U.S. patent number 3,718,040 [Application Number 05/178,131] was granted by the patent office on 1973-02-27 for method and apparatus for evaluating railroad track structure and car performance.
This patent grant is currently assigned to Bessemer and Lake Erie Railway Company, Quebec Cartier Mining, United States Steel Corporation. Invention is credited to William H. Freeman, Leavitt A. Peterson, Joseph M. Wandrisco.
United States Patent |
3,718,040 |
Freeman , et al. |
February 27, 1973 |
METHOD AND APPARATUS FOR EVALUATING RAILROAD TRACK STRUCTURE AND
CAR PERFORMANCE
Abstract
Method and apparatus for determining dynamic lateral and
vertical wheel-rail forces. Axle bending sensors and axle load
cells provide signals to a computer programmed to calculate lateral
and vertical wheel-rail forces. These forces are used as the basis
for comparing the effects of a variety of car truck design criteria
and track conditions. Comparison of forces developed by the same
equipment on different runs over the same trackage discloses track
condition changes between runs.
Inventors: |
Freeman; William H. (Port
Cartier, Quebec, CA), Peterson; Leavitt A. (Glenshaw,
PA), Wandrisco; Joseph M. (Lower Burrell, PA) |
Assignee: |
Bessemer and Lake Erie Railway
Company (BY SAID Peterson) N/A)
Quebec Cartier Mining (BY SAID Freeman) N/A)
United States Steel Corporation (BY SAID Wandrises)
N/A)
|
Family
ID: |
22651331 |
Appl.
No.: |
05/178,131 |
Filed: |
September 7, 1971 |
Current U.S.
Class: |
73/146 |
Current CPC
Class: |
G01L
5/20 (20130101); G01L 5/1627 (20200101); G01M
17/10 (20130101) |
Current International
Class: |
G01L
5/20 (20060101); G01L 5/16 (20060101); G01M
17/08 (20060101); G01M 17/10 (20060101); G01m
019/00 () |
Field of
Search: |
;73/146 ;33/144,146 |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Woodiel; Donald O.
Claims
We claim:
1. Apparatus for evaluating railroad track structure and car
performance comprising a railroad car truck including two side
frames, an axle and a pair of wheels mounted on said axle, a pair
of load cells mounted between said axle and each side frame for
providing a first signal and a second signal responsive to the
vertical load of each truck side frame on said axle, a pair of
electrical resistance strain gauges mounted on said axle for
providing a third signal and a fourth signal responsive to the
bending moments in said axle as it rotates under load, and means
connected to said load cells and said strain gauges for combining
said signals and calculating the vertical and lateral wheel-rail
forces at each wheel.
2. Apparatus according to claim 1 in which said pair of strain
gauges are mounted on a common axis on said axle parallel to the
axis of said axle and spaced apart from, and in near proximity to,
each wheel.
3. Apparatus according to claim 1 in which said means for
calculating includes a signal amplifier and conditioner module, an
analog to digital converter connected to said module and a general
purpose digital process computer connected to said converter.
4. Apparatus according to claim 3 which includes a trackside
reference location detector connected to said module.
5. Apparatus according to claim 3 which includes a printer
connected to said computer, a cathode ray tube display connected to
said computer, a digital tape recorder connected to said computer,
and an analog recorder connected to said module and said
converter.
6. A method for evaluating railroad track structure and car
performance comprising the steps of continuously determining axle
bending moments of a truck axle under load at two locations on the
axle between the wheels of said axle as the wheels roll over a
desired section of track, continuously determining the bearing
force each truck side frame exerts on the car axle under load as
the wheels roll over said section of track, and calculating the
vertical and lateral wheel-rail forces for each wheel from said
axle bending moments and said bearing forces.
7. A method according to claim 6 in which the bending moments are
determined by strain gauges and the bearing forces are determined
by load cells and which includes statically calibrating the strain
gauges and load cells, measuring the distance between the axis of
the axle and the surface of the tread of the wheel, measuring the
distance between each load cell and the adjacent wheel, measuring
the distance between each strain gauge and the adjacent wheel,
measuring the distance between the wheel treads, and said
calculating of vertical and lateral forces are performed by
combining an output from each strain gauge, an output from each
load cell and said measurements.
8. A method according to claim 7 in which the step of calculating
the vertical and lateral forces comprises the steps of
simultaneously solving the four free-body equations
M.sub.L = L.sub.L c + B.sub.L (a + d) - V.sub.L d
M.sub.R = L.sub.L c + B.sub.L (a + (f - e)) - V.sub.L (f - e)
M.sub.R = L.sub.R c + B.sub.R (b + e) - V.sub.R e
M.sub.L = L.sub.R c + B.sub.R (b + (f - d)) - V.sub.R (f - d)
for L.sub.L, L.sub.R, V.sub.L, and V.sub.R where L.sub.L is the
lateral force on the left wheel, L.sub.R is the lateral force on
the right wheel, V.sub.L is the vertical force on the left wheel,
and V.sub.R is the vertical force on the right wheel and in which
equations M.sub.L is the calibrated strain gauge output adjacent
the left wheel, M.sub.R is the calibrated strain gauge output
adjacent the right wheel, B.sub.L is the calibrated load cell
output adjacent the left wheel, B.sub.R is the calibrated load cell
output adjacent the right wheel, a is the measured distance between
the left wheel and the adjacent load cell, b is the measured
distance between the right wheel and the adjacent load cell, c is
the measured distance between the axis of the axle and the surface
of the tread of the wheel, d is the measured distance between the
left wheel and the adjacent strain gauge, e is the measured
distance between the right wheel and the adjacent strain gauge, and
f is the measured distance between the wheel treads.
