
Figure 3a shows a side view of a two unit articulated autorack rail road car;
Figure 3b shows a side view of an alternate autorack rail road car to that of
Figure
3a, having a cantilevered articulation;
Figure 4a shows a side view of a three unit autorack rail road car;
Figure 4b shows a side view of an alternate three unit autorack rail road car
to the
articulated rail road unit car of Figure 4a, having cantilevered
articulations;
Figure 4c shows an isometric view of an end unit of the three unit autorack
rail road
car of Figure 4b;
Figure 5a is a partial side sectional view of the draft pocket of the coupler
end of any
of the rail road cars of Figures 2a, 3a, 3b, 4a, or 4b taken on '5a ¨ 5a' as
indicated in Figure 2a; and
Figure 5b shows a top view of the draft gear at the coupler end of Figure 5a
taken on
`5b ¨ 5b' of Figure 5a;
Figure 6a shows a swing motion truck as shown in Figure la, but lacking a
transom;
Figure 6b shows a cross-sectional detail of a bottom spring seat of the truck
of
Figure 6a;
Figure 6c shows a cross-sectional detail of a bottom spring seat of the truck
of
Figure 6a;
Figure 7a shows a swing motion truck having an upper rocker as in the swing
motion truck of Figure la, but having a rigid spring seat, and being free of
a transom;
Figure 7b shows a cross-sectional detail of the upper rocker assembly of the
truck
of Figure 7a;
Figure 8 shows a swing motion truck similar to that of Figure 7a, but having
doubled bolster pockets and wedges;
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Figure 9a shows an isometric view of a three piece truck for the autorack rail
road
cars of Figures 2a, 3a, 3b, 4a or 4h;
Figure 9b shows a side view of the three piece truck of Figure 9a;
Figure 9c shows a top view of half of the three piece truck of Figure 9b;
Figure 9d shows a partial section of the three piece truck of Figure 9b taken
on '9d ¨
9d';
Figure 9e shows a partial isometric view of the truck bolster of the three
piece truck
of Figure 9a showing friction damper seats;
Figure 9f shows a force schematic for dampers in the side frame of the truck
of
Figure 9a.
Figure 10a shows a side view of an alternate three piece truck to that of
Figure 9a;
Figure 10b shows a top view of half of the three piece truck of Figure 10a;
and
Figure 10c shows a partial section of the three piece truck of Figure 10a
taken on
'10c ¨ 10c'.
Figure ha shows an alternate version of the bolster of Figure 9e, with a
double
sized damper pocket for seating a large single wedge having a welded insert;
Figure llb shows an alternate optional dual wedge for a truck bolster like
that of
Figure 11a;
Figure 11c shows an alternate bolster, similar to that of Figure 9a, having a
pair of
spaced apart wedge pockets, and pocket inserts with both primary and
secondary wedge angles;
Figure lid shows an alternate bolster, similar to that of Figure 11c, and
split
wedges;
Figure 12 shows an optional non-metallic wear surface arrangement for dampers
such as used in the bolster of Figure 11b;
Figure 13a shows a bolster similar to that of Figure 11c, having a wedge
pocket
having primary and secondary angles and a split wedge arrangement for use
therewith;
Figure 13b shows an alternate stepped single wedge for the bolster of Figure
13a;
Figure 13c is a view looking along a plane on the primary angle of the split
wedge of
Figure 13a relative to the bolster pocket taken on '13c ¨ 13c';
Figure 13d is a view looking along a plane on the primary angle of the stepped
wedge of Figure 13b relative to the bolster pocket taken on '13d ¨ 13d';
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Figure 14a shows an alternate bolster and wedge arrangement to that of Figure
11b,
having secondary wedge angles;
Figure 14b shows an alternate, split wedge arrangement for the bolster of
Figure
14a;
Figure 14c shows a cross-section of a stepped damper wedge for use with a
bolster
as shown in Figure 14a;
Figure 14d shows an alternate stepped damper to that of Figure 14c;
Figure 15a is a section of Figure 9b showing a replaceable side frame wear
plate;
Figure 15b is a sectional view on of the side frame of Figure 15a 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 15a;
Figure 15c shows a compound bolster pocket for the bolster of Figure 15a;
Figure 15d shows a side view detail of the bolster pocket of Figure 15c, as
installed,
relative to the main springs and the wear plate;
Figure 15e 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
15c;
Figure 15f shows an alternate, stepped steeper angle profile for the primary
angle of
the wedge of the bolster pocket of Figure 15d;
Figure 15g shows a welded insert having a profile for mating engagement with
the corresponding face of the bolster pocket of Figure 15d;
Figure 16a shows an exploded isometric view of an alternate bolster and side
frame
assembly to that of Figure 9a, in which horizontally acting springs drive
constant force dampers;
Figure 16b shows a side-by-side double damper arrangement similar to that of
Figure 16a;
Figure 17a shows an isometric view of an alternate railroad car truck to that
of
Figure 9a;
Figure 17b shows a side view of the three piece truck of Figure 17a.
Figure 17c shows a top view of the three piece truck of Figure 17a.
Figure 17d shows an end view of the three piece truck of Figure 17a.
Figure 17e shows a schematic of a spring layout for the truck of Figure 17a.
,
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DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
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.
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 railcar 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 railcar 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
railcar 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 Lbs GWR and the "70 Ton Special" low profile truck sometimes used for
autorack 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
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may generally also be intended to describe the other half as well, allowing
for differences
between right hand and left hand parts.
Portions of this application refer to friction dampers, and multiple friction
damper
systems. There are several types of damper arrangement as shown at pages 715 -
716 of
the 1997 Car and Locomotive Encyclopedia. Double damper arrangements are shown
and described in my co-pending US Patent application. Each of the arrangements
of
dampers shown at pp. 715 to 716 of the 1997 Car and Locomotive Encyclopedia
can be
modified to employ a four cornered, double damper arrangement of inner and
outer
dampers.
Figures 2a, 3a, 3b, 4a, and 4b, show different types of rail road freight cars
in the
nature of autorack rail road cars, all sharing a number of similar features.
Figure 2a (side
view) shows a single unit autorack rail road car, indicated generally as 20.
It has a railcar
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 railcar body 22. Body 22 has
a housing
structure 30, including a pair of left and right hand sidewall structures 32,
34 and 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 2b, 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 bi-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
1
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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 Autorack Car
Similarly, Figure 3a shows a two unit articulated autorack rail road car,
indicated
generally as 80. It has a first railcar unit body 82, and a second railcar
unit body 83, both
supported for rolling motion in the longitudinal direction (i.e., along the
rails) upon
railcar trucks 84, 86 and 88. Railcar trucks 84 and 88 are mounted at main
bolsters at
respective coupler ends of the first and second railcar unit bodies 82 and 83.
