Note: Descriptions are shown in the official language in which they were submitted.
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RAIL ROAD FREIGHT CAR WITH RESILIENT SUSPENSION
Field of the Invention
This invention relates to the field of auto rack rail road cars for carrying
motor
vehicles.
Background of the Invention
Auto rack rail road cars are used to transport automobiles. Most often,
although not always, they are used to transport finished automobiles from a
factory or
a port to a distribution center. Typically, auto-rack rail road cars are
loaded in the
"circus loading" manner, by driving vehicles into the cars from one end, and
securing
them in places with chocks, chains or straps. When the trip is completed, the
chocks
are removed, and the cars are driven out.
Automobile manufacturers would like to be able to have new cars driven into
the auto-rack cars, and then to be held in place using the parking brake of
the car
alone, without the need for chocks or chains. At present the operating
characteristics
of auto-rack cars are not generally considered to be gentle enough to permit
this do be
done reliably. That is, a long standing concern has been the frequency of
damage
claims arising from high accelerations imposed on the lading during train
operation.
It has been suggested that the maximum design load condition of some
automobile
components occurs during the single journey of the automobile on the rail car.
Damage due to dynamic loading in the railcar may tend to arise principally in
two ways. First, there are the longitudinal input loads transmitted through
the draft
gear due to train line action or shunting. Second, there are vertical, rocking
and
transverse dynamic responses of the rail road car to track perturbations as
transmitted
through the rail car suspension.
In this context, slack includes (a) the free slack in the couplers; and (b)
the
travel of the draft gear of successive rail road cars under the varying buff
and draft
loads. Slack run-out occurs, for example, as a train climbs a long upgrade,
and all of
the slack is taken out of the couplings as the train stretches. Once the train
clears the
crest, and begins its descent, the rail road cars at the end of the train may
tend to
accelerate downhill into the cars in front, closing up the slack. This slack
run-in and
run-out can result in significant longitudinal accelerations. These
accelerations are
transmitted to the automobiles carried in the auto-rack cars.
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Historically, the need for slack was related, at least in part, to the
difficulty of
using a steam locomotive to "lift" (that is, move from a standing start) a
long string of
cars with journal bearings, particularly in cold weather. Steam engines were
reciprocating piston engines whose output torque at the drive wheels varied as
a
function of crank angle. By contrast, presently operating diesel-electric
locomotives
are capable of producing high tractive effort from a standing start, without
concern
about crank angle or wheel angle. For practical purposes, presently available
diesel-
electric locomotives are capable of lifting a unit train of one type of cars
having little
or no slack.
Switching is another process having a long history. Two common types of
switching are "flat switching" and "humping". Humping involves running freight
cars successively over a raised portion of track, and then allowing the car to
run
down-hill under gravity along various leads and sidings to couple with other
cars as a
train consist is assembled. For this type of operation the coupling speeds can
be
excessive, resulting in similarly excessive car body accelerations. For many
types of
rail road car, humping is now forbidden due to the probability of damaging the
lading.
An alternate form of switching is "flat switching" in which a locomotive is
used to
give a push to a rail road car, and then to send it rolling under its own
inertia down a
chosen siding to couple with another car. Particularly when done at night, the
desirability of making sure that a good coupling is made tends to encourage
rail yard
personnel to make sure that the rail road cars are given an extra generous
push. This
often less than gentle habit tends to lead to rather high impact loads during
coupling at
impacts in the 5 m.p.h. (or higher) range. Forces can be particularly severe
when
there is an impact between a low density lading rail road car, such as an auto
rack car,
and a high density lading car (or string of cars) such as coal or grain cars.
Given this history, rail road car draft gear are designed to cope with slack
run-
out and slack run-in during train operation, and also to cope with 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 Lb.
(roughly
2.2 x 106 N). Typically, these draft gear have a travel of 2 3/4 to 3 '/4
inches in buff
before reaching the 500,000 Lb. 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
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lading, increases sharply. While this may be acceptable for coal or grain, it
is
undesirably severe for more sensitive lading, such as automobiles or auto
parts, paper,
and other high value consumer goods such as household appliances.
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. Draft
gear
development has tended to be directed toward providing longer travel on impact
to
reduce the peak acceleration. In the development of sliding sills, and
latterly,
hydraulic end of car cushioning (EOCC) units, the same impact is accommodated
over 10, 15, or 18 inches of travel. As a result, for example, by the end of
the 1960's
nearly all auto rack cars, and other types of special freight cars had EOCC
units.
Further, of the approximately 45,000 auto-rack cars in service in 1997,
virtually all
were equipped with end of car cushioning units. A discussion of the
developments of
couplers, draft gear and EOCC equipment is given the 1997 Car and Locomotive
Cyclopedia (Simmons-Boardman Books, Inc., Omaha, 1997 ISBN 0-9 1 1 3 82-20-8)
at pp. 640 - 702. In summary, there has been a long development of long travel
draft
gear equipment to protect relatively fragile lading from end impact loads.
In light of the foregoing, it is counter-intuitive to employ short-travel, or
ultra
short travel, draft gear for carrying wheeled vehicles. However, by
eliminating, or
reducing, the accumulation of slack, the use of short travel buff gear may
tend to
reduce the relative longitudinal motion between adjacent rail road cars, and
may tend
to reduce the associated velocity differentials and accelerations between
cars. The use
of short travel, or ultra-short travel, buff gear also has the advantage of
eliminating the
need for relatively expensive, and relatively complicated EOCC units, and the
fittings
required to accommodate them. This may tend to permit savings both at the time
of
manufacture, and savings in maintenance during service.
Further, as noted above, given the availability of locomotives that develop
continuous high torque from a standing start, it is possible to re-examine the
issue of
slack action from basic principles. The use of vehicle carrying rail road cars
in unit
trains that will not be subject to operation with other types of freight cars,
that will not
be subject to flat switching, and that may not be subject to switching at all
when
loaded, provides an opportunity to adopt a short travel, reduced slack
coupling system
throughout the train. The conventional approach has been to adopt end of car
equipment with sufficient travel to cope with existing slack accumulation
between
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cars. In doing so, the long travel end of car equipment has tended to add to
the range
of slack action in the train that is to be accommodated by the draft gear
along the
train. The opposite approach is to avoid a large accumulation of slack in the
first
place. If a large amount of slack is not allowed to build up along the train,
then the
need for long-travel draft gear and other end of car equipment is also
reduced, or,
preferably, eliminated.
One way to reduce slack action is to use fewer couplings. To that end, since
articulated connectors are slackless, use of articulated rail road cars
significantly
reduces the slack action in the train. Some releasable couplings are still
necessary, to
permit the composition of a train to change, if desired. Further, it is
necessary to be
able to change out a car for repair or maintenance when required.