9. A method according to claim 8 which includes the steps of
converting the outputs of the strain gauges and load cells from
analog to digital form and calculating force relationships and in
which said calculations are made by a general purpose digital
computer.
10. A method according to claim 9 which includes recording the
outputs of the strain gauges and the load cells in digital
form.
11. A method according to claim 10 which includes the steps of
simultaneously recording with said outputs periodic trackside
reference location indicator signals.
12. A method according to claim 9 which includes the steps of
determining the vertical and lateral forces for all axles of a
truck.
13. A method according to claim 9 which includes the steps of
determining the lateral and vertical forces for all axles of a
car.
14. A method according to claim 8 which includes calculating the
net lateral force on the two wheels of a single axle according to
the equation
L.sub.N = L.sub.R - L.sub.L
where L.sub.R is the lateral force on the right wheel, L.sub.L is
the lateral force on the left wheel, and L.sub.N is the net lateral
force which will have a positive sign if toward the right wheel and
a negative sign if toward the left wheel.
15. A method according to claim 8 including calculating the total
dynamic axle load B', according to the equation
B' = B.sub.R ' + B.sub.L '
where B.sub.R ' is the deviation of the right load cell output from
static value expressed as a force and B.sub.L ' is the deviation of
the left load cell output from static value expressed as a
force.
16. A method according to claim 8 including the steps of
determining dynamic vertical forces V.sub.L ' and V.sub.R ' where
V.sub.L ' is the deviation from static value of V.sub.L and V.sub.R
' is the deviation from static value of V.sub.R.
17. A method according to claim 8 including calculating dynamic
axle rock according to the equation
R.sub.R ' = B.sub.R ' - B.sub.L '
where B.sub.R ' is the deviation of the right load cell output from
static value expressed as a force, B.sub.L ' is the deviation of
the left load cell output from static value expressed as a force,
and R.sub.R ' is the dynamic axle rock base on the right wheel
expressed as a force.
18. A method according to claim 8 including calculating the lateral
to vertical force ratio, R, according to the equation
R = L/V
where L and V are the calculated lateral and vertical forces.
19. A method according to claim 13 including calculating the
transfer of vertical forces to and from leading and trailing
axles.
20. A method according to claim 8 including determining the
frequencies of occurrence of repetitive deviations from static
values of lateral and vertical forces.
21. A method according to claim 9 in which said calculations are
performed for increments of travel over said sections of track
thereby providing a series of values for vertical forces, a series
of values for lateral forces and a series of values for each force
relationship and which includes placing each value of a series in a
range group, determining the frequency of occurrence of values in
said range group, determining the mean of values of said group and
determining one standard statistical deviation from the mean of
values of said group.
22. A method according to claim 9 in which said calculations are
performed for increments of travel over said section of track
thereby providing a series of lateral and vertical force values,
and which includes measuring the transfer of vertical forces from
one rail to the other rail on curved track.
23. A method according to claim 9 in which said calculations are
performed for increments of travel over said section of track
thereby providing a series of lateral and vertical force values and
which includes grouping said values according to track
curvature.
24. A method according to claim 9 in which the vertical forces,
lateral forces, and force relationships are determined for a first
car truck design criterion and a second car truck design criterion
and which includes the steps of comparing said forces and force
relationships for the first criterion and the second criterion.
25. A method according to claim 24 in which the design criterion is
the size of the suspension springs and the damping of the
suspension system.
26. A method according to claim 24 in which the design criterion is
the effectiveness of devices controlling movement between truck
side frames and axles.
27. A method according to claim 24 in which the design criterion is
the lateral clearance in the car axle assembly.
28. A method according to claim 24 in which the design criterion is
the lateral clearance between truck bolster and side frames.
29. A method according to claim 24 in which the design criterion is
the wheel base of the car truck.
30. A method according to claim 24 in which the design criterion is
the wheel tread contour.
31. A method according to claim 24 in which the design criterion is
the derailment tendency of the axle.
32. A method according to claim 24 in which the design criterion is
the spacing between wheels on an axle.
33. A method according to claim 9 in which the vertical forces,
lateral forces, and force relationships are determined for a car
passing over a first section of track having a first condition and
for said car passing over a second section of track having a second
condition and which includes the steps of comparing said forces and
force relationship for the first track condition and the second
track condition.
34. A method according to claim 33 in which the track conditions
are different track gauges on tangent track.
35. A method according to claim 33 in which the track conditions
are different track gauges on curved track.