Truck 86 is
mounted beneath articulated connector 90 by which bodies 82 and 83 are joined
together.
Each of bodies 82 and 83 has a housing structure 92, 93, including a pair of
left and right
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
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(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, 133 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
autorack
rail road car when doors 132 or 133 are opened to permit circus loading of the
decks.
For the purposes of this description, the cross-section of Figure 2b 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 2b 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
main bolster. The sections of Figures 2b and 2c 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 2c 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.
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Three or More Unit Articulated Autorack Car
Figure 4a shows a three unit articulated autorack rail road car, generally as
140.
It has a first end railcar unit body 142, a second end railcar unit body 144,
and an
intermediate railcar unit body 146 between railcar unit bodies 142 and 144.
Railcar unit
bodies 142, 144 and 146 are supported for rolling motion in the longitudinal
direction
(i.e., along the rails) upon railcar trucks 148, 150, 152, and 154. Railcar
trucks 148 and
150 are "coupler end" trucks mounted at main bolsters at respective coupler
ends of the
first and second railcar 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. Railcar
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 railcar body 144.
Body 146 has a housing structure 166 like that of Figure 2b, 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.
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 autorack 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
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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 3b
and
4b. In Figure 3b, a two unit articulated autorack rail road car is indicated
generally as
200. It has first and second railcar unit bodies 202, 204 supported for
rolling motion in
the longitudinal direction by three rail road car trucks, 206, 208 and 210
respectively.
Railcar unit bodies 202 and 204 are joined together at an articulated
connector 212. In
this instance, while railcar bodies 202 and 204 share the same basic
structural features of
railcar body 22, in terms of a through center sill, cross-bearers, side sills,
walls and
canopy, and vehicles decks, railcar body 202 is a "two-truck" body, and
railcar body 204
is a single truck body. That is, railcar 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, railcar 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 railcar bodies 202 and
204,
longitudinally outboard of railcar 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 /
614,815 for a Rail Road Car with Cantilevered Articulation filed July 12,
2000, and may
tend to permit a longer car body for a given articulated rail road car truck
center distance
as therein described.
Figure 4b 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 4c 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
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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
railcar body, or parts of different railcar bodies of a multi-unit rail road
car joined by
articulated connections, draw-bars, or a combination of articulated
connections and draw-
bars.
Figures 5a and 5b show an example of a draft gear arrangement that may be used
at a first coupler end 300 of rail road car 20, coupler end 300 being
representative of
either of the coupler ends 300, 330 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, 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 railcar 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. The draft pocket arrangement may include a flared bell-mouth and other
features
differing from the illustrated example.
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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 % 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 5a and 5b, 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 examples herein, all of the trucks may have wheels that are
greater
than 33 inches in diameter. The wheels can advantageously be 36 inches or 38
inches in
diameter, or possibly larger depending on deck height geometry, and are
preferred to be
36 inch wheels. Although it is advantageous for the wheels of all of the
trucks to be of
the same diameter, it is not necessary. That is, one or more trucks, such as
the
intermediate truck or trucks in an articulated autorack rail road car
embodiment can have
wheels of a larger diameter than 33 inches such as 36 or 38 inches, for
example, whereas
the other trucks, namely the end trucks can have 33 inch or other wheels.
Weight Distribution
In each of the autorack railcar embodiments described above, each of the car
units
has a weight, that weight being carried by the railcar trucks with which the
car is
equipped. In each of the embodiments of articulated railcars described above
there is a
number of railcar 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
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of articulated car units is one more than the number of articulations, and one
less than the
number of trucks. In the event that some of the cars units are 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 case 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 4b, 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 full
weight. The full
weight is usually limited by the truck capacity, whether 70 ton (33 inch
diameter wheels),
100 ton (36 inch diameter wheels), 110 ton (36 inch diameter wheels, 286,000
Lbs.) or
125 ton (38 inch diameter wheels). 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 railcar unit is generally taken as the body weight
of
the car, including any ballast, as described below, plus that portion of the
weight of the
truck bearing on the springs, that portion most typically 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. The unsprung weight of a three piece truck may generally be about
8,800 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,
and is not to exceed the rated value for the truck.
In each of the embodiments described above, each of the railcar 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
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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 autorack rail road car 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, Wm, 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
railcar 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 4b and 4c
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 railcar
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 railcar
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
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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
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. In alternate embodiments plate thickness T can be a thickness greater
than 1/4 inches,
whether 1/4 inches, I inch, or 1 'A inches, or some other thickness. Further,
the ballast
plate need not be a monolithic cut sheet, but can be made up of a plurality of
plates
mounted at appropriate locations to yield a mass (or weight) of ballast of
suitable
distribution.
Similar weight distributions can be made for other capacities of truck whether
100
Ton, 110 Ton or 125 Ton. With an increase in truck capacity beyond "70 Ton",
there is
correspondingly an opportunity to add more ballast up to the truck capacity
limit. As
noted above, 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.
In the example of Figure 6a and 6b, a truck embodying an aspect of the present
invention is indicated as 410. Truck 410 differs from truck A20 of Figure la
insofar as it
is free of a rigid, unsprung lateral connecting member in the nature of
unsprung cross-
bracing such as a frame brace of crossed-diagonal rods, lateral rods, or a
transom (such as
transom A60) running between the rocker plates of the bottom spring seats of
the
opposed sideframes. Further, truck 410 employs gibs 412 to define limits to
the lateral
range of travel of the truck bolster 414 relative to the sideframe 416. In
other respects,
including the sideframe geometry and upper and lower rocker assemblies, truck
410 is
intended to have generally similar features to truck A20, although it may
differ in size,
pendulum length, spring stiffness, wheelbase, window width and window height,
and
damping arrangement. The determination of these values and dimensions may
depend on
the service conditions under which the truck is to operate.
CA 02860202 2016-12-20
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As with other trucks described herein, it will be understood that since truck
410
(and trucks 420, 520, and 600, described below) are symmetrical about both
their
longitudinal and transverse axes, the truck is shown in partial section. In
each case,
where reference is made to a sideframe, it will be understood that the truck
has first and
second sideframes, first and second spring groups, and so on.
In Figures 7a and 7b, for example, a truck is identified generally as 420.