To reduce overall slack, it would be advantageous to adopt a reduced slack, or
slackless, coupler, (as compared to AAR Type E). Although reduced slack AAR
Type F couplers have been known since the 1950's, and slackless "tightlock"
AAR
Type H couplers became an adopted standard type on passenger equipment in
1947,
AAR Type E couplers are still predominant. AAR Type H couplers are expensive,
(and are used for passenger cars), as were the alternate standard Type CS
controlled
slack couplers. According to the 1997 Cyclopedia, supra, at p. 647 "Although
it was
anticipated at one time that the F type coupler might replace the E as the
standard
freight car coupler, the additional cost of the coupler and its components,
and of the
car structure required to accommodate it, have led to its being used primarily
for
special applications". One "special application" for F type couplers is in
tank cars,
another is in rotary dump coal cars.
The difference between the nominal 3/8" slack of a Type F coupler and the
nominal 25/32" slack of a Type E coupler may seem small in the context of EOCC
equipped cars having 10, 15 or 18 inches of travel. By contrast, that
difference,
13/32", seems proportionately larger when viewed in the context of the
approximately
11/16" buff compression (at 700,000 lbs.) of Mini-BuffGear. It should be noted
that
there are many different styles of Type E and Type F couplers, whether short
or long
shank, whether having upper or lower shelves, as described in the Cyclopedia,
supra.
There is a Type E/F having a Type E coupler head and a Type F shank. There is
a
Type E50ARE knuckle which reduces slack from 25/32 to 20/32". Type F herein is
intended to include all variants of the Type F series, and Type E herein is
intended to
include all variants of the Type E series having 20/32" of slack or more.
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Another way to reduce slack action in the draft gear is to employ stiffer
draft
gear. Short travel draft gear are presently available. As noted above, most M-
901-G
draft gear have an official rating travel of 2 3/4" to 3 1/4" under a buff
load of 500,000
Lbs. Mini-BuffGear, as produced by Miner Enterprises Inc., of 1200 State
Street,
Geneva Illinois, appears to have a displacement of less than 0.7 inches at a
buff load
of over 700,000 lbs., and a dynamic load capacity of 1.25 million pounds at 1
inch
travel. This is nearly an order of magnitude more stiff than some M-901-G
draft gear.
Miner indicates that this "special BuffGear gives drawbar equipped rail cars
and trains
improved lading protection and train handling", and further, "[The resilience
of the
Mini-BuffGear] reduces the tendency of the draw bar to bind while negotiating
curves. At the same time, the Mini-BuffGear retains a high pre-load to reduce
slack
action. Elimination of slack between coupler heads, plus Mini-Buff Gear's high
pre-
load and limited travel, provide ultralow slack coupling for multiple-unit
well cars
and drawbar connected groups of unit train coal cars." Notably, unlike vehicle
carrying rail cars, coal is unlikely to be damaged by the use of short travel
draft gear.
In addition to M-901-G draft gear, and Mini-BuffGear, it is also possible to
obtain draft gear having less than 1 3/4 inches of deflection at 400,000 Lbs.,
one type
having about 1.6 inches of deflection at 400,000 Lbs. This is a significant
difference
from most M-901-G draft gear.
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 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 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
conicity
of the wheels tends not only to give the trucks a measure of self-steering
ability, but
tends also to cause the truck to oscillate transversely between the rails.
During
hunting, the trucks tend most often to deform in a parallelogram manner.
Lateral
perturbations in the rails sometimes arise where the rails widen or narrow
slightly, or
one rail is more worn than another, and so on.
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There are both geometric and historic factors to consider related to these
loading conditions. One is the near universal usage of the three-piece style
of freight
car truck in North America. While other types of truck are known, such as an H-
frame truck or single axle fixed truck as used in Europe, the three piece
truck has
advantages that have made it overwhelmingly dominant in freight service in
North
America. First, it can carry greater loads than a fixed, single axle truck,
and permits
greater longitudinal truck spacing than a single axle truck. The three piece
truck is
simple. It employs only three main component elements, namely a truck bolster
and a
pair of side frames. The side frame castings are inexpensive relative to
alternative H-
frame designs. Manufacture of the side frame requires a relatively small mold
as
compared to an H-frame truck, and may tend to be less prone to molding
defects. 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. The three piece
truck can
operate in a wide range of environmental conditions, over a long period of
time, with
relatively little maintenance. When maintenance is required, the springs and
axles can
be changed out relatively easily. In terms of 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. By contrast, an H frame truck requires both a
primary
suspension and secondary suspension at each of the wheels. In summary, 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 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 must tend not to fall
outside a
given range, typically about 2 inches, if the couplers are to perform
satisfactorily in
interchange service. In addition, rail road car suspensions have a dynamic
range in
operation, including a reserve 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 full loaded general purpose flat
car. In each
case, the design lading tended to be very heavy relative to the rail car
weight. 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
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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
damaged badly by excessive vibration. In addition, coal and grain tend to have
a
relatively low value per unit weight. 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 two reasons. First, the weight
to be back
hauled empty is kept low, reducing the fuel cost of the backhaul. Second, when
the
ratio of lading to car weight increases, a higher proportion of hauling effort
goes into
hauling lading, as opposed to hauling the deadweight of the railcars
themselves.
By contrast, an autorack car has the opposite loading profile. A two unit
articulated autorack car as presently in service may have a light car weight
of 165,000
lbs., and a lading weight when fully loaded of only 35 - 40,000 lbs. The
lading
typically has a high, or very high, ratio of value to weight. Generally, while
coal may
account for as much as 40 % of all car loadings, it may generate only about 25
% of
freight revenues. By comparison, automobiles may account for only about 2 % of
car
loadings, yet may account for about 10 % of freight revenues. Similarly,
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 incur a greater expense to obtain superior ride quality to
that
suitable for coal or grain.
Historically auto rack 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. 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. 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.
Deliberately increasing the mass of any kind of freight car is, itself,
counter
intuitive, since many years of effort has gone into reducing the weight of
rail cars
relative to the weight of the lading for the reasons noted above. One
manufacturer,
for example, advertises a light weight aluminium auto-rack 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, auto rack rail cars tend to be tall,
long, and thin,
with the upper deck loads carried at a relatively high location as measured
from top of
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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.
Decreasing the spring rate involves further considerations. Historically the
deck height of a flat car tended to be very closely related to the height of
the upper
flange of the center sill. This height was itself established by the height of
the cap of
the draft pocket. The size of the draft pocket was standardised on the basis
of the
coupler chosen, and the allowable heights for the coupler knuckle. The deck
height
usually worked out to about 40 or 41 inches above top of rail. For some time
auto
rack cars were designed to a 19 ft height limit. To maximise the internal
loading
space, it has been considered desirable to lower the main deck as far as
possible,
particularly in tri-level cars. Since the lading is relatively light, the
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. Since the
American
Association of Railroads (AAR) specifies a minimum clearance of 5" above the
wheels, the combination of low deck height, deck clearance, and minimum wheel
height set an effective upper limit on the spring travel, and reserve spring
travel range
available. If softer springs are used, the remaining room for spring travel
below the
decks may well not be sufficient to provide the desired reserve height. In
consequence, the present inventor proposes, contrary to lowering the main
deck, that
the main deck be higher than 42 inches to allow for more spring travel.