36. A method according to claim 33 in which the track conditions
are the amount of super-elevation on curved track.
37. A method according to claim 33 in which the track conditions
are track having welded joints and tracks having spliced
joints.
38. A method according to claim 33 in which the track conditions
are the amount of lateral profile.
39. A method according to claim 33 in which the track conditions
are the amount of vertical profile.
40. A method according to claim 33 in which the track conditions
are ballast resiliency of a first consistency and ballast
resiliency of a second consistency.
41. A method according to claim 33 in which the track conditions
are the amount of curvature in the track.
42. A method according to claim 9 which includes the steps of
recording the values of the vertical forces, lateral forces, and
force relationships for said section of track, determining a second
set of vertical and lateral force values with the same rolling
equipment, and comparing the recorded values with the second set of
values.
43. A method according to claim 9 which includes the steps of
establishing acceptable standards for values of said forces and
force relationships for a test car, operating said test car on a
second section of track, calculating said forces and force
relationships for said test car for said second section of track
and comparing said calculated values of said forces and force
relationships for said second section of track with said standard
values.
Description
This invention relates to a method and apparatus for evaluating
railroad track structure and car performance and more particularly,
to a method and apparatus for determining the dynamic forces
between wheels and rails, analyzing the results and applying the
results of the analysis to evaluation of truck design and track
conditions.
There is a continuous and steady effort within the railroad
industry to carry heavier loads, to achieve higher speeds and to
reduce derailments while at the same time to decrease car and track
maintenance. The ability to set track standards to meet these
requirements, as for example, for a given unit train hauling
operation or for a defined high speed passenger run in order to
prevent severe and accelerated rail deterioration or a relatively
high ratio of derailments, is highly depended upon knowledge of the
dynamic response of components of cars to track input. The dynamic
response is significantly affected by car conditions, such as the
dimensional relationships and characteristics of the bearings,
springing and snubbing in the suspension system of the cars and
track conditions, such as alignment, vertical profile, lateral
profile, gauge, cross-level and ballast characteristics.
There have been studies made of wheel-rail interaction. Two
articles, one entitled "Lateral Forces Between Wheels and Rails" by
P. E. Olson and S. Johnsson, A.S.M.E. Publication 60-RR-6 dated
Apr. 20, 1960 and the other entitled "Lateral Loading Between
Locomotive Truck Wheels and Rail Due To Curve Negotiation" by L. F.
Koci and H. A. Marta, A.S.M.E. Publication 65-WA/RR-4 dated Nov.
11, 1965, describe limited studies on lateral forces. Both studies
use the technique of placing strain gauges on the plate of a wheel
at a radial distance supposedly insensitive to change in vertical
load to measure lateral loads. Vertical loading was determined from
static measurements. Interpretation of the data was inconclusive
and apparently had limited usefulness. An article entitled
"Research on the Operation Stresses in PATH Rail Car Axles, Drive
Systems, Wheels and Rail Benders" by M. Youtar, A.S.M.E.
Publication 66-RR-6 dated May 4, 1966, describes a study directed
to identify causes of premature cracking of axles with inboard
bearings and based on axle strain measurements. Since the study was
oriented towards axle cracking, the disclosure on the dynamics of
wheel-rail relationships was incidental and not complete.
According to our invention, railroad car trucks were equipped with
axle bending sensors and axle load cells for different truck
configurations. Test equipment including signal conditioning
equipment, a tape recorder, an analog to digital converter, a
computer and readout equipment was located on an equipment car and
connected to the car truck undergoing test. Test runs were made
over a section of track having a wide variety of curve radii for
each type of truck configuration under study. The computer was
programmed to develop dynamic vertical and lateral forces
throughout the run and then to provide statistical comparison of
the data of car behavior for different track inputs.
The data is then presented in form to evaluate such items as the
desirability of making a vehicle suspension system more resilient,
the desirability of reducing the unsprung weight of a car, altering
car design characteristics which might influence natural periods of
oscillation, reduction of static and dynamic axle loading,
adjustment of track condition parameters to match operations and
decreasing the frequency and level of lateral wheel-rail forces in
curves. The data is then used for altering track conditions and for
car design criteria to minimize lateral wheel-rail forces and to
achieve the objective of moving cars along the track centered on
the rail with uniform loading on all wheels on curves and tangent
track.
It is therefore an object of our invention to provide apparatus for
continuously measuring dynamic vertical and lateral forces between
railroad wheels and rails.
Another object is to provide apparatus for analyzing the dynamic
and lateral forces between railroad wheels and rails.
A further object is to provide a method for determining the dynamic
vertical and lateral forces between railroad wheels and rails.
A still further object is to provide a method of evaluating design
criteria for railroad car trucks.
Still another object is to provide a method for evaluating railroad
track conditions.
Yet another object is to provide a method for evaluating track
maintenance and detecting track defects.
These and other objects will become more apparent after referring
to the following drawings and specification in which:
FIG. 1 is a schematic drawing of the testing apparatus of our
invention.
FIG. 2 is a force diagram illustrating the force calculations of
our method.