Inasmuch as truck 420 is symmetrical about the truck center both from side-to-
side and
lengthwise, the bolster, identified as 422, and the sideframes, identified as
424 are shown
in part. Truck 420 differs from truck A20 of the prior art, described above,
in that truck
420 has a rigid bottom spring seat 444 rather than a lower rocker as in truck
A20, as
described below, and is free of a rigid, unsprung lateral connection member
such as an
underslung transom A60, a frame brace, or laterally extending rods.
Sideframe 424 has a generally rectangular window 426 that accommodates one of
the ends 428 of the bolster 422. The upper boundary of window 426 is defined
by the
sideframe arch, or compression member identified as top chord member 430, and
the
bottom of window 426 is defined by a tension member identified as bottom chord
432.
The fore and aft vertical sides of window 426 are defined by sideframe columns
434.
The ends of the tension member sweep up to meet the compression member. At
each of the swept-up ends of sideframe 424 there are sideframe pedestal
fittings 438.
Each fitting 438 accommodates an upper rocker identified as a pedestal rocker
seat 440.
Pedestal rocker seat 440 engages the upper surface of a bearing adapter 442.
Bearing
adapter 442 engages a bearing mounted on one of the axles of the truck
adjacent one of
the wheels. A rocker seat 440 is located in each of the fore and aft pedestal
fittings 438,
the rocker seats 440 being longitudinally aligned such that the sideframe can
swing
transversely relative to the rolling direction of the truck in a "swing
hanger" arrangement.
Bearing adapter 442 has a hollowed out recess 441 in its upper surface that
defines a bearing surface for receiving rocker seat 440. Bearing surface 441
is formed on
a radius of curvature RI. The radius of curvature R1 is preferably in the
range of less
than 25 inches, may be in the range of 5" to 15", and is preferably in the
range of 8 to 12
inches, and most preferably about 10 inches with the center of curvature lying
upwardly
of the rocker seat. The lower face of rocker seat 440 is also formed on a
circular arc,
having a radius of curvature R2 that is less than the radius of curvature R1
of the recess of
CA 02860202 2016-12-20
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surface recess 441. R2 is preferably in the range of 1/4 to 3/4 as large as
121, and is
preferably in the range of 3 ¨ 10 inches, and most preferably 5 inches when R1
is 10
inches, ie., R2 is one half of RI. Given the relatively small angular
displacement of the
rocking motion of R2 relative to R1 (typically less than +/- 10 degrees) the
relationship is
one of rolling contact, rather than sliding contact.
The bottom chord or tension member of sideframe 424 has a basket plate, or
lower spring seat 444 rigidly mounted to bottom chord 432, such that it has a
rigid
orientation relative to window 426, and to sideframe 424 in general. That is,
in contrast
to the lower rocker platform of the prior art swing motion truck A20 of Figure
la, as
described above, spring seat 444 is not mounted on a rocker, and does not rock
relative to
sideframe 424. Although spring seat 444 retains an array of bosses 446 for
engaging the
corner elements 450, namely springs 454 and 455 (inboard), 456 and 457
(outboard) of a
spring set 448, there is no transom mounted between the bottom of the springs
and seat
444. Seat 444 has a peripheral lip 452 for discouraging the escape of the
bottom ends of
the springs.
The spring group, or spring set 448, is captured between the distal end 428 of
bolster 422 and spring seat 444, being placed under compression by the weight
of the
railcar body and lading that bears upon bolster 422 from above.
Friction damping is provided by damping wedges 462 that seat in mating bolster
pockets 464 that have inclined damper seats 466. The vertical sliding faces
470 of the
friction damper wedges 462 then ride up and down on friction wear plates 472
mounted
to the inwardly facing surfaces of sideframe columns 434. Angled faces 474 of
wedges
462 ride against the angled face of seat 466. Bolster 422 has inboard and
outboard gibbs
476, 478 respectively, that bound the lateral motion of bolster 422 relative
to sideframe
columns 434. This motion allowance may advantageously be in the range of +/- 1-
1/8 to
1-3/4 inches, and is most preferably in the range of 1-3/16 to 1-9/16 inches,
and can be
set, for example, at 1 1/2 inches or 1 1/4 inches of lateral travel to either
side of a neutral,
or centered, position when the sideframe is undeflected.
As in the prior art swing motion truck A20, a spring group of 8 springs in a
3:2:3
arrangement is used. Other configurations of spring groups could be used, such
as those
described below.
CA 02860202 2016-12-20
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In the embodiment of Figure 8, a truck 520 is substantially similar to truck
420,
but differs insofar as truck 520 has a bolster 522 having double bolster
pockets 524, 526
on each face of the bolster at the outboard end. Bolster pockets 524, 526
accommodate a
pair of first and second, laterally inboard and laterally outboard friction
damper wedges
528, 529 and 530, 531, respectively. Wedges 528, 529 each sit over a first,
inboard
corner spring 532, 533, and wedges 530, 531 each sit over a second, outboard
corner
spring 534, 535. In this four corner arrangement, each damper is individually
sprung by
one or another of the springs in the spring group. The static compression of
the springs
under the weight of the car body and lading tends to act as a spring loading
to bias the
damper to act along the slope of the bolster pocket to force the friction
surface against the
sideframe. As such, the dampers co-operate in acting as biased members working
between the bolster and the side frames to resist parallelogram, or lozenging,
deformation
of the side frame relative to the truck bolster. A middle end spring 536 bears
on the
underside of a land 538 located intermediate bolster pockets 524 and 526. The
top ends
of the central row of springs, 540, seat under the main central portion 542 of
the end of
bolster 522.
The lower ends of the springs of the entire spring group, identified generally
as
544, seat in the lower spring seat 546. Lower spring seat 546 has the layout
of a tray with
an upturned rectangular peripheral lip. Lower spring seat 546 is rigidly
mounted to the
lower chord 548 of sideframe 549. In this case, spring group 544 has a 3 rows
x 3
columns layout, rather than the 3:2:3 arrangement of truck 420. A 3 x 5 layout
as shown
in Figure 17e (described below) could be used, as could other alternate spring
group
layouts. Truck 520 is free of any rigid, unsprung lateral sideframe connection
members
such as transom A60.
It will be noted that bearing plate 550 mounted to vertical sideframe columns
552
is significantly wider than the corresponding bearing plate 472 of truck 420
of Figure 7a.
