AN ATTEMPT

at a comparative assessment of the influence of external forces on track-gauge widening.

By the lecturer of the Tomsk Technological Institute,
engineer L. N. Lyubimov.

Introduction.

The question of track-gauge widening has repeatedly been given due attention, both in foreign and in our domestic technical literature. The valuable investigations of Couard*), Engerth and Spitz**) on French and German railways, on the one hand, and the detailed report of engineer Lebedev at the XV Advisory Congress of Engineers of the Track Service of the Russian railways, on the other, have in many respects illuminated this essentially important weak point of our railway track economy.

From a consideration of the measures applied, both on Russian and on foreign railways, against track-gauge widening, it becomes clear, however, that its complete elimination has not yet been achieved by any means. Nevertheless, even merely reducing the magnitude of the widening to possibly negligible dimensions is unquestionably necessary not only in the sense of track improvement, but also as a circumstance representing—

*) Coüard. Revue générale des chemins de fer. Août. 1896.
**) „Ueber das Wandern von Schienen bei Eisenbahngleisen“. Nach Vorträgen von H. Freiherr von Engerth und M. Spitz. Organ für die Fortschritte des Eisenbahnwesens.

which is undoubtedly a very substantial reduction in the expenditures for eliminating this evil, sometimes reaching rather impressive figures*). In all the above-mentioned investigations, however, no attention has been paid to that aspect of the question whose proper elucidation might perhaps facilitate the search for a more effective device for combating track creep. Namely: nowhere has there been a proper analysis of which of all the external forces acting on the railway track has the predominant significance? The aim of the present investigation is, as far as possible, to fill this substantial gap.

§ 1. Manifestation of creep. The longitudinal displacement of rails along the axis of the track, or so-called “creep,” is expressed in two ways: either in the form of movement of the rails along the axis over the sleepers and tie plates, or as displacement together with the sleepers. Creep, as is known, leads to complete derangement of the track, creates the need for extensive repairs, and serves as one of the principal causes of premature deterioration of the permanent way. Since the displacement of the rails occurs either simultaneously in both rail lines, more or less in parallel, or, conversely, one line runs ahead of the other, the very nature of the injuries caused by creep is connected with the character of such movement. Damage to the permanent way due to the first kind of creep consists chiefly in those defects caused by excessive increase or decrease of joint gaps, such as: fractures and breaks of rails at bolt holes, lateral curvature of the track, damage to splice bars, bolts, chairs, and screws, as well as derangement of the ballast and deterioration of the sleepers through the turning over of the latter.

The second kind of creep causes skewing of the joints and, depending on this, unequal loading of the sleepers under both rail lines, which in turn entails an uneven track.

§ 2. Magnitude of the greatest annual creep. This magnitude is evidently quite conditional: the circumstances that are the principal factors in the matter

*) Up to 150 rubles per verst, for example, on the Moscow–Yaroslavl Railway.

coal, can, on different railways, depending on local conditions, give such varied combinations that the greatest consumption may differ on each railway, not only for each place, but also for each year.

The greatest values of annual longitudinal displacements on foreign railways and on 7 Russian ones are indicated in Engineer Lebedev’s report to the XV Congress of Engineers of the Service of Ways*). In the table appended below are shown the values of the greatest annual creep which we happened personally to observe on two of the busiest railways of the Moscow railway junction: the Kursk and the Nizhny Novgorod. The values of maximum creep, on both the one and the other railway, are expressed almost identically, namely they reach a value of 0.143 sazhen = 305 m/m for the Kursk**) and 0.14 sazhen = 229 m/m.***) for the Nizhny Novgorod Railway. A comparison of these two values with the greatest creep of foreign railways and Russian ones, cited in the above-mentioned report, shows that it is equal to the greatest creep on the Prussian railways, on the Transcaucasian and Southeastern ones, and less than the creep on the Privislinsky railways.††) Unfortunately this comparison is, of course, very conditional, since it could have been made with a greater degree of reliability only after reducing all annual creep, so to speak, to a single denominator, i.e. by calculating it as a function, for example, of each thousand passing trains.