FIG. 3 is a partial side view of a truck showing a pad which will
absorb part of the lateral energy transmitted to the wheel.
FIG. 4 is a partial computer produced graph for a type C truck,
first axle, trailing truck, tangent track, welded rail at 35 miles
per hour.
FIG. 5 is a partial computer produced graph for a type A truck,
first axle, lead truck, 7.degree. curve, jointed rail at 35 miles
per hour.
FIG. 6 is a partial computer produced graph showing various forces
for a type A truck, first axle, trailing truck, tangent track,
welded rail at 35 miles per hour.
FIG. 7 is a partial computer produced graph showing various forces
for a type B truck, first axle, trailing truck, tangent track,
welded rail at 35 miles per hour.
FIG. 8 is a series of charts showing the cumulative frequency
distribution of lateral wheel forces in thousands of pounds for
three types of trucks as percent of time less than ordinate
value.
FIG. 9 is a series of charts showing the cumulative frequency
distribution of the lateral to vertical force ratio for three types
of trucks on a 5.degree. curve, welded rail, at 35 miles per hour
as percent of time less than ordinate value.
FIG. 10 is a series of charts showing the cumulative frequency
distribution of the lateral wheel force (A), dynamic vertical wheel
force (B) and lateral to vertical force ratio (C) on the high rail
side for a type A truck on a 5.degree. curve at 35 miles per hour
for welded rail and jointed rail as percent of time less than
ordinate value.
FIG. 11 is a series of charts showing the cumulative frequency
distribution of the net lateral force (A), dynamic axle rock (B)
and dynamic axle loading (C) in thousands of pounds of type A, B,
and C trucks on a 5.degree. curve, welded rail at 35 miles per hour
as percent of time less than ordinate value.
FIG. 12 is a chart showing the net lateral force, at one standard
deviation, 1.sigma., above the mean of the frequency distribution
of the measured values for three trucks at different curvatures on
jointed rails at 35 miles per hour.
FIG. 13 is a chart showing the total axle load, at one standard
deviation, 1.sigma., above the mean of the frequency distribution
of the measured values for three type trucks of different
curvatures on welded rails at 35 miles per hour.
FIG. 14 is a chart showing the load transfer across an axle at one
standard deviation, 1.sigma., above the mean of the frequency
distribution of the measured values on various degrees of track
curvatures for different type trucks at different speeds for both
jointed and welded rail.
FIGS. 15A and B are grossly exaggerated depictions of the forces on
rails and wheels due to alternate axle loading and unloading.
FIG. 16 is a chart showing the maximum wheel lateral to vertical
ratio at one standard deviation, 1.sigma., above mean of the
frequency distribution of the calculated values on different track
curvatures at 35 miles per hour on jointed rails for two types of
trucks.
FIG. 17 is a chart showing the maximum wheel lateral to vertical
ratio at one standard deviation, 1.sigma., above mean of the
frequency distribution of the calculated values on different track
curvatures at 35 miles per hour on welded rails for three types of
trucks.
FIG. 18 is a chart showing the average vertical truck load shifting
in thousands of pounds to high and low rails on different curves
for different trucks at 20 and 35 miles per hour.
FIG. 19 is a partial computer produced graph showing various forces
in thousands of pounds for a type A truck, second axle, lead truck,
tangent track, jointed rail at 20 miles per hour correlated with
rail joints.
FIG. 20 is a partial computer produced graph showing various forces
in thousands of pounds for a type A truck, first axle, lead truck,
tangent track, jointed rail at 35 miles per hour correlated with a
turn out.
Referring now to FIG. 1, reference numeral 2 generally indicates
the truck of a railroad car with side frames 4, wheels 6, and axles
8 to be tested. Two conventional electrical resistance strain
gauges 10 are mounted on each axle on a common axis parallel to the
axis of rotation of the axle and spaced apart from the wheels as
shown by distances d and e in FIG. 2. Leads 12 from strain gauges
10 are connected to slip rings 14 through holes in the axle. The
slip rings 14 are connected through leads M to a car mounted signal
amplifier and conditioner module 15. Four conventional compression
load cells using electrical resistance strain gauges 18 (only two
are shown in FIG. 1) are located between side frames 4 and axles 8
and are connected through lines B to module 15. Module 15 is
connected to an analog signal recorder 16 and to a conventional
analog to digital converter 20 which is in turn connected to a
general purpose digital computer 22. Computer 22 may include a
conventional cathode ray tube display 24 and a conventional
teleprinter 26. A digital tape recorder 27 is connected to computer
22. A track side reference location detector 28, such as a mile
post detector, is connected to module 15. Detector 28 may be either
the metal detector type in which detecting the presence of a metal
plate attached to a tie at the desired location provides a signal
to the recorder 16, a mechanical trip type which has a track side
arm at each desired location for tripping a car mounted switch to
provide a signal to the recorder or light source and photocell
receiver. Recorder 16 is connected to converter 20.
Determination of the dynamic vertical and lateral force reactions
at each wheel is a particularly critical part of the measurements.