As noted above, many previous auto rack cars have been built to a 19 ft
height. Another major trend in recent years has been the advent of "double
stack"
intermodal container cars capable of carrying two shipping containers stacked
one
above the other in a well or to other freight cars falling within the 20 ft 2
in. height
limit of AAR plate F. Many main lines have track clearance profiles that can
accommodate double stack cars. Consequently, it is now possible to use auto
rack
cars built to the higher profile of the double stack intermodal container
cars. The
present inventor has chosen to increase the height of the car generally to
provide both
a suitable internal height for the lading, and to permit the use of softer
springs.
While decreasing the primary vertical bounce natural frequency appears to be
advantageous for auto rack rail road cars generally, including single car unit
rail road
cars, articulated auto rack cars may also benefit not only from adding
ballast, but from
adding ballast preferentially to the end units near the coupler end trucks. As
explained more fully in the description below, the interior trucks of
articulated cars
tend to be more heavily burdened than the end trucks, primarily because the
interior
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trucks share loads from two adjacent car units, while the couple end trucks
only carry
loads from one end of one car. There are a number of reasons why it would be
advantageous to even out this loading so that the trucks have roughly similar
vertical
bounce frequencies.
Three piece trucks currently in use tend to use friction dampers, sometimes
assisted by hydraulic dampers such as can be mounted, for example, in the
spring set.
Friction damping has most typically been provided by using spring loaded
blocks, or
snubbers, mounted with the spring set, with the friction surface bearing
against a
mating friction surface of the columns of the side frames, or, if the snubber
is
mounted to the side frame, then the friction surface is mounted on the face of
the
truck bolster. There are a number of ways to do this. In some instances, as
shown at
p. 847 of the 1984 Car & Locomotive Cyclopedia lateral springs are housed in
the end
of the truck bolster, the lateral springs pushing horizontally outward on
steel shoes
that bear on the vertical faces of the side columns of the side frames. This
provides
roughly constant friction (subject to the wear of the friction faces), without
regard to
the degree of compression of the main springs of the suspension.
In another approach, as shown at p. 715 of the 1997 Car & Locomotive
Cyclopedia, one of the forward springs in the main spring group, and one of
the
rearward springs in the main spring group bears 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
rail car. During hunting the side frames tend to want to rotate about a
vertical axis to
a non-perpendicular angular orientation relative to the truck bolster
sometimes called
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"parallelogramming". This will tend to cause lateral deflection of the spring
group,
and will tend to generate a squeezing force on opposite diagonal sides of the
wedges,
causing them to tend to bear against the side frame columns. This diagonal
action
will tend to generate a restoring moment working against the angular
deflection. The
moment arm of this restoring force is proportional to half the width of the
wedge,
since half of the friction plate lies to either side of the centreline of the
side frame.
This tends to be a relatively weak moment connection, and the wedge, even if
wider
than normal, tends to be positioned over a single spring in the spring group.
Typically, for a truck of fixed wheelbase length, there is a trade-off between
wheel load equalisation and resistance to hunting. Where a car is used for
carrying
high density commodities at low speeds, there may tend to be a higher emphasis
on
maintaining wheel load equalisation. Where a car is light, and operates at
high speed
there will be a greater emphasis on avoiding hunting. In general the
parallelogram
deformation of the truck in hunting is deterred by making the truck laterally
more
stiff. Another method is to use a transom, typically in the form of a channel
running
from between the side frames below the spring baskets.
One way to raise the hunting threshold is to employ a truck having a longer
wheelbase, or one whose length is proportionately great relative to it width.
For
example, at present two axle truck wheelbases may range from about 5' - 3" to
6' -
0". However, the standard North America track gauge is 4' - 8 '/2", giving a
wheelbase to track width ratio possibly as small as 1.12. At 6' - 0" the ratio
is
roughly 1.27. It would be preferable to employ a wheelbase having a longer
aspect
ratio relative to the track gauge. As described herein, one aspect of the
present
invention employs a truck with a longer wheelbase, preferably about 86 inches,
giving
a ratio of 1.52. This increase in wheelbase length may tend also to be benign
in terms
of wheel loading equalisation.
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 two wedges, of comparable size to those previously used, the two wedges
being placed side by side and each 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.
Summary of the Invention
In an aspect of the invention there is a rail road freight car having at least
one
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rail car unit. The rail road freight car is supported by three piece rail car
trucks for
rolling motion along rail road tracks. Each of the three piece trucks has a
rigid truck
bolster and a pair of first and second side frame assemblies. The bolster has
first and
second ends and the side frames are mounted at either end of the truck
bolster. The
three piece trucks each have a resilient suspension mounted between the truck
bolster
and the side frames. The rail road freight car has a sprung mass. A first
portion of the
sprung mass is carried by a first of the rail car trucks. The resilient
suspensions of the
first of the trucks has a vertical bounce spring rate. The rail car truck
suspension has a
natural vertical bounce frequency. The frequency is the square root of the
value
obtained by dividing the first spring rate by the first portion of the sprung
mass. The
natural vertical bounce frequency of the rail road car is less than 4.0 Hz.
when the rail
road car is unloaded.
In an additional feature of that aspect of the invention, each of the trucks
bears
a respective portion of the sprung mass of the rail road car. Each of the
trucks has a
vertical bounce spring rate, and each respective natural vertical bounce
frequency of
each of the trucks is less than 3.0 Hz. when the rail road car is empty.
In another additional feature, each of the trucks bears a respective portion
of
the sprung mass of the rail road car. Each of the trucks has a vertical bounce
spring
rate, and the rail road car has an overall natural vertical bounce frequency
of less than
2.0 Hz. when the road car is empty.
In yet another additional feature, the first rail car truck has a gross rail
load
limit. The first rail car truck carries a first live load when the rail road
car is fully
loaded. The gross rail limit for the first truck is at least as great as the
first portion of
the rail car mass and the first live load when added together. The first rail
car truck
has a natural vertical bounce frequency less than 1.5 Hz. when the rail road
car is fully
loaded.
In still yet another additional feature, the rail road car has a fully loaded
live
load mass, and when fully loaded, the rail road car has a natural vertical
bounce
frequency of less than 1.5 hz. In a further additional feature, the rail road
car has a
natural vertical bounce frequency of less than 1.4 Hz. In still a further
additional
feature, the rail road car has at least one end-loading deck for carrying
wheeled
vehicles. In yet a further additional feature, the rail road car is an auto
rack car. In
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another additional feature, the rail road car is an articulated rail road car.
In still
another additional feature, the rail road car is a three pack auto rack rail
road car.
In yet another additional feature, the three pack autorack rail road car has a
center unit and first and second end units joined at articulated connectors to
the center
unit. The center unit has two of the trucks mounted thereunder, and each of
the end
units has a single one of the trucks mounted thereunder. The articulated
connectors
are longitudinally offset from the trucks mounted under the center unit.