The conclusion that may be drawn from the table cited below, in general, only confirms conclusions made earlier by other investigators, and it would therefore be superfluous to dwell on them longer here.

*) Proceedings of the XV Congress of Engineers of the Service of Ways, p. 71.—1898.

**) 167th verst, Moscow track.

***) 17th verst, Nizhny Novgorod track.

††) 420 m/m.

Table of the annual runoff coefficient on the Moscow–Kursk and Nizhny Novgorod Railways (track on angular embankments).

§ 3. Determination of the resistance force of the superstructure at points.

The force of longitudinal displacement of the rails, as was said above (§ 1), may manifest itself in two ways: a) it moves only the rails along their axis, or b) it displaces the entire superstructure, including the ties.

If the weight of the superstructure over the length of one rail-track panel is denoted by \(P_o\); by \(L_o\), the weight of the locomotive; and by \(F_o\), the coefficient of friction—then the friction corresponding to such a weight may be expressed as:

\[ W_o = (P_o + L_o) F_o \ . \ . \ . \ . \ (I) \]

The weight of the superstructure over the length of one panel of both lines of 28-foot rail, with the weight of the latter taken as \(24\frac{1}{3}\) pounds per running foot, is composed of:

\[ \text{Weight of 2 rails} = \frac{2 \times 28 \times 24.33}{40} = 34.06 \text{ poods.} \]

\[ \text{„ 8 angle fish-plates} = \frac{22 \times 8}{40} = 4.40 \text{ „} \]

\[ \text{„ 24 bearing plates} = \frac{6.35 \times 24}{40} = 3.82 \text{ „} \]

\[ \text{„ 8 bolts} = \frac{1.6 \times 8}{40} = 0.32 \text{ „} \]

\[ \text{Total} \ . \ . \ . \ . \ . \ . \ . \ . \ 42.60 \text{ poods.} \]

This weight is increased by the weight of the rolling stock.

Assuming that the heaviest part of the first one, i.e. a 60-ton locomotive, is located over the panel, and taking the coefficient \(F_o = 0.25\), formula (I) takes the form:

\[ W_o = (42.60 + 60 \times 61.05) \times 0.25 \text{ poods} = (42.60 + 60 \times 61.05) \times 16.38 \times 0.25 \text{ kilograms} = 15174.43 \text{ kilograms} \ . \ . \ . \ . \ (A) \]

In the case of rails lying directly on the ties, without bearing plates, the coefficient of friction of the rail on the tie will be different. If it is denoted by \(F\), then formula (I) will take the form:

\[ W'_o = (P'_o + L_o)F \ . \ . \ . \ . \ (II) \]

\(P'_o\) will be determined from the preceding as:

\[ 42.60 - 3.82 = 38.78 \text{ poods} \]

\[ W'_o = (38.78 + 60 \times 61.05) \times 16.38 \times 0.50 = 30317.58 \ . \ . \ . \ (Б) \]

In deriving expressions (A) and (Б), the restraining force of the ties was neglected. Experience shows that, if one takes into account both this latter force and, in general, all the resistance represented by the heart-piece, then the quantities \(W_o\) and \(W'_o\) may be taken *) up to 50 percent higher, i.e., one may take

\[ W_o = \text{up to } 22000 \text{ kilograms} \qquad \ldots \qquad (\text{В}) \]

\[ W'_o = \text{up to } 45000 \text{ kilograms} \qquad \ldots \qquad (\text{Г}) \]

b) In the second case, i.e., when the force of the angle acts to shear the entire upper structure, including the sleepers, the friction opposing this shear will be:

\[ W''_o = (P_o + L_o + nG)\times F_o \qquad \ldots \qquad (\text{III}) \]

or:

\[ W'''_o = (P'_o + L_o + nG)\times F \qquad \ldots \qquad (\text{IV}) \]