This additional width corresponds to the additional overall damper span width
measured
fully across the damper pairs, plus lateral travel as noted above, typically
allowing
roughly 1 1/2 (+/-) inches of lateral travel (i.e. for an overall total of
roughly 3" travel) of
the bolster relative to the sideframe to either side of the undefiected
central position.
That is, rather than having the width of one coil, plus allowance for travel,
plate 550 has
the width of three coils, plus allowance to accommodate 1 1/2 (+/-) inches of
travel to
either side. Plate 550 is significantly wider than the through thickness of
the sideframes
more generally, as measured, for example, at the pedestals.
I I
CA 02860202 2016-12-20
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Damper wedges 528 and 530 sit over 44 % (+/-) of the spring group i.e., 4/9 of
a 3
rows x 3 columns group as shown in Figure 8, whereas wedges 470 only sat over
2/8 of
the 3:2:3 group in Figure 7a. For the same proportion of vertical damping,
wedges 528
and 530 may tend to have a larger included angle (i.e., between the wedge
hypotenuse
and the vertical face for engaging the friction wear plates on the sideframe
columns 552.
For example, if the included angle of friction wedges 472 is about 35 degrees,
then,
assuming a similar overall spring group stiffness, and single coils, the
corresponding
angle of wedges 528 and 530 could advantageously be in the range of 50 ¨ 65
degrees, or
more preferably about 55 degrees. In a 3 x 5 group such as group 976 of truck
970 of
Figure 17e, for coils of equal stiffness, the wedge angle may tend to be in
the 35 to 40
degree range. The specific angle will be a function of the specific spring
stiffnesses and
spring combinations actually employed.
The use of spaced apart pairs of dampers 528, 530 may tend to give a larger
moment arm, as indicated by dimension "2M", for resisting parallelogram
deformation of
truck 520 more generally as compared to trucks 420 or A20. Parallelogram
deformation
may tend to occur, for example, during the "truck hunting" phenomenon that has
a
tendency to occur in higher speed operation.
Placement of doubled dampers in this way may tend to yield a greater
restorative
"squaring" force to return the truck to a square orientation than for a single
damper alone,
as in truck 420. That is, in parallelogram deformation, or lozenging, the
differential
compression of one diagonal pair of springs (e.g., inboard spring 532 and
outboard spring
535 may be more pronouncedly compressed) relative to the other diagonal pair
of springs
(e.g., inboard spring 533 and outboard spring 534 may be less pronouncedly
compressed
than springs 532 and 535) tends to yield a restorative moment couple acting on
the
sideframe wear plates. This moment couple tends to rotate the sideframe in a
direction to
square the truck, (that is, in a position in which the bolster is
perpendicular, or "square",
to the sideframes) and thus may tend to discourage the lozenging or
parallelogramming,
noted by Weber.
Figures 9a, 9b, 9c, 9d and 9e all relate to a three piece truck 600 for use
with the
rail road cars of Figure 2a, 3a, 3b, 4a or 4b. Figures 2c and 2d show the
relationship of
this truck to the deck level of these rail road cars. Truck 600 has three
major elements,
those elements being a truck bolster 602, symmetrical about the truck
longitudinal
centreline, and a pair of first and second side frames, indicated as 604. Only
one side
CA 02860202 2016-12-20
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frame is shown in Figure 9c given the symmetry of truck 600. Three piece truck
600 has
a resilient suspension (a primary suspension) provided by a spring groups 605
trapped
between each of the distal (i.e., transversely outboard) ends of truck bolster
602 and side
frames 604.
Truck bolster 602 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 606). A center plate or center bowl 608 is located at
the truck
center. An upper flange 610 extends between the two ends 606, being narrow at
a central
waist and flaring to a wider transversely outboard termination at ends 606.
Truck bolster
602 also has a lower flange 612 and two fabricated webs 614 extending between
upper
flange 610 and lower flange 612 to form an irregular, closed section box beam.
Additional webs 615 are mounted between the distal portions of upper flange
610 and
614 where bolster 602 engages one of the spring groups 605. The transversely
distal
region of truck bolster 602 also has friction damper seats 616, 618 for
accommodating
friction damper wedges as described further below.
Side frame 604 is a casting having bearing seats 619 into which bearing
adapters
620, bearings 621, and a pair of axles 622 mount. Each of axles 622 has a pair
of first
and second wheels 623, 625 mounted to it in a spaced apart position
corresponding to the
width of the track gauge of the track upon which the railcar is to operate.
Side frame 604
also has a compression member, or upper beam member 624, a tension member, or
lower
beam member 626, and vertical side columns 628 and 630, each lying to one side
of a
vertical transverse plane 601 bisecting truck 600 at the longitudinal station
of the truck
center. A generally rectangular opening in the nature of a sideframe window
627 is
defined by the co-operation of the upper and lower beam members 624, 626 and
vertical
columns 628, 630. The distal end of truck bolster 602 can be introduced into
window
627. The distal end of truck bolster 602 can then move up and down relative to
the side
frame within this opening. Lower beam member 626 (the tension member) has a
bottom
or lower spring seat 632 upon which spring group 605 can seat. Similarly, an
upper
spring seat 634 is provided by the underside of the distal portion of bolster
602 to
engages the upper end of spring group 605. As such, vertical movement of truck
bolster
602 will tend to compress or release the springs in spring group 605.
For the purposes of this description the swivelling, 4 wheel, 2 axle truck 600
has
first and second sideframes 604 that can be taken as having the same upper
rocker
CA 02860202 2016-12-20
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assembly as truck 520, and has a rigidly mounted lower spring seat 632, like
spring seat
544, but having a shape to suit the 2 rows x 4 columns spring layout rather
than the 3 x 3
layout of truck 520. It may also be noted that sideframe window 627 has
greater width
between sideframe columns 628, 630 than window 526 between columns 552 to
accommodate the longer spring group footprint, and bolster 602 similarly has a
wider end
to sit over the spring group.