In still yet another additional feature, the rail road car includes at least
one rail
car unit. The rail car unit has a light car weight and a fully loaded weight,
and the
light car weight is at least half as great as the fully loaded weight.
In still another additional feature, the rail road car is an articulated auto
rack
rail road car including at least two auto rack rail car units joined at an
articulated
connection. At least one of the auto rack rail car units is an end unit. The
end unit
has a sprung weight of at least 65,000 lbs.
In a further additional feature, the rail road car is an articulated rail road
car
including at least two rail car units joined at an articulated connection. At
least two of
the rail car units are first and second end units. Each end unit has a first
end having a
releasable coupler mounted thereto, and a second end connected by the
articulated
connection to an adjacent rail car unit. The first end unit has one of the
three piece
trucks mounted thereunder closer to the first end having the releasable
coupler than to
the second end joined by the articulated connector to the adjacent car. The
first end
unit has a weight, and a weight distribution of the weight biased toward the
coupler
end thereof.
In another additional feature, the end unit has at least one ballast member
mounted closer to the coupler end thereof than to the articulated connector
end
thereof. In still another additional feature, the ballast member is a deck
plate. In yet
another additional feature, as unloaded, at least 60 % of the weight is
carried by the
truck mounted closer to the coupler end than to the articulated connector end.
In still
yet another additional feature, as unloaded, at least 2/3 of the weight is
carried by the
truck mounted closer to the coupler end than to the articulated connector end.
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In a further additional feature, the rail road car has a three piece truck
mounted
closer to the articulation connection end of the end rail car truck than any
other truck
of the rail road car. When the rail road car is empty, the three piece truck
mounted
closer to the coupler end of the end car unit bears a dead sprung load D1. The
three
piece truck closest to the articulated connector bears a dead sprung load D2.
D1 lies
in the range of 2/3 of D2 to 4/3 of D2.
In still a further additional feature, D1 is in the range of 4/5 to 6/5 of D2.
In
another additional feature, D 1 is in the range of 90 % of D2 to 110 % of D2.
In still
another additional feature, the first three piece truck has a wheelbase of
greater than
72 inches. In yet another additional feature, the first three piece truck has
a wheelbase
of greater than 80 inches. In still yet another additional feature, the first
three piece
truck has a track width corresponding to a railroad gauge width, and a
wheelbase
length. The ratio of the wheelbase length to the gauge width is at least as
great as 1.3
: 1Ø In still another additional feature, the ratio is at least as great as
1.4 : 1Ø In
another additional feature, the first rail car truck has a set of wheels for
engaging a rail
road track. The rail road car has a body having a clearance above the wheels
of more
than 5 inches. In yet another additional feature, the clearance is at least 7
inches.
In still another additional feature, the car has a light weight corresponding
to a
first mass Ml when unloaded, and is rated to carry a live load of a maximum
mass
M2, and the ratio of M1 : M2 is at least as great as 1.2 : 1. In still yet
another
additional feature, the ratio is at least as great as 1.5 : 1. In a further
additional
feature, the rail road car has a deck for carrying lading above the first rail
car truck.
The deck for lading lies at a height of greater than 42 inches relative to top
of rail. In
yet a further additional feature, the first rail car truck has a rating at
least as great as
"70 Ton". In still a further additional feature, the car exceeds 19' - 0" in
height
measured from top of rail.
In still yet a further additional feature, the rail road car has a first
coupler end
and a second coupler end. A draft gear is mounted to the railcar at the first
coupler
end, and a releasable coupler is mounted to the draft gear. The draft gear has
a
deflection of less than 2 %2 inches under a buff load of 500,000 Lbs. In
another
additional feature, the resilient suspension includes a spring group mounted
between
one end of the truck bolster and one of the side frames, and a second spring
group
mounted between the other end of the truck bolster and the other side frame.
Each of
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the spring groups has a spring rate constant lying in the range of 6,000
lbs/in to
10,000 lbs/in. In yet another additional feature, the spring rate constant of
each of the
groups has a value lying in the range of 7000 lbs/in and 9500 lbs/in.
In another aspect of the invention there is a articulated rail road freight
car. At
least a first rail car unit and a second rail car unit is joined at an
articulated
connection. The articulated rail road car is carried by rail car trucks for
rolling motion
along rail road tracks. At least two of the rail car units are end units. The
first rail car
unit is one of the end units. The first end unit has a first end and a second
end. The
first end of the first rail car unit has a releasable couple mounted thereto
and the
second end is joined by the articulated connection to the second rail car
unit. A first
of the trucks is mounted to the first rail car unit at a first truck center.
The first truck
center lies closer to the first end of the first rail car unit than to the
second end. A
second of the trucks is mounted closer to the articulation between the first
and second
rail car units than any other of the trucks. The first car unit has a weight
and a dead
load weight distribution. The dead load weight distribution of the first rail
car unit is
biased toward the first end of the first rail car unit.
In an additional feature of that aspect of the invention, as empty, at least
60 %
of the weight of the first rail car unit is borne by the first truck. In
another additional
feature, as empty, at least 2/3 of the weight of the first rail car unit is
borne by the first
truck. In still another additional feature, the second rail car unit has a
weight
distributed between the second rail car truck and a third rail car truck. When
the rail
road car is empty, the first rail car truck bears a first dead load, D1. The
second rail
car truck bears a second dead load, D2, and Di is in the range of 2/3 to 4/3
of D2. In
yet another additional feature, D1 is in the range of 90 % to 110 % of D2.
In another aspect of the invention there is an articulated rail road freight
car
comprising a number of rail car units connected at a number of articulated
connectors.
The rail car units are supported for rolling direction along rail road tracks
by a number
of rail car trucks. The number of articulated connectors is one less than the
number of
railcar units. Each articulated connector is located between two adjacent ones
of the
rail car units. The number of rail car trucks is one greater than the number
of rail car
units. The rail car units each have a dead sprung weight. The dead sprung
weights of
the rail cars is distributed among the trucks. An average dead sprung weight
per
truck, WO, is equal to the total dead sprung weight of all of the rail car
units divided
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by the total number of the trucks. Each of the rail car truck bears a dead
sprung
weight, WDS. For each of the trucks WDS lies in the range of 2/3 to 4/3 of WO.
In
an additional feature of that aspect of the invention, for each of the trucks
WDS lies in
the range of 90 % to 110 % of WO. In another additional feature, each of the
trucks
has a resilient suspension having an overall vertical bounce spring rate in
the range of
13,000 to 20,000 lbs per inch.
In still another additional feature, each of the trucks has a resilient
suspension
having an overall vertical bounce spring rate, k, and the value of the square
root of the
dividend obtained by dividing k by a mass equal to W0/g yields a natural
frequency of
less than 2 Hz when the articulated freight car is unloaded. In yet another
additional
feature, at least one of the rail car trucks has a wheelbase to track gauge
width ratio
greater than 1.3.