where \(G\) is the weight of one sleeper, and \(n\) is the number of such sleepers under one rail link. Let, as before, \(P_o = 42.60\) poods; \(P'_o = 38.78\) poods; \(G = 3\) (the weight of a pine sleeper of section \(5 \times 3\) vershoks, length 1.25 sazhens); then:

\[ W''_o = (42.60 + 60 \times 61.05 + 12 \times 3)\times 16.38 \times 0.25 = 15321.85 \text{ kilograms.} \; (\text{Д}) \]

\[ W'''_o = (38.78 + 60 \times 61.05 + 12 \times 3)\times 16.38 \times 0.50 = 30612.42 \text{ kilograms.} \; (\text{Е}) \]

To this is also added the resistance of the ballast contained in the boxes between the sleepers, caused by the shearing of the latter. If the resistance of the ballast to the shear of one sleeper (per unit of its length) is denoted by \(E\), then, with sleeper length \(=1\) and the number of sleepers \(=n\), for the case of rails lying on tie-plates the following expression is obtained:

\[ W''''_o = (P_o + L + nG)F_o + nlE \qquad \ldots \qquad (\text{V}) \]

And for the case of rails lying directly on the sleepers:

\[ W'''''_o = (P'_o + L + nG)F + nlE \qquad \ldots \qquad (\text{VI}) \]

Since

\[ E = \frac{1}{2}h^2 \Delta \operatorname{tg}^2(45^\circ + \varphi/2) \]

**), where \(\Delta\) is the weight of a cubic unit of ballast; \(h\) is the height of the sleeper; \(\varphi\) is the angle of friction, then under the condition: sleeper height \(= 3\) vershoks \(= 0.0625\) sazhen; length \(= 1.25\) sazhens; \(\Delta\)—the weight of one cubic sazhen of ballast \(=1000\) poods, and \(\varphi = 37^\circ\)

) Proceedings of the XV Congress of Engineers of the Track Service, p. 69.
) Heusinger von Waldegg. Kalender für Eisenbahn-Techniker*. 1898, p. 1.

\[ nE = n\frac{h^{2}}{2}\Delta \tg^{2}\left(45^\circ+\frac{\varphi}{2}\right) =12\times1.25\times0.0625\,\frac{2}{2}\times1000\times\tg^{2}63^\circ30' = \]

\[ =12\times1.25\times0.0625\,\frac{2}{2}\times1000\times16.38\times4\ \text{kilograms} =1919.53\ \text{kg}. \]

And, consequently:

\[ W_o^{''''}=15321.85+1919.53=17241.38\ \text{kilograms}. \ldots \quad (\text{Ж}); \]

\[ W_o^{'''''}=30612.42+1919.53=32531.96\ \text{kilograms}. \ldots \quad (3). \]

Formulas (I)—(VI) thus make it possible to compute the resistance force to creeping for each particular case, and, when applied to the case of a single 28-foot rail link, give, as is evident from expressions (A) (Б) (В) (Г) (Д) (Е) (Ж) and (З), the forces counteracting creeping, of which, according to the preceding, the maximum is:

\[ W'_o=45000\ \text{kilograms}. \]

§ 4. The principal factors in the phenomenon of rail creeping. The direction of the longitudinal displacement of rails on roads with two-way traffic generally coincides, as is known, with the direction of train movement; on single-track roads this direction is almost always toward the predominant traffic. Thus, rail creeping is above all in direct dependence on those circumstances that accompany the movement of trains as they proceed along the track. Such circumstances are:

a) the force of resistance to the motion of the train.
b) the impact of the wheels of the rolling stock.
c) the force of friction during braking.
d) the design of locomotives, making itself felt almost exclusively in the uneven creeping of both rail lines of one and the same track.

Then, although to a much lesser degree, the totality of various local conditions also acts on rail creeping, such as: the influence of ballast, the construction of the superstructure of the track, climate, etc.; consideration of these is not included in the present investigation, since they have already been sufficiently thoroughly examined in their time in the works of both Russian and foreign engineers.