While the load cells 18 provide a measure of the vertical force
transmitted from car to axle, it is not a measure of the forces
transmitted from wheel to rail. In order to determine the vertical
and lateral wheel-rail forces, calibrated strain gauges 10 are used
at known locations and positions on the axle to measure
instantaneous bending moments.
Continuous measuring of bending strains at any known point in the
revolving axle 8 is made by strain gauges 10. The strain gauge will
provide a sine wave output as the location undergoes bending with
axle rotation. This output is due to gauge position and is not a
direct measurement of the maximum bending strain except at the
extreme points twice each revolution. More than one strain gauge
may be used at each location to improve the accuracy of the
determination of bending moments. Once the axle strain gauges are
calibrated by well known static methods, the vertical and lateral
forces are calculated using the following free-body based equations
based on the force diagram of FIG. 2. Under conditions of
equilibrium using the forces and moment arms shown in FIG. 2
M.sub.L = L.sub.L c + B.sub.L (a + d) - V.sub.L d (1)
where a, c, and d are distances shown in FIG. 2, M.sub.L is the
bending moment of the axle at the left point, M.sub.L where a
strain gauge 10 is located, L.sub.L is the left lateral force,
B.sub.L is the left bearing force as measured by a load cell 18,
and V.sub.L is the left vertical force. In addition,
M.sub.R = L.sub.L c + B.sub.L (a + (f - e)) - V.sub.L (f - e)
(2)
where a, c, e, and f are distances shown in FIG. 2, M.sub.R is the
bending moment of the axle at the right point, M.sub.R where a
strain gauge 10 is located and B.sub.L and V.sub.L are as
previously described. And,
M.sub.R = L.sub.R c + B.sub.R (b + e) - V.sub.R e (3)
where b, c, and e are distances shown in FIG. 2, L.sub.R is the
right lateral force, B.sub.R is the right bearing force as measured
by a load cell 18, V.sub.R is the right vertical force and M.sub.R
is as previously described. And,
M.sub.L - L.sub.R c + B.sub.R (b + (f - d)) - V.sub.R (f - d)
(4)
where b, c, d, and f are distances shown in FIG. 2 and M.sub.L,
L.sub.r, B.sub.R, and V.sub.R are as previously described.
The equations are solved simultaneously Rpair to provide V.sub.L,
V.sub.R, L.sub.L, and L.sub.R as follows:
V.sub.L = (M.sub.L - M.sub.R)/(f-d-e) + B.sub.L (5) V.sub.R =
(M.sub.R - M.sub.L)/(f-d -e) + B.sub.R (6)
L.sub.L = (M.sub.L - M.sub.R)/(f-d-e) (b - e/c) + (M.sub.R /c) -
B.sub.L (a/c) (7) L.sub.R = (M.sub.R - M.sub.L)/(f-d -e) (b - d/c)
+ (M.sub.L/ c) - B.sub.R (8) c)
In addition by way of definition,
L.sub.N = L.sub.R - L.sub.L (9)
where L.sub.R is the right lateral force, L.sub.L is the left
lateral force and L.sub.N is the net lateral force which will have
a sign indicating its direction.
The desired vertical and lateral forces are obtained by
substituting the known dimensions and measured quantities for the
appropriate terms. A resultant computed value having a negative
sign indicates that the force vector is in the opposite direction
to that shown in FIG. 2.
Although the physical dimensions a, b, c, d, e, and f represent
known measured constants, it is recognized that force application
loci, such as the vertical rail-wheel locations, actually do shift
on the wheel tread as the car rolls along the rails and that the
force vectors in general are not always oriented on or acting in a
simple manner as illustrated in the diagram. However, the magnitude
of error introduced by treating the moment arms as constant appears
to be well within acceptable limits.
For the determination of truck design and performance
characteristics, a number of trucks are equipped with strain gauges
and load cells, at least one truck for each type of truck for which
the particular design and performance characteristics are to be
compared. The instrumented trucks are used on loaded or empty
freight cars and the connections B and M of FIG. 1 are made by
cable to a test car in which the remainder of the components of
FIG. 1 are installed. The test car may be conveniently located in a
train consist of loaded and/or empty freight cars and the test
car.
The test train is then run along a section of track at various
speeds. The section of track preferably includes an adequate
variety of curvatures and other track conditions to insure the
accumulation of sufficient representative data. For example, a
15-mile section of track which included 6 miles of tangent track,
curves from 1.degree. to 8.degree. and grades up to 0.8 percent
gave very satisfactory results. Speeds should be relatively
constant, such as a run at 20 miles per hour and a run at 35 miles
per hour, with other runs over specific track sections at 25, 30,
40, and 45 miles per hour. Instrumented trucks should be placed in
both leading and trailing positions.
The strain gauges and load cells are calibrated statically before a
run. During the run the data is continuously fed into the module
15. The track side reference location detector 28 simultaneously
records location signals for later synchronization. Module 15
includes amplifiers for altering all the signals recorded to be
compatible with the inputs to the analog to digital converter. The
signals are then processed by the computer in specific ways for
specific purposes.