In another aspect of the invention there is a three piece freight car truck
comprising a rigid truck bolster having a first end and a second end. A first
side
frame is mounted at the first end of the truck bolster. A second side frame is
mounted
at the second end of the bolster. A first spring group is mounted between the
first side
frame and the first end of the bolster. A second spring group is mounted
between the
second side frame and the second end of the truck bolster. Wheel sets each
have a
first and second wheel mounted on a pair of first and second axles. The first
and
second wheels are spaced apart from each other a distance corresponding to a
track
gauge width. The first and second axles are mounted between the first and
second
side frames. The wheel sets have a wheel base length that is (a) greater than
72 inches
and (b) at least 1.3 times as great as the track gauge width.
In an additional feature of that aspect of the invention, the truck has a load
carrying capacity at least as great as an AAR 70 Ton truck, and each of the
spring
groups has a vertical spring rate constant of less than 10,000 lbs./in.
In another aspect of the invention there is a three piece freight car truck
comprising a rigid truck bolster having a first end and a second end. The
truck bolster
has a center plate and a truck center. The truck bolster extends in along a
transverse
axis defined through the truck center. A first side frame is mounted at the
first end of
the truck bolster. A second side frame is mounted at the second end of the
bolster.
The side frames extend in a longitudinal direction relative to the truck
bolster. A first
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spring group is mounted between the first side frame and the first end of the
bolster.
A second spring group is mounted between the second side frame and the second
end
of the truck bolster. Wheel sets each having a first and second wheel is
mounted on a
pair of first and second axles. The first and second axles are mounted between
the
first and second side frames and spaced in a longitudinal direction relative
to each
other. Friction dampers are mounted to provide damping to the spring groups
during
motion of the side frames relative to the truck bolster. Each of the side
frames has a
first pair of friction dampers and a second pair of friction dampers. The
first pair of
friction dampers are mounted longitudinally to one side of a vertical
transverse plane
passing through the truck center of the truck bolster. The second pair of
friction
dampers are mounted to the other side of the vertical transverse plane. The
first pair
of friction dampers includes a first inboard damper and a first outboard
damper. The
first outboard damper is located transversely outboard of the first inboard
damper.
The second pair of friction dampers includes a second inboard damper and a
second
outboard damper. The second outboard damper is located transversely outboard
of
the second inboard damper. Each of the first inboard and first outboard
friction
dampers are independently sprung. Each of the second inboard and second
outboard
dampers is independently sprung.
In an additional feature of that aspect of the invention, each of the first
and
second side frames has a lower frame member, an upper frame member, and fore
and
aft vertical columns, the upper frame member. The lower frame member and the
columns co-operate to define an opening in the side frame through which one
end of
the truck bolster is introduced. The lower frame member has a spring seat. The
spring group has an inboard row of springs and an outboard row of springs
seated in
the spring seat of the lower frame member. Each of the columns has an inboard
friction bearing surface portion and an outboard friction bearing surface
portion.
Brief Description of the Drawings
Figure 1 a shows a side view of a single unit auto rack rail road car;
Figure lb shows a cross-sectional view of the auto-rack rail road car of
Figure 1 a
in a bi-level configuration, one half section of Figure lb being taken
through the main bolster and the other half taken looking at the cross-tie
outboard of the main bolster;
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Figure 1c shows a half sectioned partial end view of the rail road car of
Figure la
illustrating the wheel clearance below the main deck, half of the section
being taken through the main bolster, the other half section being taken
outboard of the truck with the main bolster removed for clarity;
Figure id shows a partially sectioned side view of the rail road car of Figure
1c
illustrating the relationship of the truck, the bolster and the wheel
clearance, below the main deck.
Figure 2a shows a side view of a two unit articulated auto rack rail road car;
Figure 2b shows a side view of an alternate auto rack rail road car to that of
Figure 2a, having a cantilevered articulation;
Figure 3a shows a side view of a three unit auto rack rail road car;
Figure 3b shows a side view of an alternate three unit auto rack rail road car
to
the articulated rail road unit car of Figure 3a, having cantilevered
articulations;
Figure 3c shows an isometric view of an end unit of the three unit auto rack
rail
road car of Figure 3b;
Figure 4a is a partial side sectional view of the draft pocket of the coupler
end of
any of the rail road cars of Figures la, 2a, 2b, 3a, or 3b taken on `4a -
4a' as indicated in Figure 1 a; and
Figure 4b shows a top view of the draft gear at the coupler end of Figure 4a
taken on `4b - 4b' of Figure 4a;
Figure 5a shows a side view of a three piece truck for the auto rack rail road
cars
of Figures la, 2a, 2b, 3a or 3b;
Figure 5b shows a top view of half of the three piece truck of Figure 5a;
Figure 5c shows a partial section of the three piece truck of Figure 5a taken
on
`5c - 5c';
Figure 5d shows a partial isometric view of the truck bolster of the three
piece
truck of Figure 5a showing friction damper seats;
Figure 6a shows a side view of an alternate three piece truck to that of
Figure 5a;
Figure 6b shows a top view of half of the three piece truck of Figure 6a; and
Figure 6c shows a partial section of the three piece truck of Figure 6a taken
on
`6c - 6c'.
DETAILED DESCRIPTION OF THE INVENTION
The description that follows, and the embodiments described therein, are
20915751.1
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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
rail car unit about a horizontal axis perpendicular to the longitudinal
direction.
Yawing is angular motion about a vertical axis. Roll is angular motion about
the
longitudinal axis.
Portions of this description relate to rail car trucks. Several AAR standard
truck sizes are listed at page 711 in the 1997 Car & Locomotive Cyclopedia. As
indicated, for a single unit rail car having two trucks, a "40 Ton" truck
rating
corresponds to a maximum gross car weight on rail 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.
Figures la, 2a, 2b, 3a, and 3b, show different types of auto rack rail road
car,
all sharing a number of similar features. Figure la (side view) shows a single
unit
autorack rail road car, indicated generally as 20. It has a rail car body 22
supported
for rolling motion in the longitudinal direction (i.e., along the rails) upon
a pair of rail
car trucks 23 and 24 mounted at main bolsters at either of the first and
second ends
26, 28 of rail car body 22. Body 22 has a housing structure 30, including a
pair of left
and right hand sidewall structures 32, 34 and a canopy, or roof 36 that co-
operate to
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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. Body 22 can have
staging
in either a bi-level configuration, as shown in Figure lb, in which a second,
or upper
deck 40 is mounted above main deck 38 to permit two layers of vehicles to be
carried; or a tri-level configuration 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, 54 extend to either side of center sill 50, terminating at side sills 56,
58. Main
deck 38 is supported above cross-bearers 52, 54 and between side sills 56, 58.
Sidewall structures 32, 34 each include an array of vertical support members,
in the
nature of posts 60, that extend between side sills 56, 58, and top chords 62,
64. A
corrugated sheet roof 66 extends between top chords 62 and 64 above deck 38
and
such other decks as employed. Radial arm doors 68, 70 enclose the end openings
of
the car, and are movable to a closed position to inhibit access to the
interior of car 20,
and to an open position to give access to the interior. Each of the decks has
bridge
plate fittings (not shown) to permit bridge plates to be positioned between
car 20 and
an adjacent car when doors 68 or 70 are opened to permit circus loading of the
decks.