§ 5. Rail creeping from resistance to the movement of the train. The friction between the wheels of the rolling stock and the rails, which causes the movement of the train, gives rise to a force striving, as stated above, to cree-

to crush the joints in the direction of motion. The magnitude of this force may be expressed by a formula of the form:

\[ \Phi = F_o P \ .\ .\ .\ .\ (\mathrm{VII}) \]

where \(\Phi\) is the tractive force; \(F_o\) is the coefficient of friction between the bandage and the rail, and \(P\) is the weight of the train. The coefficient \(F_o\), according to Galton’s experiments, may be taken on average as equal to \(0.25^*)\).

If we further take the weight of the locomotive with tender as equal to 100 tons**), then, for the case of motion of this locomotive alone, the tractive force is expressed as:

\[ \Phi = 0.25 \times 100 \times 1000 = 25000 \text{ kilograms} \ .\ .\ (\mathrm{И}) \]

Since such an expression for the tractive force would be only approximately correct, in view of the circumstance that formula (VII) is derived entirely apart from the speed of motion, the profile of the track, and the influence of the other part of the train, then for a more exact determination of the tractive force from the resistance of the train to motion it is best to apply Professor Petrov’s formula***).

We have, for a freight train:

\[ W = (4.3 + 0.15v + 0.001v^2)L + 1.2Q + 0.9nv + 0.03(1 + 0.04n)v^2 + \]

\[ + \left(\pm i + 21\frac{4\frac{1}{2} + l^2}{R - 45} + 0.2 - 0.015T\right)(L + Q) \ .\ .\ .\ .\ (\mathrm{VIII}). \]

Here \(W\) is the total resistance of the whole train, expressed in kilograms;

\(L\)—weight of the locomotive with tender in tons.

\(n\)—number of cars.

\(l\)—distance between axles not changing their relative position in the car, expressed in meters.

\(\pm i\)—number of thousandths on an ascent or descent****); in this, the sign \((+)\) refers to ascents, and \((-)\) to down grades.

\(R\)—radius of curvature of the track in meters.

\(T\)—temperature of the atmosphere in degrees Celsius.

\(v\)—speed of the whole train in kilometers per hour.

*) This coefficient varies within the limits \(0.14\)—\(0.33\).

**) Circular of the Department of Railways of January 15, 1896, No. 753.
Weight of locomotive 60 tons. Length between buffers 9.3 m. Weight of tender 37.50 tons. Length 6.8 m.

***) “Resistance of a Train on the Railway,” p. 332.

****) i.e., for example, with a gradient \(=0.008\); \(i = 8\).

Let, as before, \(L=100\) tons; \(v=40\) kilometers; \(n=30\); \(Q=480^*)\) tons; \(i=8\); \(R=800\) meters; \(l=3.8\); \(T=20^\circ\) C. Then formula (VIII) will take the form:

\[ \begin{aligned} W={}&(4+0.15\times 40+0.001\times 40^2)\times 100+1.2\times 480+0.9\times 30\times \\ &\times 40+0.03(1+0.04\times 30)\times 40^2+\left(-8+21\times \frac{4\times 3.8+3.8^2}{800-45}+\right.\\ &\left.+0.2-0.015\times 20\right)(100+480)=1190+576+1080+1056+\\ &+(-8+0.824+0.2-0.3)\times 580=-1268.48\ \text{or}\ +8011.52\ \text{kilogr.} \end{aligned} \]

It follows from this that the driving force for both threads of the rail track in the case of a downgrade \(=1266.48\) kilograms . . . . . . . . . . . . . . (K)
And in the case of the train moving on the same upgrade \(=8011.52\) kil. . . (L)

This force acts simultaneously over the length, properly speaking, of the whole train, i.e. \(16.1^m+30\times 7.6^m=244.1^m=114.48\) sazhen, or (with 28-foot rails)—over 29 joints and, consequently, only a part of it falls simultaneously on each joint. But in view of the circumstance that the successive passage of the wheels of the whole train over one and the same joint takes place in a very short interval of time—it is more correct to assume that the full driving force, calculated according to (K) or (L), acts on the joint.