In the embodiment of Figure 9a, spring group 605 has two rows of springs 636,
a
transversely inboard row and a transversely outboard row, each row having four
large (8
inch +/-) diameter coil springs 636, 637, 638, 639 giving vertical bounce
spring rate
constant, k, for group 605 of less than 10,000 lbs / inch. This spring rate
constant can be
in the range of 6,000 to 10,000 lbs / in., and is advantageously in the range
of 7,000 to
9,500 lbs / in, and preferably in the range of 8,000 ¨ 8,500 lbs./in., giving
an overall
vertical bounce spring rate for the truck of double these values, preferably
in the range of
14,000 to 18,500 lbs / in, or more narrowly, 16,000 ¨ 17,000 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 640,
641, 642
and 643 that engage the sockets, or seats 616, 618 in a four-cornered
arrangement. The
corner springs in spring group 605 bear upon a friction damper wedge 640, 641,
642 or
643. Each of vertical columns 628, 630 has a friction wear plate 650 having
transversely
inboard and transversely outboard regions against which the friction faces of
wedges 640,
641, 642 and 643 can bear, respectively. Bolster gibs 651 and 653 lie inboard
and
outboard of wear plate 650 respectively. Gibs 651 and 653 act to limit the
lateral travel
of bolster 602 relative to side frame 604. 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.
CA 02860202 2016-12-20
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As seen in the top view of Figure 9c, and in the schematic sketch of Figure 9f
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 9e, the damper seats are shown as being
segregated by
a partition 652. If a longitudinal vertical plane 654 is drawn through truck
600 through
the center of partition 652, it can be seen that the inboard dampers lie to
one side of plane
654, and the outboard dampers lie to the outboard side of plane 654. 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 654 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 640
against column
628 can be thought of as a point load offset transversely relative to the
diagonally
outboard friction face of wedge 643 against column 630 by a distance that is
notionally
twice dimension `L' shown in the conceptual sketch of Figure 9f. 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 + F4)]L. As indicated by the formulae on the conceptual
sketch of Figure
9f, the difference between the inboard and outboard forces on each side of the
bolster is
proportional to the angle of deflection E of the truck bolster relative to the
side frame, and
CA 02860202 2016-12-20
- 44 -
since the normal forces due to static deflection xo may tend to cancel out, MR
=
41(cTan(E)Tan(0)L, where 0 is the primary angle of the damper, and IQ 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 9b, it
is 50 % of
the group total (i.e., 4 of 8 equal springs). The wedge angle of wedges 640,
642 is
significantly greater than 35 degrees. 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, whether in the range of between
roughly 37 to
40 or 45 degrees, to roughly 60 or 65 degrees.
In this example, damper wedges 640, 641 and 642, 643 sit over 50 % of the
spring
group i.e., 4/8 namely springs 636, 637, 638, 639. For the same proportion of
vertical
damping as in truck 420, wedges 640, 641 and 642, 643 may tend to have a
larger
included angle, possibly about 60 degrees, although angles in the range of 45
to 70
degrees could be chosen depending on spring combinations and spring
stiffnesses. Once
again, in a warping condition, the somewhat wider damping region (the width of
two full
coils plus lateral travel of 1
(+/-)) of sideframe column wear plates 627, 629 lying
between inboard and outboard gibbs 611, 613, 615, 617 relative to truck 20 (a
damper
width of one coil with travel), sprung on individual springs (inboard and
outboard in
truck 600, as opposed to a single central coil in truck 20), may tend to
generate a moment
couple to give a restoring force working on a moment arm. This restoring force
may tend
to urge the sideframe back to a square orientation relative to the bolster,
with diagonally
opposite pairs of springs working as described above. In this instance, the
springs each
work on a moment arm distance corresponding to half of the distance between
the centers
of the 2 rows of coils, rather than half the 3 coil distance shown in Figure
8.
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
CA 02860202 2016-12-20
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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.
The size of the spring group embodiment of Figure 9b yields a side frame
window
opening having a width between the vertical columns of side frame 604 of
roughly 33
inches. This is relatively large compared to existing spring groups, being
more than 25 %
greater in width. In an alternate 3 x 5 spring group arrangement of 5 1/2"
diameter
springs, the opening between the sideframe columns is more than 27 1/2 inches
wide, in
one preferred embodiment being between 29 and 30 inches wide, namely about 29
¨ 'A
inches.
Truck 600 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 1/2 - 17" or more.
In Figures 10a, 10b and 10c, there is an alternate embodiment of soft spring
rate,
long wheelbase three piece truck, identified as 660. Truck 660 employs
constant force
inboard and outboard, fore and aft pairs of friction dampers 666 mounted in
the distal
ends of truck bolster 668. In this arrangement, springs 670 are mounted
horizontally in
pockets in the distal ends of truck bolster 668 and urge, or bias, each of the
friction
dampers 666 against the corresponding friction surfaces of the vertical
columns of side
frames 664.
The spring force on friction damper wedges 640, 641, 642 and 643 varies as a
function of the vertical displacement of truck bolster 602, since they are
driven by the
CA 02860202 2016-12-20
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vertical springs of spring group 605. By contrast, the deflection of springs
670 does not
depend on vertical compression of the main spring group 672, but rather is a
function of
an initial pre-load. Although the arrangement of Figure 10a, 10b and 10c still
provides
inboard and outboard dampers and independent springing of the dampers, the
embodiment of Figures 9b is preferred to that of Figures 6a, 6b and 6c.
Damper Variations
Figures ha and lib show a partial isometric view of a truck bolster 680 that
is
generally similar to truck bolster 600 of Figure 9a, except insofar as bolster
pocket 682
does not have a central partition like web 652, but rather has a continuous
bay extending
across the width of the underlying spring group, such as spring group 636. A
single wide
damper wedge is indicated as 684. Damper 684 is of a width to be supported by,
and to
be acted upon, by two springs 686, 688 of the underlying spring group. In the
event that
bolster 680 may tend to deflect to a non-perpendicular orientation relative to
the
associated side frame, as in the parallelogramming phenomenon, one side of
wedge 684
will tend to be squeezed more tightly than the other, giving wedge 684 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 686, 688, yielding a restoring moment both
to the
twisting of wedge 684 and to the non-square displacement of truck bolster 680
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 llb is an alternate pair of damper wedges 690, 692.
This
dual wedge configuration can similarly seat in bolster pocket 682, and, in
this case, each
wedge 690, 692 sits over a separate spring. Wedges 690, 692 are in a side-by-
side
independently displacable vertically slidable relationship relative to each
other along the
primary angle of the face of bolster pocket 682. When the truck moves to an
out of
square condition, differential displacement of wedges 690, 692 may tend to
result in
differential compression of their associated springs, e.g., 686, 688 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
CA 02860202 2016-12-20
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pockets. To alleviate, or ameliorate, this situation, consumable wear plates
694 can be
mounted in bolster pocket 682 (with appropriate dimensional adjustments) as in
Figure
11b. Wear plates 694 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
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 696 in
Figure
16a.