Two-Unit Articulated Auto Rack Car
Similarly, Figure 2a shows an articulated two unit auto rack rail road car,
indicated generally as 80. It has a first rail car unit body 82, and a second
rail car unit
body 83, both supported for rolling motion in the longitudinal direction
(i.e., along
the rails) upon rail car trucks 84, 86 and 88. Rail car trucks 84 and 88 are
mounted at
main bolsters at respective coupler ends 300 and 330 of the first and second
rail car
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.
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Each of bodies 82, 83 has staging in the nature of a main deck 102 (or 103)
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
1 b, or a
tri-level configuration. Other than brake fittings, and other minor fittings,
car unit
bodies 82 and 83 are substantially the same, differing only in that car body
82 has a
pair of female side-bearing arms adjacent to articulated connector 90, and car
body
83 has a co-operating pair of male side bearing arms adjacent to articulated
connector
90.
Each of car unit bodies 82 and 83 has a through center sill 110 that extends
between the first and second ends 104, 106 (105, 107). A set of cross-bearers
112,
114 extend to either side of center sill 110, terminating at side sills 116,
118. Main
deck 102 (or 103) is supported above cross-bearers 112, 114 and between side
sills
116, 118. Sidewall structures 94, 96 and 95, 97 each include an array of
vertical
support members, in the nature of posts 120, that extend between side sills
116, 118,
and top chords 126, 128. A corrugated sheet roof 130 extends between top
chords
126 and 128 above deck 102 and such other decks as employed.
Radial arm doors 132, 134 enclose the coupler end openings of car bodies 82
and 83 of rail road car 80, and are movable to respective closed positions to
inhibit
access to the interior of rail road car 80, and to respective open positions
to give
access to the interior thereof. Each of the decks has bridge plate fittings
(upper deck
fittings not shown) to permit bridge plates to be positioned between car 80
and an
adjacent auto rack rail road car when doors 132 or 134 are opened to permit
circus
loading of the decks.
For the purposes of this description, the cross-section of Figure lb can be
considered typical also of the general structure of the other rail car unit
bodies
described below, whether 82, 83, 202, 204, 142, 144, 146, 222, 224 or 226. It
should
be noted that Figure lb shows a stepped section in which the right hand
portion
shows the main bolster 75 and the left hand section shows a section looking at
the
cross-tie 77 outboard of the main bolster. The sections of Figures lb and lc
are
typical of the sections of the end units described herein at their coupler end
trucks,
such as trucks
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232, 148, 84, 88, 210, 206. The upward recess in the bolster provides vertical
clearance for the side frames (typically 7" or more).
Three or More Unit Articulated Auto Rack Car
Figure 3a shows a three unit articulated autorack rail road car, generally as
140. It has a first end rail car unit body 142, a second end rail car unit
body 144, and
an intermediate rail car unit body 146 between rail car unit bodies 142 and
144. Rail
car unit bodies 142, 144 and 146 are supported for rolling motion in the
longitudinal
direction (i.e., along the rails) upon rail car trucks 148, 150, 152, and 154.
Rail car
trucks 148 and 150 are "coupler end" trucks mounted at main bolsters at
respective
coupler ends of the first and second rail car bodies 142 and 144. Trucks 152
and 154
are "interior" or "intermediate" trucks mounted beneath respective articulated
connectors 156 and 158 by which bodies 142 and 144 are joined to body 146. For
the
purposes of this description, body 142 is the same as body 82, and body 144 is
the
same as body 83. Rail car body 146 has a male end 159 for mating with the
female
end 160 of body 142, and a female end 162 for mating with the male end 164 of
rail
car body 144.
Body 146 has a housing structure 166 like that of Figure ib, 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 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 minor fittings, car bodies 142 and 144
are
substantially the same, differing only in 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.
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Other articulated auto-rack cars of greater length can be assembled by using a
pair of end units, such as male and female end units 82 and 83, and any number
of
intermediate units, such as intermediate unit 146, as may be suitable. In that
sense,
rail road car 140 is representative of multi-unit articulated rail road cars
generally.
Alternate Configurations
Alternate configurations of multi-unit rail road cars are shown in Figures 2b
and 3b. In Figure 2b, a two unit articulated auto-rack rail road car is
indicated
generally as 200. It has first and second rail car unit bodies 202, 204
supported for
rolling motion in the longitudinal direction by three rail road car trucks,
206, 208 and
210 respectively. Rail car unit bodies 202 and 204 are joined together at an
articulated connector 212. In this instance, while rail car bodies 202 and 204
share
the same basic structural features of rail car body 22, in terms of a through
center sill,
cross-bearers, side sills, walls and canopy, and vehicles decks, rail car body
202 is a
"two-truck" body, and rail car body 204 is a single truck body. That is, rail
car body
202 has main bolsters at both its first, coupler end, and at its second,
articulated
connector end, the main bolsters being mounted over trucks 206 and 208
respectively. By contrast, rail car body 204 has only a single main bolster,
at its
coupler end, mounted over truck 210. Articulated connector 212 is mounted to
the
end of the respective center sills of rail car bodies 202 and 204,
longitudinally
outboard of rail car truck 208. The use of a cantilevered articulation in this
manner, in
which the pivot center of the articulated connector is offset from the nearest
truck
center, is described more fully in U. S. patent no. 7,047,889 for a Rail Road
Car with
Cantilevered Articulation filed Jul. 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 3b shows a three-unit articulated rail road car 220 having first end
unit
222, second end unit 224, and intermediate unit 226, with cantilevered
articulated
connectors 228 and 230. End units 222 and 224 are single truck units of the
same
construction as car body 204. Intermediate unit 226 is a two truck unit having
similar
construction to car body 202, but having articulated connectors at both ends,
rather
than having a coupler end. Figure 3c shows an isometric view of end unit 224
(or
222). Analogous five pack articulated rail road cars having cantilevered
articulations
can also be produced. Many alternate configurations of multi-unit articulated
rail
CA 02354611 2001-08-01
-23-
road cars employing cantilevered articulations can be assembled by re-
arranging, or
adding to, the units illustrated.
In each of the foregoing descriptions, each of rail road cars 20, 80, 140, 200
and 220 has a pair of first and second coupler ends by which the rail road car
can be
releasably coupled to other rail road cars, whether those coupler ends are
part of the
same rail car body, or parts of different rail car bodies of a multi-unit rail
road car
joined by articulated connections, draw-bars, or a combination of articulated
connections and draw-bars.