§ 6. Driving of rails by impacts of the wheels of rolling stock when joints deflect. The impact of the wheels of rolling stock occurs, as is known, chiefly from insufficient rigidity of the connection at rail joints: when the joint deflects (see fig. 1 on the separate sheet of drawings), the wheel, after lowering one of the joint sleepers in its motion, must, as it were, climb up an incline at the following joint sleeper; in this case the force of the impact which the wheel produces on the rails must inevitably push the latter forward. This will occur especially in those cases when the clearance between adjacent rails is too large: then, as the wheel approaches the joint (fig. 2), the end of the rail against the direction of travel will be higher than that of the rail in the direction of travel, thereby forming a ledge against which the edge of the tire will strike.

The magnitude of the impact force is evidently proportional, chiefly, to the weight of the rolling stock and to the speed of its motion.

If one considers only a single locomotive and assumes its connection with the tender and the remaining part of the train to be perfectly elastic, and takes, as before, the weight of the locomotive equal to \(L\) tons; the speed of motion \(=v\) ki-

*) Taking the weight of a car with full load as 16 tons.

kilometers per hour; \(g\) is the acceleration of gravity—then the force of the impact at the joint of the two wheels may be expressed by the formula:

\[ T=\frac{Lv}{g}\ldots\ldots (IX) \]

With the weight of the locomotive equal to 60 tons \(=60{,}000\) kilograms; \(v=40\) kilometers \(=40{,}000\) meters per hour, and \(g=9.81\) meters per second,

\[ T=\frac{60{,}000\times40{,}000}{9.81\times60\times60}=67957.87 \text{ kilograms}\ldots\ldots (H). \]

It is obvious that the magnitude of the impact force may also be influenced by the mass of the remaining part of the train.

§ 7. Lateral displacement from impacts of the wheels of rolling stock due to canting of the rail ends. From the mismatch of the rail profiles at a joint, owing to the transverse inclination of such rails under the load of the wheels, there occurs, as it were, a twisting of the transverse end of the rail. As is known, under the pressure of the wheels the rails incline somewhat toward the inside of the track, which thereby narrows. According to observations on French railways*) it appears that such an inclination amounts on average to from \(1—2.5\) m/m, and this inclination at the rail ends is greater than in the middle, i.e. there occurs, as it were, a “canting” of its end. Since the fishplates do not possess sufficient rigidity (even under the condition of their closest fit to the rail web) to transmit this twisting force to the adjacent rail, the transverse end at the joint (fig. 3) inclines more than the opposite one, thus forming a step from which the wheel runs off. As a result of this, especially at high speeds, the wheels fall onto the opposite rail end at a certain distance (from \(1 \tfrac{1}{2}''—3''\)) from the joint (fig. 4), forming indentations in the heads of these ends, impacts against which will, evidently, contribute to displacement.

§ 8. Displacement of rails from impacts of the wheels of rolling stock against steps at joints, occurring as a result of mismatch of the rail profiles. The increase of the impact force at joints is undoubtedly affected also by the unequal height of adjacent rail ends, which is the consequence either of careless rolling of new rails, or else of laying, during single replacement, old rails with an unequal degree of wear. According to the latest technical conditions for acceptance of rails*) a difference in their height of not more than \(0.5\) m/m is permitted. Even such a comparatively

) According to Cuoir’s observations on the Paris—Lyon—Méditerranée railway.
) “Temporary technical conditions for the supply of steel rails,” § 2—Bulletin of the Ministry of Railways, Series No. 1, 1900.

...an altogether negligible deflection nevertheless has a tangible effect: thus

when a wheel with a pressure of \(\frac{15}{2} = 7.5\) tons strikes such a deflection, the living force developed is \(7.5 \times 1000 \times 0.0005 = 3.75\) kilogram-meters.