The data shown hereafter includes measurements from three different
types of railroad trucks designated as a type A car equipped with
standard plain bearings, a type B car equipped with standard roller
bearings and a type C car equipped with standard roller bearings
and lateral energy absorbing pads. FIG. 3 shows a lateral energy
absorbing pad 30 placed between axle 8 and side frame 4. Four
lateral pads are provided on each truck.
FIGS. 4 and 5 are computer produced graphs illustrating one method
of displaying the data acquired on test runs and processed by the
computer. FIG. 4 is for a type C car, first axle, trailing truck on
tangent track with welded rails at 35 miles per hour. FIG. 5 is for
a type A car, first axle, lead truck on a 7.degree. curve with
jointed rails at 35 miles per hour speed. The abscissae are shown
marked in units, each 357 units represents one second as shown in
FIG. 4A. The same scale is used in the remainder of FIG. 4 and
FIGS. 5, 6, 7, 19 and 20. Only a small fragment of the total run is
shown.
FIG. 4A shows the output of one of the strain gauges 10 expressed
as a bending moment in foot pounds.
FIG. 4B shows the total axle load in pounds as determined from
gauges 18 in FIG. 1. The total axle load may be expressed as
B' = B.sub.R ' + B.sub.L' (10)
where B.sub.R ' and B.sub.L ' are the algebraic deviations from
static values. B' is a measure of car bounce.
FIG. 4C shows the calculated vertical forces V.sub.R ' and V.sub.L
' as deviations from static values.
FIG. 4D shows a measure of car rock in pounds and may be expressed
as
R.sub.R ' = B.sub.R ' - B.sub.L ' (11) where R.sub.R ' is a measure
of rock based on the right wheel and B.sub.R ' and B.sub.L ' are
deviations from static values.
FIG. 4E shows the dynamic lateral forces L.sub.L and L.sub.R where
a positive value indicates an inward force applied to the wheel and
a negative value indicates an outward force applied to the
wheel.
FIG. 4F shows the dynamic net lateral force expressed as
L.sub.N = L.sub.R - L.sub.L (9)
where L.sub.N is the dynamic net lateral force based on the right
wheel and L.sub.R and L.sub.L are the dynamic lateral forces.
FIG. 4G shows the ratio of the lateral to vertical forces for each
wheel expressed as
Ratio.sub.R = L.sub.R /V.sub. R (12)
for the right wheel, and
Ratio.sub.L = L.sub.L /V.sub. L (13)
for the left wheel. The ratios are a measure of potential derailing
tendencies.
FIG. 5 shows the same data as FIG. 4, but for a type A car, first
axle, lead truck on a 7.degree. curve, jointed rail at 35 miles per
hour.
By comparing many of the instantaneous lateral and vertical forces
and through appropriate combinations and calculations instantaneous
values of other forces and force relationships, such as ratios,
rack, bounce, and other factors, it is possible to determine many
design criteria for car trucks and to evaluate track
conditions.
However, we have found that merely comparing average values of
force did not disclose the true details of the relationships
between car and track. Accordingly, the data was rearranged to
distribute instantaneous force measurements into various range
groupings and to determine the percent of time the forces were at
each of these designated levels. This analysis was programmed into
the computer.
Such range groupings of forces are the true measure of car behavior
on a given section of track. Table 1 shows the data of FIG. 4
separated into frequency distribution groups; averaged, and the
value of first standard deviation from the mean determined by
conventional statistical methods. When these techniques were
applied to the accumulated data, we found that each car, and each
minor modification to car characteristics, has a unique and
identifiable response to a given set of track input conditions
which would repeat under the same set of circumstances.
##SPC1##
VERTICAL FORCE COMPARISONS
Transmittal of the load vertically to the axle bearings and
eventually to the wheel-rail contact area is a much more complex
occurrence when a car is moving as compared to when a car is
standing still. The dynamic pattern of load transmittal as a car
traverses a given section of track is determined chiefly by the
suspension system of the truck, such as bolsters, springs,
bearings, wedges or snubbers, and the magnitude may vary
considerably depending upon the characteristics and clearance
between these components.
Dynamic load transfer from one side of an axle to the other is
termed "rock," equation 11, and has been long recognized as a
potential cause for derailment. Rock is associated with rail joint
spacing and truck type and the relation with truck type is shown in
FIG. 11B. FIGS. 4D and 5D illustrate rock by the computer produced
graph method. FIG. 14 also shows the nature of rock for different
type cars at different speeds for both jointed and welded rail, the
type C car exhibits the highest capability for absorbing
shocks.
Bounce, equation 10, is the simultaneous and synchronous loading of
all bearings of the trucks of a car. Car bounce is shown in FIGS.
4B and 5B, and FIG. 13 shows the bounce comparison of the three
type trucks, type C exhibits the smallest bounce tendency. Table 2
shows a summary of vertical loadings and high frequency peak to
peak shocks.
TABLE 2
Vertical Axle Loading (Thousands of Lbs.)