Figures 4a and 4b show the draft gear at a first coupler end 300 of rail road
car
20, coupler end 300 being representative of either of the coupler ends and
draft gear
arrangement of rail road car 20, and of rail road cars 80, 140, 200 and 220
more
generally. Coupler pocket 302 houses a coupler indicated as 304. It is mounted
to a
coupler yoke 308, joined together by a pin 310. Yoke 308 houses a coupler
follower
312, a draft gear 314 held in place by a shim (or shims, as required) 316, a
wedge 318
and a filler block 320. Fore and aft draft gear stops 322, 324 are welded
inside
coupler pocket 302 to retain draft gear 314, and to transfer the longitudinal
buff and
draft loads through draft gear 314 and on to coupler 304. In the preferred
embodiment, coupler 304 is an AAR Type F70DE 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,
supra.,
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
rail car body
of rail road car 20, (or of the other rail road cars noted above) the
construction
described and illustrated is free of other long travel draft gear, sliding
sills and EOCC
devices, and the fittings associated with them.
. Mini-BuffGear has between 5/8 and 3/4 of an inch in buff at a compressive
force greater than 700,000 Lbs. Other types of draft gear can be used that
will 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-
20915751.1
CA 02354611 2009-08-31
-24-
BuffGear is preferred, other draft gear is available having a travel of less
than 1 3/4
inches at 400,000 Lbs., buff load, 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 6a
and 6b,
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 F70DE 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 E50ARE coupler can
be used. In particular, in alternative embodiments with appropriate housing
changes
where required, AAR Type F79DE and Type F73BE, with or without top or bottom
shelves; AAR Type CS; or AAR Type H couplers can be used to obtain reduced
slack relative to AAR Type E couplers.
In each of the autorack rail car embodiments described above, each of the car
units has a weight, that weight being carried by the rail car trucks with
which the car
is equipped. In each of the embodiments of articulated rail cars described
above there
is a number of rail car units joined at a number of articulated connectors,
and carried
for rolling motion along rail car tracks by a number of rail car trucks. In
each case the
number of articulated car units is one more than the number of articulations,
and one
less than the number of trucks. In the event that 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 3b, the
articulated
connector is mounted at a longitudinal offset distance (the cantilever arm CA)
from
the truck center. In each case, each of the car units has an empty weight, and
also a
design full weight. The full weight is usually limited by the truck capacity,
whether
70 ton, 100 ton, 110 ton (286,000 lbs.) or 125 ton. In some instances, with
low
density lading, the volume of the lading is such that the truck loading
capacity cannot
CA 02354611 2001-08-01
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be reached without exceeding the volumetric capacity of the car body.
The dead sprung weight of a rail car unit is generally taken as the body
weight
of the car, including any ballast, as described below, plus that portion of
the weight of
the truck bearing on the springs, that portion most typically taken as being
the weight
of the truck bolsters. The unsprung weight of the trucks is, primarily, the
weight of
the side frames, the axles and the wheels, plus ancillary items such as the
brakes,
springs, and axle bearings. The unsprung weight of a three piece truck may
generally
be about 8800 lbs. The live load is the weight of the lading. The sum of the
live load
and the dead sprung load and the unsprung weight of the trucks is the gross
railcar
weight on rail, and must not exceed the rated value for the trucks.
In each of the embodiments described above, each of the rail car units has a
weight and a weight distribution of the dead sprung weight of the carbody
which
determines the dead sprung load carried by each truck. In each of the
embodiments
described above, the sum of the sprung weights of all of the car bodies of an
articulated car is designated as WO. (The sprung mass, MO, is the sprung
weight WO
divided by the gravitational constant, g. In each case where a weight is given
herein,
it is understood that conversion to mass can be readily made in this way,
particularly
as when calculating natural frequencies). For a single unit symmetrical rail
road car,
such as car 20, the weight on both trucks is equal. In all of the articulated
auto rack
rail road car 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 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 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, WDS,
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 maximum live
load
carried by each of the trucks be roughly similar such that the overall truck
loading is
about the same, and ideally equal. In any case, for the embodiments described
above,
the design live load for and one truck can be taken as being at least 60 % of
the load
of the next adjacent truck, and advantageously 75 % of the load. In terms of
overall
dead and live loads, in each of the embodiments described the overall sprung
load is
at least 70 % of the nearest adjacent truck, advantageously 80 % or more, and
preferably 90 % of the nearest adjacent truck.
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CA 02354611 2001-08-01
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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 is achieved in
the
embodiments described above by adding ballast to the end car units. That is,
the dead
sprung weight distribution of the end car units is biased toward the coupler
end, and
hence toward the coupler end truck (e.g. 84, 88, 206, 210, 150, 232). For
example, in
the embodiments described above, a first ballast member is provided in the
nature of a
main deck plate 350 of unusual thickness T that forms part of main deck 38 of
the rail
car unit. Plate 350 extends across the width of the end car unit, and from the
longitudinally outboard end of the deck a distance LB. In the embodiment of
Figures
3b and 3c for example, the intermediate of interior truck 234 may be a 70 ton
track
near its sprung load limit of about 101,200 lbs., on the basis of its share of
loads from
rail car units 222 and 226 (or, symmetrically 224 and 226 as the case may be),
while,
without ballast, end trucks 232 would be at a significantly smaller sprung
load, even
when rail car 220 is fully loaded. In this case, thickness T can be 1 1/2
inches, the
width can be 112 inches, and the length LB can be 312 inches, giving a weight
of
roughly 15,220 lbs., centered on the truck center of the end truck 232. This
gives a
dead load of end car unit 222 of roughly 77,000 lbs., a dead sprung load on
end truck
232 of about 54,000 lbs., and a total sprung load on truck 232 can be about
84,000 lbs.
By comparison, center car unit 226 has a dead load of about 60,000 lbs., with
a dead
sprung load on interior truck 234 of about 55,000 lbs., and the 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.
Figures 5a, 5b, 5c and 5d all relate to a three piece truck 400 for use with
the
rail road cars of Figure la, 2a, 2b, 3a or 3b. Figures 1c and ld show the
relationship
of this truck to the deck level of these rail road cars. Truck 400 has three
major
elements, those elements being a truck bolster 402 , symmetrical about the
truck
longitudinal centreline, and a pair of first and second side frames, indicated
as 404.
Only one side frame is shown in Figure 5b given the symmetry of truck 400.
Three
piece truck 400 has a resilient suspension (a primary suspension) provided by
a spring
groups 405 trapped between each of the distal (i.e., transversely outboard)
ends of
truck bolster 402 and side frames 404. The clearance `x' in Figure 1c being 7
inches
in one embodiment between the side frames and the bolster.
20915751.1
CA 02354611 2009-08-31
-27-
Truck bolster 402 is a rigid, fabricated beam having a first end for engaging
one side frame assembly, a second end for engaging the other side frame
assembly
(both ends being indicated as 406) a center plate, or center bowl 408 located
at the
truck center, an upper flange 410 extending between the two ends 406, being
narrow
at a central waist and flaring to a wider transversely outboard termination at
ends
406. Truck bolster 402 also has a lower flange 412 of similar profile to upper
flange
410, and two fabricated webs 414 extending between upper flange 410 and lower
flange 412 to form an irregular closed section box beam. Additional webs 413
are
mounted between the distal portions of upper flange 410 and 414 where bolster
402
engages the one of the spring groups 405. The transversely distal region of
truck
bolster 402 also has friction damper seats 416, 418 for accommodating friction
damper wedges as described further below.