§ 9. Derailment from impacts of improperly worn tires. It is not uncommon for the tires of the wheels of rolling stock to wear down far from uniformly. The impacts received by the wheels when passing over joints, and especially the friction against the rails of wheels that are too strongly braked, very often alter the running surface of the tires, sometimes giving them a form that deviates considerably from the shape of a proper cone having its geometrical axis—the axis of rotation of the wheels—and sometimes even producing actual flat spots in them.

In this respect the action of air brakes, which produce pressure on the wheels of the whole train more or less uniformly, is much weaker than the action of hand brakes.

During the entire motion of a rolling wheel, so long as its contact with the rails occurs on the regular parts of the surface of the tire, the center of gravity of the wheel will describe a line equidistant from the surface of the rail, and therefore the force pressing the wheel against the rail will do no work. But as soon as the wheel touches the rail with the flattened part of the tire, an impact follows, which in turn forms on the surface of the rails a whole series of small dents, thanks to which subsequently rolling stock, even with regular tires, will strike these dents and thereby contribute to derailment.

§ 10. Derailment from braking force. When braking wheels rolling on rails, as is known, the resistance offered by the latter to translational motion increases. This resistance \(R\) cannot exceed the adhesion force \(FP\) (§ 5), otherwise there would occur simple sliding of the wheels of the rolling stock along the rails. The force \(R\) obviously reaches its greatest value when the brake shoes are pressed to the highest limit at which the wheels still roll along the rails, and with the slightest increase of which they would begin to slide. The coefficient of adhesion of wheels with rails varies depending on the condition of the rail surface, does not depend on the speed of motion, and fluctuates, as was said above, within the limits from \(0.14\)—\(0.33\), and on average may be taken equal, for dry rails, to \(0.25\).

Taking this latter value of the coefficient \(F\), it is not difficult to see that with \(N\%\) of braked axles in the train, the greatest horizontal eff...

... transmitted during braking of an entire train may reach the value

\[ R = 0.25 \times N \times P \ . \ . \ . \]

or

\[ R = 0.25 \times N[L+Q] \ . \ . (X) \]

where, as before, \(L\) is the weight of the locomotive with tender, and \(Q\) is the weight of the cars in tons.

For the particular case considered in § 5, \(L = 100\) tons, \(Q = 480\) tons; \(n\), the number of cars, is 30.

If we take into account that, according to our ministerial regulations, at a running speed of 40 versts per hour and a grade of 0.008, the smallest number of braked axles must be 23, then for the present case

\[ N = \frac{23}{2 \times n} = 0.38 \]

and

\[ R = 0.25 \times 0.38 \times (100 + 480) \times 1000 = 55100 \text{ kilograms} \ . \ . (M) \]

At the present time, when the system of continuous brakes has already become so widely used that on some railways, in courier and fast trains, all car axles are already equipped with brake shoes, the braking force can obviously attain, in other cases, far greater dimensions than those indicated by expression (M), and, consequently, can contribute still more strongly to creeping.

§ 11. Comparison of the found creeping forces.

From consideration of the present table it is not difficult to draw the following conclusions:

I. The resistance of the superstructure, under other favorable conditions, may prove sufficient for adequately counteracting the angle resulting from the resistance to the motion of the train.

II. The resistance of the superstructure is not in a condition to contend with the horizontal forces caused by the blows of the wheels of the rolling stock and by braking.

III. The greatest pressure on the rail angle is, all other circumstances being equal, the same horizontal force generated by the blows of the wheels of the rolling stock.

IV. With an increase in the number of brake axles it is possible that the angle force from braking not only equals the angle force from the blows of the wheels of the rolling stock, but even exceeds it.