Range of Total Axle Load During Long Characteristic Dynamic Term
Bounce Cycles Peak to Peak "Shocks" Truck Tangent 5.degree. Curve
Tangent 5.degree. Curve Type A 44 - 84 56 - 72 0.5 5 B 58 - 70 59 -
69 1.5 5 C 58 - 70 60 - 68 1.0 3
the type A truck was equipped with a free spring travel system and
the type B truck was equipped with friction snubbing in the spring
group. The total axle load as shown in FIG. 6 shows a relatively
smooth sinusoidal curve of large magnitude whereas the snubbed car
in FIG. 7 shows the effects of snubbing by reducing the magnitude
but creating high frequency vibrations, about 60-80 Hz.
In-depth examination of the dynamic vertical forces in relationship
to the simultaneous lateral forces showed that the outboard bearing
on axles transforms pure vertical force input at the bearing into
well defined lateral as well as vertical force elements at the
wheel-rail contact area. Vertical loading and unloading of the
axle, such as rock or bounce, will produce a reciprocating lateral
reaction at the rail tending to force the rail inward and outward
respectively. This effect is grossly exaggerated diagrammatically
in FIG. 15A for loading and FIG. 15B for unloading. An axle will
not, of course, bend as shown in FIG. 15B, but the axle has a
tendency towards being, thus creating the wheel-rail forces as
shown.
LATERAL FORCE COMPARISONS
The type B truck, equipped with roller bearings, allows only a
small amount of lateral movement between the side frames and axles.
Such truck is commonly called a "rigid" truck as compared to the
"flexible" plain bearing truck, type A, which has considerable more
lateral movement tolerance, or the type C which has the lateral
pads to facilitate lateral movement. The differences are shown in
FIGS. 8, 11A, and 12. These comparisons may also be made for any
particular change in dimension to optimize the lateral force
effects.
The dynamic lateral force frequencies correlates with the distances
(7 to 12 inches) between rail corrugations at the predominant train
speed.
This type of analysis may also be used to determine the
effectiveness of lateral energy absorbing pads by comparing
different dimensions, different materials, and different
compressibilities.
LATERAL TO VERTICAL RATIOS
A potential derailing tendency for each type of truck was indicated
by the ratio of dynamic lateral to vertical forces for each wheel.
This is shown in computer produced graph form in FIG. 4G and FIG.
5G. FIGS. 9 and 10C compare the ratio on two different types of
track and on high and low side of curves. FIG. 16 compares the
ratio for types A and B trucks over jointed rail at 35 miles per
hour over several curves, and FIG. 17 is a similar comparison, but
on welded rails. Obviously, the lateral to vertical ratio could be
calculated using net lateral force on each axle, or total net truck
lateral force as well as different vertical forces could be used to
develop potential derailment tendencies. FIGS. 16 and 17 show that
under certain conditions the type A truck has a much higher
probability of derailment on curves than the type B truck because
it develops higher ratios of instantaneous lateral to vertical
forces at each wheel, which means that for a given lateral force
attempting to push one of its wheels off the rail, it has a lower
vertical force tending to keep the wheel seated on the rail, for a
greater portion of time. Car bounce, as determined by equation 10,
significantly contributes to potential derailment tendencies and
effects the lateral to vertical ratio as shown in FIG. 5.
TRUCK DESIGN CRITERIA
The car equipped with type B truck as compared to the cars equipped
with type A truck or the type C truck exhibited a definite pitching
action on curves, i.e., the weight was being transferred back and
forth between the rear and the front truck. This tendency is
determined by comparing the instantaneous vertical forces on all
four axles. In addition, the type B truck had a tendency to rotate,
as through it were a rigid rectangle, about the lead wheel on the
high rail side of the curve. On the other hand, the type A and C
trucks allow lateral movement of the axles relative to the side
frame and truck bolsters have the capability of absorbing the minor
lateral profile changes of the rail by a back and forth hunting
movement of the wheel and axle units only, without the heavier mass
of the car and lading. This independent lateral freedom allows such
trucks to accommodate to track inputs, such as lateral
irregularities, curve alignment and gauge variations, without the
higher mass reinforced pivoting and pitching action of more rigid
designs. The relative effect of pivoting action may be determined
by comparing the charts of FIG. 8 and is further shown in FIG.
11A.
Other car truck design criteria that may be determined in a similar
manner include the effectiveness of suspension systems and damping
devices, optimal wheelbase, spacing between wheels on an axle,
allowable lateral motion of truck components, clearances between
truck components and wheel tread contours. This technique can also
be used to determine that best compromise in two or more car truck
design features. For example, referring to FIGS. 6 and 7, the type
A car has a higher dynamic axle load than the type C car and
therefore more truck wear and damage potential, but the type C car,
because of the snubbing effect produces lower axle forces at high
frequency levels which can result in a corrugated rail wear
pattern.
TRACK CONDITIONS
The rails, roadbed, and general track structure conditions act as
stimuli for inducing various reactions in passing cars. Track
geometry, such as alignment, vertical profile, lateral profile,
gauge, cross-level rail joint conditions and ballast
characteristics very noticeably affect car behavior. Statically
measured track dimensions do not necessarily yield on accurate
index as to what a car will sense under moving or dynamic
conditions because physical measurements of gauge, cross-level,
alignment and the like under no-load conditions do not necessarily
reflect the acceptability of the track under operation conditions.
