separate a piece for the cap; the brasses are then simple half bushes or rings, held in position by being just cut away at the bolt holes, and, therefore, being held by the bolts. In the arrangement of Fig. 191, the rod end is enlarged to an oblong section, and the brasses B and B' are held between the end and the inside of a strap, S, which surrounds them and the oblong portion of the rod end, the strap is fixed by a cotter, C, and a gib, G, so that when the cotter is driven in, the gib draws the

strap towards the rod, and brings the brass B nearer the brass B'. This use of a gib and cotter has already been referred to in § 67, but owing to the cost of machining, it is more expensive than the design with bolts and nuts. It has recently been cheapened by cutting the slot for the cotter quite at the rod end, so that there is no metal between the brass and the slot. The strap is then rigidly bolted to the rod through a part removed from the brasses, and the gib is placed next the brass B', and

bearing against it, so that when the cotter is driven in it bears hard against the rod and strap on one edge, and on the other against the gib moving it and the brass B' towards the brass B.

When a rod has its crank pin end of the design of Fig. 191, its crosshead end is generally similar, and a rod so made is used in the horizontal engine of Ex. B, the crank pin end of the rod in Ex. A being of the form of Fig. 189, end C.P. (see Fig. 178).

(182) Lubrication of Connecting-Rod Brasses.-In Fig. 191 a siphon oil cup (see Fig. 153) is screwed into the top of the strap over the centre of the brasses, and oil is conducted to the front and back parts of the brasses which we know to be the seats of pressure, through narrow grooves cut in the top halves, leading from the hole under the oil cup; a horizontal groove being also cut nearly across the brasses exactly at the centre line. For vertical engines (say for the crank pin end) it is customary to fix a separate oil cup to the connecting-rod, and lead a small pipe from it to a hole drilled through the T end of the rod and the brass next it, a groove being cut from this hole around the brasses to the other side of the journal.

(183) Proportions of Parts.-In obtaining the sizes for the crossheads, guides, and connecting-rods, we require to know the total forces acting upon them when working. The stresses due to the steam pressure upon the piston alone are easily found, but these are increased by the effect of the inertia of the parts,

especially with high speeds, to an extent which must not be neglected, but which can only be found for particular examples by a method far too complicated for insertion here. We can, however, usefully show how to find the stresses caused by the steam pressure, and then by allowing a low value for the working stress per square inch, give some idea of how the sizes of the parts are obtained.

In Fig. 192, P R is the piston-rod, G and G' the guide blocks, CR the connecting-rod, and C the crank, S being the shaft

centre. (The crank and connecting-rod are drawn in the position when at right angles to each other.) If CT be drawn at right angles to RS, and, therefore, parallel to the pressure on the guides, then if the length of RT represents the total pressure on the piston, the length of CR will represent the force along the connecting-rod, and of CT the pressure on the botton guide. By drawing arrows representing the direction of the forces, it will be seen that in both the inward- and outward-stroke the guide pressure is on the bottom guide G when the engine runs in the direction of the arrow, the rod being subject to tensile stresses during the inward-stroke, and to compressive stresses during the outwardstroke. We thus see that a top guide is unnecessary so far as the pressure upon it is concerned, except when the engine reverses. Now draw CV to represent a shorter connecting-rod, then for the same position of the crank we see that, if the length of VT represents the total pressure on the piston, the length of CT represents, as before, the pressure on the guide. But this means that the guide pressure is the same as before, with a less pressure on the piston, or what amounts to the same thing, the shorter the connecting-rod, the greater the pressure on the guides. A short connecting-rod also gives a more irregular twisting force at the crank pin than a long rod, so that as long a connecting-rod as is conveniently possible should always be secured.

(184) Length of Connecting-Rods.-The length of connecting-rods between the centres for such engines as we are dealing with varies from five to seven times the crank length, the usual standard being six times.

Turning again to Fig. 192 we see that since the angles RCS and RTC are right angles, the triangles SCR and CTR are similar (CR being common), that is

also that

SC: CR RS:: CT: TR: RC,

RS = √RC2 + C S2, which = con.-rod length2 + crank length2,

and if the rod is 6 cranks length this equals√37, or 6·08. So that as RC is the total pressure on the piston when RS is the force along the connecting-rod, we see that with the usual length of rod the maximum force along the rod may be taken as equal to the total pressure on the piston (or more exactly to 6.08 = 1.013 times the piston pressure); also as CS represents

6

the maximum guide pressure we have, that this is equal to

CS CR

or th of the total pressure on the piston, or in a general case length of crank is equal to x total pressure on piston.* length of rod As the piston will not have reached the point of cut-off in the engines we are dealing with, when the rod and crank are at right angles, the pressure on the piston will be taken at its greatest value-that is, p × d2 x where p = effective pressure of steam 4'

per square inch before cut-off, and d = piston diameter.