Thus all measures for combating the angle caused by the first three causes considered above in § 4 must chiefly be directed, above all, toward eliminating the influence of the force of the blows of the wheels of the rolling stock by means of:

a) The unconditional withdrawal from service of wheels with broken flanges.

b) The possible strengthening of the rigidity of joints by improving the systems of fishplates, the quality of the ballast, and various kinds of devices capable, at least indirectly, of contributing to the force of resistance of the superstructure.

c) Strict sorting before laying new rails.

d) Strict selection of old rails in the case of individual replacement.

e) Paying proper attention to combating the action of braking, namely: 1) imposing on engine drivers and conductor brigades of freight trains equipped with hand brakes the duty of braking on gradients, as far as possible, in advance and uniformly, and not of suddenly applying the brakes at the last, i.e. critical, minute, when, owing to an excessively developed speed, it becomes necessary to brake with full force; 2)—the unconditional prohibition to engine drivers of passenger and especially express trains equipped with continuous brakes—of braking almost at the last minute before the very stopping of the train at a station, before a place on the track for which a warning of slow running has been issued, before large bridges, etc.

§ 12. Anti-creep fishplate of Engineer Lyubimov’s system*).

The fishplate of this system belongs to the number of means against creeping listed under rubric б in the preceding paragraph. When—

*) Protective certificate No. 10370.

this device consists (see the separate sheet of drawings, figs. 5, 6, 7 and 8) of two sheets A and A¹ of boiler iron, ²⁄₈ inch thick, 21 inches high and 8.4 inches wide each. To sheet A¹, by means of 6 rivets 1/4 inch in diameter, are riveted two strips of boiler iron a and a¹ (figs. 5 and 6), 21 inches long, 1¹⁄₂ inches wide and ²⁄₈ inch thick. These strips form a groove into which sheet A may be inserted. To the same sheets A and A¹, by means of 5 rivets, are riveted strips B and B¹, 21 inches long, 2 inches wide and ¹⁄₂ inch thick, provided at their upper ends with lugs of the same iron (2″ by 1/2″). These lugs are formed by twisting the natural continuation of the same strips B and B¹ and thus constitute, with the latter, one whole. In each of them two holes, 1 inch in diameter, are made, through which bolts D, ⁷⁄₈ inch in diameter and 5¹⁄₈ inches long, may be passed.

The method of using the anti-derailment fishplate of the described type is as follows: in places of the track where an angularity of the rails is observed, the track is first brought in the proper manner into its normal condition; then, depending on the magnitude of the angularity existing at the given place, either at every joint, or at every other joint, or finally at every two joints, anti-derailment fishplates of the described type are placed in the joints. For this purpose one pair of bolts D in the shaped fishplates bb of the joint (fig. 7) is unbolted on any one side of the joint; both halves A and A¹ of the anti-derailment fishplate are lowered into the ballast. By striking lightly with a sledgehammer, sheet A is made to enter the groove aa¹ of sheet A¹, and the lugs are bolted with the above-mentioned bolts D to the shaped fishplates bb¹, and consequently also to the very rail of the track.

In the event that the ballast has become too compacted and does not permit the direct insertion of the halves A and A¹, proceed in the following manner: in the box between the joint sleepers a small pit is dug out, and sheets A and A¹ are inserted into it, proceeding thereafter as indicated above, and then merely ensuring that the ballast poured back is tightly tamped.

In the described form this device may be manufactured entirely by the railway’s own means, both because of the simplicity of its construction and because of the constant availability, within any railway, of those materials from which the fishplate of the indicated system is made. Its installation in place is carried out at the rail joint, and not somewhere in the middle between such joints; that is, in other words, no additional holes are made in the rails that could weaken their section and serve as a cause of

such fractures. The fishplate does not hinder the proper tamping of the ends of the sleepers.