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 9a, lla
and 17a,
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 12 in which
a
damper wedge is shown generically as 700. The replaceable, consumable wear
members
are indicated as 702, 704. The wedges and wear members have mating male and
female
mechanical interlink features, such as the cross-shaped relief 703 formed in
the primary
angled and vertical faces of wedge 700 for mating with the corresponding
raised cross
shaped features 705 of wear members 702, 704. Sliding wear member 702 is
preferably
made of a non-metallic, low friction material.
Although Figure 12 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 ha. 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 706 indicated in Figure ha.
The underside of the wedges described herein, wedge 700 being typical in this
regard, has a seat, or socket 707, for engaging the top end of the spring
coil, whichever
CA 02860202 2016-12-20
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spring it may be, spring 762 being shown as typically representative. Socket
707 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 16a as item 708.
Thus far only primary angles have been discussed. Figure 11c shows an
isometric
view of an end portion of a truck bolster 710, generally similar to bolster
600. As with all
of the truck bolsters shown and discussed herein, bolster 710 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 railcar
longitudinal center
line). Bolster 710 has a pair of spaced apart bolster pockets 712, 714 for
receiving
damper wedges 716, 718. Pocket 712 is laterally inboard of pocket 714 relative
to the
side frame of the truck more generally. Consumable wear plate inserts 720, 722
are
mounted in pockets 712, 714 along the angled wedge face.
As can be seen, wedges 716, 718 have a primary angle, a as measured between
vertical sliding face 724, (or 726, as may be) and the angled vertex 728 of
outboard face
730. 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 716 (or 718) when taken in any plane parallel to the plane of
outboard end
face 730. This same angle a is matched by the facing surface of the bolster
pocket, be it
712 or 714, and it defines the angle upon which displacement of wedge 716, (or
718) is
intended to move relative to that surface.
A secondary angle p gives the inboard, (or outboard), rake of the hypotenuse
surface of wedge 716 (or 718). 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 730. 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 730 of outboard wedge 718 outboard against
the
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opposing outboard face of bolster pocket 714. Similarly, the inboard face of
wedge 716
will tend to be biased toward the inboard planar face of inboard bolster
pocket 712.
These inboard and outboard faces of the bolster pockets are preferably lined
with a low
friction surface pad, indicated generally as 732. The left hand and right hand
biases of
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 710 includes a middle land 734 between pockets 712, 714, against which
another spring 736 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 11c, with
or without
(though preferably with) wear inserts.
In the case where a central land, such as land 734 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 noimal 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 11a, lib and lie are arranged in order of incremental
increases in complexity. The Example of Figure lid again provides a further
incremental increase in complexity. Figure lid shows a bolster 740 that is
similar to
bolster 710 except insofar as bolster pockets 742, 744 each accommodate a pair
of split
wedges 746, 748. Pockets 742, 744 each have a pair of bearing surfaces 750,
752 that are
inclined at both a primary angle and a secondary angle, the secondary angles
of surfaces
750 and 752 being of opposite hand to yield the damper separating forces
discussed
above. Surfaces 750 and 752 are also provided with linings in the nature of
relatively
low friction wear plates 754, 756. Each of pockets 742 and 744 accommodates a
pair of
split wedges 758, 760. Each pair of split wedges seats over a single spring
762. Another
spring 764 bears against central land 766.
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The example of Figure 13a shows a combination of a bolster 770 and biased
split
wedges 772, 774. Bolster 770 is the same as bolster 740 except insofar as
bolster pockets
776, 778 are stepped pockets in which the steps, e.g., items 780, 782, 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 11d, which
are left and
right handed. Thus the outboard pair of split wedges 784 has a first member
786 and a
second member 788 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 770). Similarly, the
inboard pair
of split wedges 790 has a first member 792 and a second member 794 each having
primary angle a, and secondary angle 13, except that the sense of secondary
angle f3 is in
the opposite direction such that members 792 and 792 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 13c, a replaceable monolithic
stepped wear insert 796 is welded in the bolster pocket 780 (or 782 if
opposite hand, as
the case may be). Insert 796 has the same primary and secondary angles a and
13 as the
split wedges it is to accommodate, namely 786, 788 (or, opposite hand, 792,
794). When
installed, and working, the more outboard of the wedges, 788 (or, opposite
hand, the
more inboard of the wedges 792) has a vertical and longitudinally planar
outboard face
800 that bears against a similarly planar outboard face 802 (or, opposite
hand, inboard
face 804) 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 780. The
hypotenuse face 806 of member 788 bears against the opposing outboard land 810
of
insert 796. The overall width of outboard member 788 is greater than that of
outboard
land 810, such that the inboard planar face of member 788 acts as an abutment
face to
fend inboard member 786 off of the surface of the step 812 in insert 796.
In similar manner inboard wedge member 786 has a hypotenuse face 814 that
bears against the inboard land portion 816 of insert 796. The total width of
bolster pocket
780 is greater than the combined width of wedge members, such that a gap is
provided
between the inboard (non-contacting) face of member 786 and the inboard planar
face of
pocket 780. The same relationship, but of opposite hand, exists between pocket
782 and
members 792, 794. The interface between member 786 and 788 is identified as
face 808.
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In an optional embodiment, a low friction pad, or surfacing, can be used at
the
interface of members 786, 788 (or 792, 794) to facilitate sliding motion of
the one
relative to the other.
In this arrangement, working of the wedges, i.e., members 786, 788 against the
face of insert 796 will tend to cause both members to move in one direction,
namely to
their most outboard position. Similarly, members 792 and 794 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 ajiii 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 13b and 13d, a single,
stepped
wedge 820, or 822, is used in place of the pair of split wedges e.g., members
786, 788. A
corresponding wedge of opposite hand is used in the other bolster pocket.
In the further alternative embodiment of Figures 14a, a truck bolster 830 has
welded bolster pocket inserts 832 and 834 of opposite hands welded into
accommodations in its distal end. In this instance, each bolster pocket has an
inboard
portion 836 and an outboard portion 838. Inboard and outboard portions 836 and
838
share the same primary angle a, but have secondary angles p that are of
opposite hand.
Respective inboard and outboard wedges are indicated as 840 and 842, and each
seats
over a vertically oriented spring 844, 846. In this case bolster 830 is
similar to bolster
680 of Figure 11a, to the extent that the bolster pocket is continuous ¨ there
is no land
separating the inner and outer portions of the bolster pocket. Bolster 830 is
also similar
to bolster 710 of Figure 11c, except that rather than the bolster pockets of
opposite hand
being separated, they are merged without an intervening land.