We see then that the force along the connecting-rod may be taken as equal to the maximum total effective pressure on the piston when the steam alone is considered, and that this force has to be resisted by the crank pin and crosshead brasses. But, as previously mentioned, we must allow for the stresses due to inertia, and this we shall do by taking such a value for the working stress as shall cover these stresses, including those due to the steam. Professor Unwin has shown that the total stress in the connecting-rod is from 14 to 1 times the greatest pressure on the piston, as caused by the steam, but that the maximum guide pressure is not much increased.

(185) Area of Rubbing Surfaces.-The following figures for the working pressure upon the rubbing surfaces are ordinary practical values, and are in lbs. per square inch. The total pressure upon the surface is to be taken as that caused by the maximum effective steam pressure upon the piston only. The required area

in square inches will, therefore, in all cases be equal to

=

where d = piston diameter in inches, p maximum effective steam pressure per square inch, 8 = working pressure per square inch upon the guide.

(186) Guide Blocks.-The maximum pressure should not exceed 50 lbs. In many cases it is much lower, as other considerations in the design may require a large surface, such, for example, as in the box crosshead and trunk guide of Fig. 188.

(187) Crosshead Pin.-The effective area is only that of the length in the brasses, and is, therefore, that of the brasses also,

If the angle CRT, and TR total effective pressure on piston P, then

=

and these have their maximum value assuming P to be constant, when the angle RSC is a right angle.

which as already seen may be either in the crosshead or the connecting-rod. The maximum pressure should not exceed 2,000 lbs., a good working value being 1,200 to 1,500 lbs. If d = diameter of the pin, and 7 = length embraced by the brasses in inches, then effective area = dx l square inches.

(188) Crank Pin.-The maximum pressure should not exceed 600 lbs., a good working value being 400 lbs. If d = diameter length in inches, then effective area = dxl square

and l inches.

=

(189) Ratio of Diameter and Length of Crosshead and Crank Pin.—The calculation of the diameter and length of a crosshead or crank pin in order that the pressure upon them shall not produce too great a stress in the material, is another one of the questions which we cannot enter into here. We may, however, indicate that an ordinary crosshead or crank pin is practically in the condition of a beam loaded uniformly and supported at each end (assuming the crank to have double webs), and is, therefore, subject to bending, the bending moment increasing with the length, a crank pin being also subject to a twisting action. These considerations are fully treated in Prof. Unwin's "Machine Design."

Crosshead pins usually have a length of from 11 to 21 times their diameter, an average ratio being 2, or 7 2 d.

=

Crank pins usually have a length of from 1 to 1 times their diameter, an average ratio being 1}, or 7 = 1 d.

EXAMPLES.

EX. A 4. Make working drawings, two views, full size, of the crosshead and slipper guide for the vertical engine of Ex. A, as in Fig. 187. The crosshead block to be of mild steel forged in one piece, with the piston-rod 11" diameter. Length of block, 23". Pressure on crosshead pin 674 lbs. per square inch, 21d, brasses 3" thick. Two bolts, total area at bottom of threads area of piston-rod. Cap 14" thick. Guide plate 7" thick, 33′′ wide, pressure per square inch 35 lbs., fixed to crosshead by 2" screws. Lock nuts on bolts. Height from centre of rod to base, 3g".*

==

(Calculate length and diameter of brasses; these have flanges" thick overlapping crosshead, therefore width of crosshead = 1 - 1". Draw in brasses in both views. Find diameter of bolts taking effective area as at bottom of threads. Bolts are kept from moving by turning the crosshead where the rod joins it, forming a collar of about 24" diameter, and then

* The following are the pressures per square inch to be worked to for the examples of A and B:-Ex. A-Initial pressure, 107 lbs.; back pressure, 17 lbs.; effective initial pressure, 90 lbs. Ex. B-Initial pressure, 98 lbs.; back pressure, 2 lbs.; effective initial pressure, 96 lbs.