This is particularly true of roadbed resiliency. The methods
already described for evaluating truck features are also used in
evaluating track conditions by comparing truck reactions with
different track inputs.
As an example, FIG. 13 clearly identifies the effect of track
curvature on total axle load for the three type trucks. FIGS. 14,
16, 17, and 18 also illustrate the effect of curves. Table 3 shows
a typical lateral force for different degrees of curvature for a
type b truck, first axle, high rail, welded rail at 35 miles per
hour.
TABLE 3
Lateral Force, (Lbs.)
% of time average maximum incurred Tangent Track 4000 10000 1.4
3.degree. Curve 6000 16000 0.2 5.degree. Curve 11500 17000 19.4
This table indicated that curves over 3.degree. generally result in
much higher forces, further indicated in FIG. 12. This information
is useful in car truck design because, if there are little or no
curves greater than 3.degree., the design could include smaller
axles and other components need not be designed to handle large
stresses.
While the railroad industry has known that welded rail has
advantages over jointed rail, there is no known complete evaluation
of the advantages of welded rail. The method of our invention
provides such an evaluation. For example, Table 4 compares maximum
values for various measurements for jointed and welded rails.
TABLE 4
Comparison of Maximum Values For Various Measurements Jointed Vs.
Welded Rail Type A Truck on 5.degree. Curves at 35 MPH
Rail % Measurement jointed welded reduc- tion 1. Vertical Dynamic
Wheel- Rail Force (1000 Lbs.) 14 8 43 2. Lateral Wheel-Rail Force
(1000 lbs.) 21 16 24 3. Net Lateral Force Per Axle (1000 lbs.) 16 8
50 4. Rock Per Axle (1000 lbs.) 15 8 47 5. L/V Force Ratio per
Wheel 1.10 0.85 23 6. Axle Bending Moment (1000 lb.-ft.) 46 38
17
In addition, FIG. 10 illustrates comparisons concerning lateral
forces, vertical forces and lateral to vertical ratios for a type A
truck.
While car rock has already been discussed, our method showed that
repetitive car rocking occurred on jointed rail but only at widely
separated locations on welded rail. Maximum rocking amplitudes on
jointed track consistently correspond to joint half intervals of
191/2 feet for staggered joints of 39 foot rail lengths during
speed ranges of 18 to 25 miles per hour. Above 25 miles per hour,
rail joint input produced very short duration shock impacts instead
of the longer term resonant reactions at lower speeds. This is
illustrated in FIGS. 19 and 20.
An analysis of the behavior of the three type trucks when
travelling at 35 miles per hour over three separate but statically
similar 5.degree. curves showed that the true equilibrium speed is
not solely a function of the degree of curvature and the amount of
super-elevation of the curve. On the first curve, the type B and
type C trucks had lateral and vertical forces indicating speeds
above equilibrium. On the second curve, the type C truck appeared
to be at equilibrium speed and the type A and type B trucks
appeared to be below equilibrium speed. All of the trucks appeared
to be below equilibrium speed on the third curve. This points out
that there is a difference between static, no-load measured
super-elevation and the elevation sensed by different types of
moving cars. FIG. 18 shows one method of comparing equilibrium
behavior based on average vertical truck load shifts between high
and low rails on selected curves at constant speeds of 20 and 35
miles per hour.
There are a number of other track conditions that affect car
behavior that may be evaluated by our method. For example, as shown
in FIG. 19, the location of the rail joints is shown when car rock
is plotted on a computer production graph. FIG. 20 shows the
location of a turnout on a computer produced graph display. It is
also possible to distinguish between trailing and facing movements
in turnouts. Other track conditions that may be evaluated by this
method include gauge-curve relationship and gauge control.
Track condition in relation to a standard may be established by
using a test car equipped with a truck of known dynamic
characteristics. When the test car is run over track, the computer
may be programmed to detect any measured or calculated force or
force relationship that deviate from an established standard.
TRACK MAINTENANCE
As an aid to track maintenance track conditions may be monitored by
using the same method. A measuring car equipped with trucks having
load cells, axle strain gauges and the necessary recording
equipment is run over the section of track for which conditions are
to be monitored. For this run a characteristic base is developed
which may take the form, for example, of the information shown in
FIG. 4 converted to a cumulative distribution. The results of
another test run performed at a suitable monitor test interval, for
example monthly, is then compared with the base run and any
significant deviation in the information from the base run, such as
magnitudes, frequencies, or occurrences of forces or ratios can be
detected. By use of the track side reference location detector
signal, the location of the deviation may be determined, or by use
of the CRT display 24, the computer 22 may be programmed to display
the deviation as it occurs. This method could be used to monitor
the condition of track joints, turnouts, ballast movement, ballast
resiliency and localized track defects, such as shelled spots,
engine burns, spalling of rail head and partial fracture of head.
Where defects create their particular pattern on strip charts, the
computer may be programmed to identify the defect.
* * * * *