The anti-creep device was tested over the course of one working spring, summer, and autumn period under the following conditions: the track profile was two-verst, with a continuous gradient of 0.004. Curve \(P = 500\). Traffic was double-track. The volume of traffic was more than 30 trains on each track. There was heavy braking of passenger trains because of the need to stop at a dacha platform located on a short level stretch at the very end of the gradient. On the described part of the track, before the installation of anti-creep fishplates of the indicated system, severe rail creep was observed \((0\text{ s}.14 = 299\ \mathrm{m}/\mathrm{m})\). The track was brought into complete order, the joints were staggered according to the square; then, on a part of it, fishplates of the described system were installed across the joints. On the said part the creep ceased completely, whereas in the other places along the verst the creep resumed. No bending of the lugs B in the fishplates, nor any damage in the plates A, was observed after six months of lying in the track. The action of the fishplate may be explained, on the one hand, by a certain increase in the stiffness of the joint itself, and, on the other hand, as it were, by a certain increase in the resistance of the ballast itself. Without prejudging the possibility of the fully successful action of the same fishplate on steeper gradients as well, we confine ourselves here only to establishing the fact of its useful action under the conditions described above.

§ 13. Influence of locomotive design on creep. This influence, as has already been noted in passing (§ 4), is expressed chiefly in the unevenness of the creep of one rail line or another, i.e. in the “lead” of one relative to the other. The examination of this question has been the subject of several rather detailed articles in the foreign (chiefly German) literature; moreover, the authors of these, Borries) and Engerth and Spitz*), observed almost exclusively a lead in the left rail line (reckoning according to the direction of movement), regardless of whether it faced toward the shoulder or toward the middle of the track, and attribute such a lead to the design of locomotives on the Austro-Hungarian railways, where the crankpin of the right-hand mechanism is set 90° ahead relative to the position of the crankpin on the left side, as a result of which the left side of the locomotive—as, incidentally, was also confirmed by the greater wear of the left tires of the first wheels—worked considerably more strongly than the right.

) Organ für die Fortschritte des Eisenbahnwesens*. 1897. Heft 12, “Ueber den Einfluss der Lokomotiven auf das Wandern der Schienen.” Von Borries.

**) 1897. Heft 9, “Ueber das Wandern von Schienen bei Eisenbahngleisen.” Von H. Freiherr, von Engerth und Ingenieur M. Spitz.

For our part, we have personally had occasion to observe, on the Moscow–Nizhny Novgorod Railway, the advance of the right-hand rail line, caused likewise by a structural defect in the locomotives running on this railway, but of a somewhat different character from that just described. The point is that, before 1891, no unequal creeping was observed on the straight portions of the Moscow section of this railway. In 1891 new locomotives of Series E, built by the Kolomna works, of the Compound system, began to enter service on the railway. Some time after the start of operation of these latter, creep began to be noticed on the straight portions of the track, expressed in the right-hand line of rails overtaking the left-hand line. Since the outer line on all straight portions of this railway is elevated, in comparison with the left, by 0.002–0.003, such an advance could not find a plausible explanation in a greater settlement of the outer line on the shoulder of the roadbed, as a comparatively weaker place of the track. We then conceived the idea of seeking the cause of such an advance in the design of the above-mentioned locomotives. It turned out that locomotives of this series have, as do almost all Compound locomotives, a left cylinder of considerably greater size than the right. Equalization of the work of the one and the other cylinder is achieved by the fact that live steam from the boiler enters only the small right cylinder, while the large left one works already with exhaust steam; moreover, this steam loses its original pressure to a considerable degree, namely from a pressure of 11 atmospheres it falls to \(5 \tfrac{1}{2}\), or even 5 atmospheres. Equalization of the work of both sides of the locomotive at such greatly differing pressures is achieved, as was said above, by the unequal dimensions of the cylinders working with live and exhaust steam. But the fact is that such a circumstance, for the above-described Series E, as we have ascertained, is unconditionally true only at the limiting pressure in the boiler of 11 atmospheres. But as soon as the pressure of the live steam is even only a little below the limiting value (which in practice happens very often), the equalization of the work of the right and left sides no longer exists—and, as it turns out in practice, the work of the live steam on the right side of the locomotive is considerably greater than the work of the exhaust steam on the left. This, too, explains, as should be supposed, the advance precisely of the right-hand line of rails.

Plate II. Observations of F. [[unclear: surname]].

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