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In the further alternative of Figure 14b, split wedge pairs 848, 850 (inboard)
and
852, 854 (outboard) are employed in place of the single inboard and outboard
wedges
840 and 842.
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 14c and 14d.
Figure 15a shows the placement of a low friction bearing pad for bolster 680
of
Figure lla, 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
15a, the
truck bolster is identified as item 860 and the side frame is identified as
item 862. Side
frame 862 is symmetrical about the truck centerline, indicated as 864. Side
frame 862
has side frame columns 868 that locate between the inner and outer gibs 870,
872 of truck
bolster 860. The spring group is indicated generally as 874, 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 876, 877, 878, 879 seat
directly under
the bolster end 880. The end springs of each row, 881, 882, 883, 884 seat
under
respective friction damper wedges 885, 886, 887, 888. Consumable wear plates
889, 890
are mounted to the wide, facing flanges 891, 892 of the side frame columns,
888. As
shown in Figure 15b, plates 889, 890 are mounted centrally relative to the
side frames,
beneath the juncture of the side frame arch 892 with the side frame columns.
The lower
longitudinal member of the side frame, bearing the spring seat, is indicated
as 894.
Referring now to Figure 15c and 15e, bolster 860 has a pair of left and right
hand,
welded-in bolster pocket assemblies 900, 902, each having a cast steel,
replaceable,
welded-in wedge pocket insert 904. Insert 904 has an inboard-biased portion
906, and an
outboard-biased portion 908. Inboard end spring 882 (or 881) bears against an
inboard-
biased split wedge pair 910 having members 912, 914, and outboard end spring
884 (or
883) bears against an outboard- biased split wedge pair 916 having members
918, 920.
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.
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Each insert portion 906, 908 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 910, inboard leading member 912 has an inboard planar
face 924,
that, in use, is intended slidingly to contact the opposed vertically planar
face of the
bolster pocket. Leading member 912 has a bearing face 926 having primary angle
a and
secondary angle P. Trailing member 914 has a bearing face 928 also having
primary
angle a and secondary angle 0, and, in addition, has a transition, or step,
face 930 that has
a primary angle a and a tertiary angle cp.
Insert 904 has a corresponding array of bearing surfaces having a primary
angle a,
and a secondary angle 13, with transition surfaces having tertiary angle cp
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 cp is relatively
narrow and
steep.
In Figure 15e, it can be seen that the sloped portions of split wedge members
918,
920 extend only partially far enough to overlie a coil spring 926. In
consequence, wedge
members 918 and 920 each have a base portion 927, 929 having a fore-and-aft
dimension
greater than the diameter of spring 926, and a width greater than half the
diameter of
spring 926. Each of base portions 927, 929 has a downwardly proud, roughly
semi-
circular boss 932 for seating in the top of the coil of spring 926. The
upwardly angled
portion 934, 936 of each wedge member 918, 920 is extends upwardly of base
portion
927, 929 to engage the matingly angled portions of insert 904.
In a further alternate embodiment, the split wedges can be replaced with
stepped
wedges 940 of similar compound profile, as shown In Figure 15f. 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 15g shows a welded in insert 942 having a profile for mating
engagement with the corresponding wedge faces.
Figures 16a and 16b illustrate a bolster, side frame and damper arrangement in
which dampers 950, 951 are independently sprung on horizontally acting springs
952,
953 housed in side-by-side pockets 954, 955 in the distal end of bolster 960.
Although
CA 02860202 2016-12-20
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only two dampers are shown, it will be understood that a pair of dampers faces
toward
each of the opposed side frame columns. Dampers 950, 951 each include a block
968
and a consumable wear member 962, the block and wear member having male and
female indexing features 964 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 978 is provided in
the spring
housing to permit the spring to be pre-loaded and held in place during
installation.
Figures 17a, 17b and 17c show a preferred truck 970, having a bolster 972, a
side
frame 974, a spring group 976, and a damper arrangement 978. The spring group
has a 5
x 3 arrangement, with the dampers being in a spaced arrangement generally as
shown in
Figure 11c, 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 11c.
In one embodiment of truck 970, as might preferably be used in the location of
end trucks 88, 206, 210, or 232, there may be a 5 x 3 spring group
arrangement, the
spring group including 11 coils each having a spring rate in the range of 550
¨ 650 lb./in,
and most preferably about 580 lb./in; and 4 springs (under the dampers, in a
four corner
arrangement) having a spring rate in the range of 450 ¨ 550 lb./in, most
preferably about
500 lb./in, for which the dampers are driven by 20 ¨ 25 % of the force of the
spring
group, preferably about 24 %. The dampers may have a primary angle of 35 - 45
deg.,
preferably about 40 deg. In this preferred end truck embodiment, the overall
group
vertical spring rate is in the range of 8,000 to 8,500 lb./in., in particular
about 8,380 lb./in.
In another embodiment of truck 970, such as might preferably be used in the
location of internal truck 234, there may be a 5 x 3 spring group arrangement
in which
the spring group may include 11 outer springs having a spring rate of about
550 ¨ 650
lb./in., and most preferably about 580 lb./in; 4 springs (under the dampers,
in a four
corner arrangement) having a spring rate in the range of 550 ¨ 650 lb./in, and
most
preferably about 600 lb./in.; and six inner coils having a spring rate in the
range of 250 ¨
300 lb./in., most preferably about 280 lb./in. The overall spring rate for the
5 x 3 group is
in the range of 10,000 ¨ 11,000 lb./in., and most preferably about 10,460
lb./in. The
dampers are driven by about 20 ¨ 25 % of the total force of the spring group,
preferably
about 23 %. The dampers have a primary angle in the range of 35 ¨ 35 degrees,
preferably about 40 degrees.
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It will be appreciated that the values and ranges given for truck 970 depend
on the
expected empty weight of the railcar, the expected lading, the natural
frequency range to
be achieved, the amount of damping to be achieved, and so on, and may
accordingly vary
from the preferred ranges and values indicated above.
In the embodiments of Figures 2a, 2b, 3a, 3b, 4a and 4b, the ratio of the dead
sprung weight, WD, of the railcar 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 railcar of Figures 2a, 2b, 3a, 3b, 4a and 4b the
ratio can
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.
The embodiments described herein 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 railcars of Figures
2a, 2b, 3a,
3b, 4a and 4b is 1.5 Hz or less, with the unladen vertical bounce natural
frequency being
less 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 9a, 11a, and 17a, 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 1/4" of lateral
travel of the
CA 02860202 2016-12-20
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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 9a, ha and 17a, 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
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 although high
density lading
may tend to require a stiffer vertical response than automobiles. 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.