Side frame 404 is a casting having bearing seats 420 into which bearings 421,
and a pair of axles 422 mount. Each of axles 424 has a pair of first and
second wheels
423, 425 mounted to it in a spaced apart position corresponding to the width
of the
track gauge of the track upon which the rail car is to operate. Side frame 404
also has
an upper beam member 424, a lower beam member 426, and vertical side columns
428 and 430, each lying to one side of a vertical transverse plane 425
bisecting truck
400 at the longitudinal station of the truck center. A generally rectangular
opening is
defined by the co-operation of the upper and lower beams members 424, 426 and
vertical columns 428, 430, into which the distal end of truck bolster 402 can
be
introduced. The distal end of truck bolster 402 can then move up an down
relative to
the side frame within this opening. Lower beam member 426 has a spring seat
432
upon which spring group 405 can seat. Similarly, an upper spring seat 434 is
provided by the underside of the distal portion of bolster 402 to engages the
upper
end of spring group 405. As such, vertical movement of truck bolster 402 will
tend to
compress or release the springs in spring group 405.
Spring group 405 has two rows of springs 436, a transversely inboard row and
a transversely outboard row, each row having four large (8 inch +/-) diameter
coil
springs nested with four small diameter coil springs, giving vertical bounce
spring
rate constant, k, for the group of less than 10,000 lbs/inch. This spring rate
constant
can be in the range of 6000 to 10,000 lbs/in., and is advantageously in the
range of
7000 to 9500 lbs/in, giving an overall vertical bounce spring rate for the
truck of
double these values, preferably in the range of 14000 to 18,500 lbs/in for the
truck.
The number of springs, the number of inner and outer coils, and the spring
rate of
the.
CA 02354611 2001-08-01
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various springs can be varied to obtain the desired spring rate constant for
the loading
for which the truck is designed.
Each side frame assembly also has four friction damper wedges arranged in
first and second pairs of transversely inboard and transversely outboard
wedges 440,
442 that engage the sockets, or seats 416, 418. The corner springs in spring
group
405 bear upon a friction damper wedge 440 or 442. Each of vertical columns
428, 430
has a friction wear plate 450 having transversely inboard and transversely
outboard
regions against which the friction faces of wedges 440, 442 can bear,
respectively.
The deadweight compression of the springs will tend to work 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. The springs chosen can have an
undeflected
length of 15 inches, and a dead weight deflection of about 3 inches.
As seen in the top view of Figure 5b, the side by side friction dampers have a
much wider moment arm to resist angular deflection of the side frame relative
to the
truck bolster in the parallelogram mode than would a single such wedge located
on
the spring group centreline. 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
larger
diameter (e.g., 8 in +/-) springs, as compared to the smaller diameter of, for
example,
AAR D5 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 is double, or more, than it would have been for a
single
damper. In the illustration of Figure 5d, the damper seats are shown as being
segregated by a partition 452. If a longitudinal vertical plane 454 is drawn
through
truck 400 through the center of partition 452, it can be seen that the inboard
dampers
lie to one side of plane 454, and the outboard dampers lie to the outboard
side of plane
454. In hunting then, the normal force from the damper working against the
hunting
will tend to act in a couple in which the force on the friction bearing
surface of the
inboard pad will always be fully inboard of plane 454 on one end, and fully
outboard
on the other diagonal friction face. Put differently, the center of force
acting on the
inboard friction face of wedge 440 against column 428 is offset transversely
relative
to the diagonally outboard friction face of wedge 442 against column 430 by a
distance that is at least as great as one full diameter of the large spring
coils in the
20915751.1
CA 02354611 2001-08-01
- 29 -
spring set. This is significantly greater than found in conventional friction
dampers.
Further, in conventional 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 wedge can be less than 15 % of the group total. In the embodiment
of
Figure 5a, it is 50 % of the group total. The wedge angle of wedges 440, 442
is
significantly greater than 35 degrees. The use of more springs permit the
enclosed
angle of the wedge to be significantly larger, in the range of 45 to 60
degrees.
The size of the spring group yields an opening between the vertical columns of
side frame 404 of roughly 33 inches. This is relatively large compared to
existing
spring groups, being more than 25 % greater in width. Truck 400 has a
correspondingly greater wheelbase length, indicated as WB. WB is
advantageously
greater than 73 inches, or, taken as a ratio to the track gauge width, and is
also
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.
In Figures 6a, 6b and 6c, there is an alternate truck embodiment of soft
spring
rate, long wheelbase three piece truck, identified as 460. Although truck 400
is
thought to be preferable, there are a number of alternate possible
configurations of
truck. Truck 460 is generally similar to truck 400, but differs in having a
transom 462
in the form of an upwardly opening channel member bolted between undersides of
the
lower beam members of the left and right side frames 464 respectively. A
transom
such as transom 462 increases the rigidity of the truck against parallelogram
deformation in hunting. Truck 460 also employs constant force inboard and
outboard,
fore and aft pairs of friction dampers 466 mounted in the distal ends of truck
bolster
468. In this arrangement, springs 470 are mounted horizontally in the distal
ends of
truck bolster 468 and urge, or bias, each of the friction dampers 466 against
the
corresponding friction surfaces of the vertical columns of the side frames.
The spring force on friction damper wedges 440 and 442 varies as a function
of the vertical displacement of truck bolster 402, since they are driven by
the vertical
springs of spring group 405. By contrast, the deflection of springs 470 does
not
depend on vertical compression of the main spring group 472, but rather is a
function
20915751.1
CA 02354611 2001-08-01
- 30 -
of an initial pre-load. Although the arrangement of Figure 6a, 6b and 6c still
provides
inboard and outboard dampers and independent springing of the dampers, the
embodiment of Figures 5a is preferred.
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 la, lb, 2a, 2b, 3a and 3b, the ratio of the dead
sprung weight, WD, of the rail car unit (being the weight of the car body plus
the
weight of the truck bolster) without lading to the live load, WL, namely the
maximum
weight of lading, be at least 1:1. It is advantageous that this ratio WD : WL
lie in the
range of 1:1 to 10:3. In one embodiment of rail car of Figures la, 1b, 2a, 2b,
3a and
3b 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 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 are also
less than 3.0
Hz in the unladen vertical bounce mode. That is, the fully laden natural
vertical
bounce frequency for one embodiment of rail cars of Figures la, 1b, 2a, 2b, 3a
and
3b 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
The principles of the present invention are not limited to auto rack rail road
cars, but apply to freight cars, and 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.
20915751.1
CA 02354611 2001-08-01
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Various embodiments of the invention have now been described in detail. Since
changes in and or additions to the above-described best mode may be made
without
departing from the nature, spirit or scope of the invention, the invention is
not to be
limited to those details.
20915751.1