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It will be understood that the features of the trucks of Figures 6a, 6b, 7a,
7b, 8,
and 9a, 9f are provided by way of illustration, and that the features of the
various trucks
can be combined in many different permutations and combinations. That is, a 2
x 4
spring group could also be used with a single wedge damper per side. Although
a single
wedge damper per side arrangement is shown in Figures 6a and 7a, a double
damper
arrangement, as shown in Figures 8 and 9a may tend to provide enhanced
squaring of the
truck and resistance to hunting. A 3 x 3 or 3 x 5, or other arrangement spring
set may be
used in place of either a 3:2:3 or 2 x 4 spring set, with a corresponding
adjustment in
spring seat plate size and layout. Similarly, the trucks can use a wide
sideframe window,
and corresponding extra long wheel base, or a smaller window. Further, each of
the
trucks could employ a rocking bottom spring seat, as in Figure 6b, or a fixed
bottom
spring seat, as in Figure 7a, 8 or 9a.
As before, the upper rocker seats are inserts, typically of a hardened
material,
whose rocking, or engaging surface 480 has a radius of curvature of about five
inches,
with the center of curvature (when assembled) lying above the upper rockers
(i.e., the
surface is upwardly concave).
In each of the trucks shown and described herein, for a fully laden car type,
the
lateral stiffness of the sideframe acting as a pendulum is less than the
lateral stiffness of
the spring group in shear. In one embodiment, the vertical stiffness of the
spring group is
less than 12,000 Lbs./in, with a horizontal shear stiffness of less than 6,000
Lbs./in. The
pendulum has a vertical length measured (when undeflected) from the rolling
contact
interface at the upper rocker seat to the bottom spring seat of between 12 and
20 inches,
preferably between 14 and 18 inches. The equivalent length Leq, may be in the
range of 8
to 20 inches, depending on truck size and rocker geometry, and is preferably
in the range
of 11 to 15 inches, and is most preferably between about 7 and 9 inches for 28
inch
wheels (70 ton "special"), between about 8 1/2 and 10 inches for 33 inch
wheels (70 ton),
9 1/2 and 12 inches for 36 inch wheels (100 or 110 ton), and 11 and 13 1/2
inches for 38
inch wheels (125 ton). Although truck 520 or 600 may be a 70 ton special, a 70
ton, 100
ton, 110 ton, or 125 ton truck, it is preferred that truck 520 or 600 be a
truck size having
33 inch diameter, or even more preferably 36 or 38 inch diameter wheels.
In the trucks described herein according to the present invention, Lresultant,
as
defined above, is greater than 10 inches, is advantageously in the range of 15
to 25
inches, and is preferably between 18 and 22 inches, and most preferably close
to about 20
I]
CA 02860202 2016-12-20
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inches. In one particular embodiment it is about 19.6 inches, and in another
particular
embodiment it is about 19.8 inches.
In the trucks described herein, for their fully laden design condition which
may be
determined either according to the AAR limit for 70, 100, 110 or 125 ton
trucks, or,
where a lower intended lading is chosen, then in proportion to the vertical
sprung load
yielding 2 inches of vertical spring deflection in the spring groups, the
equivalent lateral
stiffness of the sideframe, being the ratio of force to lateral deflection
measured at the
bottom spring seat, is less than the horizontal shear stiffness of the
springs. The
equivalent lateral stiffness of the sideframe ksideliame is less than 6,000
Lbs./in. and
preferably between about 3,500 and 5,500 Lbs./in., and more preferably in the
range of
3,700 ¨ 4,100 Lbs./in. By way of an example, in one embodiment a 2 x 4 spring
group
has 8 inch diameter springs having a total vertical stiffness of 9,600 Lbs./
in. per spring
group and a corresponding lateral shear stiffness kspring shear of 4,800
lbs./in. The
sideframe has a rigidly mounted lower spring seat. It is used in a truck with
36 inch
wheels. In another embodiment, a 3 x 5 group of 5 1/2 inch diameter springs is
used, also
having a vertical stiffness of about 9,600 lbs./in. in a truck with 36 inch
wheels. It is
intended that the vertical spring stiffness per spring group be in the range
of less than
30,000 lbs./in., that it advantageously be in the range of less than 20,000
lbs./in and that it
preferably be in the range of 4,000 to 12,000 lbs./in, and most preferably be
about 6,000
to 10,000 lbs./in. The twisting of the springs has a stiffness in the range of
750 to 1,200
lbs./in. and a vertical shear stiffness in the range of 3,500 to 5,500
lbs./in. with an overall
sideframe stiffness in the range of 2,000 to 3,500 lbs./in.
In the embodiments of trucks in which there is a fixed bottom spring seat, the
truck may have a portion of stiffness, attributable to unequal compression of
the springs
equivalent to 600 to 1200 Lbs./in. of lateral deflection, when the lateral
deflection is
measured at the bottom of the spring seat on the sideframe. Preferably, this
value is less
than 1000 Lbs./in., and most preferably is less than 900 Lbs./in. The portion
of restoring
force attributable to unequal compression of the springs will tend to be
greater for a light
car as opposed to a fully laden car, i.e., a car laden in such a manner that
the truck is
approaching its nominal load limit, as set out in the 1997 Car and Locomotive
Cyclopedia at page 711.
The double damper arrangements shown above can also be varied to include any
of the four types of damper installation indicated at page 715 in the 1997 Car
and
CA 02860202 2016-12-20
- 59 -
Locomotive Cyclopedia, with appropriate structural changes for doubled
dampers, with
each damper being sprung on an individual spring. That is, while inclined
surface bolster
pockets and inclined wedges seated on the main springs have been shown and
described,
the friction blocks could be in a horizontal, spring biased installation in a
pocket in the
bolster itself, and seated on independent springs rather than the main
springs.
Alternatively, it is possible to mount friction wedges in the sideframes, in
either an
upward orientation or a downward orientation.
The embodiments of trucks shown and described herein may vary in their
suitability
for different types of service. Truck performance can vary significantly based
on the
loading expected, the wheelbase, spring stiffnesses, spring layout, pendulum
geometry,
damper layout and damper geometry.
The principles of the present invention are not limited to autorack 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 or scope of the invention, the invention is not to
be limited to
those details, but only by a purposive reading of the claims as required by
law.
