Fly-wheel or Shaft Governors.-At the Centennial Exhibition in 1876 there were shown a few steam-engines in which the governors were contained in the fly-wheel or band-wheel, the fly-balls or weights revolving around the shaft in a vertical plane with the wheel and shifting the eccentric so as automatically to vary the travel of the valve and the point of cutoff. This form of governor has since come into extensive use, especially for high-speed engines. In its usual form two weights are carried on arms the ends of which are pivoted to two points on the pulley near its circumference, 180° apart. Links connect these arms to the eccentric. The eccentric is not rigidly keyed to the shaft but is free to move transversely across it for a certain distance, having an oblong hole which allows of this movement. Centrifugal force causes the weights to fly towards the circumference of the wheel and to pull the eccentric into a position of minimum eccentricity. This force is resisted by a spring attached to each arm which tends to pull the weights towards the shaft and shift the eccentric to the position of maximum eccentricity. The travel of the valve is thus varied, so that it tends to cut off earlier in the stroke as the engine increases its speed. Many modifications of this general form are in use. For discussions of this form of governor see Hartnell, Proc. Inst. M. E., 1882, p. 408; Trans. A. S. M. E., ix. 300; xi. 1081; xiv. 92; xv. 929; Modern Mechanism, p. 399; Whitham's Constructive Steam Engineering; J. Begtrup, 4m. Mach.

Oct. 19 and Dec. 14, 1893, Jan. 18 and March 1, 1894.

Calculation of Springs for Shaft-governors. (Wilson Hartnell, Proc. Inst. M. E., Aug. 1882.)-The springs for shaft-governors may be conveniently calculated as follows, dimensions being in inches:

Let W = weight of the balls or weights, in pounds;

71 and r the maximum and minimum radial distances of the centre of the balls or of the centre of gravity of the weights;

l1 and l = the leverages, i.e., the perpendicular distances from the centre of the weight-pin to a line in the direction of the centrifugal force drawn through the centre of gravity of the weights or balls at radii r1 and r

m, and m, the corresponding leverages of the springs;

C1 and C2

the centrifugal forces, for 100 revolutions per minute, at radii r1 and r2;

P1 and P2 = the corresponding pressures on the spring;

(It is convenient to calculate these and note them down for reference.) C3 and C4 = maximum and minimum centrifugal forces;

S

mean speed (revolutions per minute);

S1 and S2 the maximum and minimum number of revolutions per minute;

P and P1 = the pressures on the spring at the limiting number of revo lutions (S, and S2);

P-PD

the difference of the maximum and minimum pressures

on the springs;

the percentage of variation from the mean speed, or the sensitive

ness;

V

The mean speed and sensitiveness desired are supposed to be given. Then

It is usual to give the spring-maker the values of P, and of v or w. Το ensure proper space being provided, the dimensions of the spring should be

calculated by the formulæ for strength and extension of springs, and the least length of the spring as compressed be determined.

With a straight centripetal line, the governor-power

For a preliminary determination of the governor-power it may be taken as equal to this in all cases, although it is evident that with a curved centripetal line it will be slightly less. The difference D must be constant for the same spring, however great or little its initial compression. Let the spring be screwed up until its minimum pressure is P. Then to find the speed PP + D,

The speed at which the governor would be isochronous would be

Suppose the pressure on the spring with a speed of 100 revolutions, at the maximum and minimum radii, was 200 lbs. and 100 lbs., respectively, then the pressure of the spring to suit a variation from 95 to 105 revolutions will be 100 X = 220.5. That is, the increase

(95)

100

90.2 and 200 X

(105) *2 =

of resistance from the minimum to the maximum radius must be 220 - 90= 130 lbs.

The extreme speeds due to such a spring, screwed up to different pressures, are shown in the following table:

Revolutions per minute, balls shut..
Pressure on springs, balls shut....

Increase of pressure when balls open fully.
Pressure on springs, balls open fully.
Revolutions per minute, balls open fully.
Variation, per cent of mean speed

80 90 95 100 110 120 64 81 90 100 121 144 130 130 130 130 130 130 194 211 220 230 251 274 98 102 105 107 112 117 10 6 5 3 1 -1

The speed at which the governor would become isochronous is 114. Any spring will give the right variation at some speed; hence in experimenting with a governor the correct spring may be found from any wrong one by a very simple calculation. Thus, if a governor with a spring whose stiffness is 50 lbs. per inch acts best when the engine runs at 95, 90 being its proper speed, then 50 X 45 lbs. is the stiffness of spring required.

2

To determine the speed at which the governor acts best, the springs may be screwed up until it begins to "hunt" and then slackened until the gov ernor is as sensitive as is compatible with steadiness.

CONDENSERS, AIR-PUMPS CIRCULATING-
PUMPS, ETC.

The Jet Condenser. (Chiefly abridged from Seaton's Marine Engineering.) The jet condenser is now uncommon in marine practice, being generally supplanted by the surface condenser. It is commonly used where fresh water is available for boiler feed. With the jet condenser a vacuum of 24 in. was considered fairly good, and 25 in, as much as was possible with most condensers; the temperature corresponding to 24 in. vacuum, or 3 lbs. pressure absolute, is 140°. In practice the temperature in the hot-well varies from 110° to 120°, and occasionally as much as 130° is maintained. To find the quantity of injection-water per pound of steam to be condensed: Let T = temperature of steam at the exhaust pressure; To temperature of the cooling.

=

water; T, temperature of the water after condensation, or of the hot-well; = pounds of the cooling-water per lb. of steam condensed; then

WH

9 R

T2- To

Another formula is: Q = in which W is the weight of steam condensed, H the units of heat given up by 1 lb. of steam in condensing, and R the rise in temperature of the cooling-water.

This is applicable both to jet and to surface condensers. The allowance made for the injection-water of engines working in the temperate zone is usually 27 to 30 times the weight of steam, and for the tropics 30 to 35 times; 30 times is sufficient for ships which are occasionally in the tropics, and this is what was usual to allow for general traders.

Area of injection orifice =

to 780.

weight of injection-water in lbs. per min. + 650

A rough rule sometimes used is: Allow one fifteenth of a square inch for every cubic foot of water condensed per hour.

Another rule: Area of injection orifice = area of piston + 250.

The volume of the jet condenser is from one fourth to one half of that of the cylinder. It need not be more than one third, except for very quickrunning engines.

Ejector Condensers.-For ejector or injector condensers (Bulkley's, Schutte's, etc.) the calculations for quantity of condensing-water is the same as for jet condensers.

The Surface Condenser-Cooling Surface.-Peclet found that with cooling water of an initial temperature of 68° to 77°, one sq. ft. of copper plate condensed 21.5 lbs. of steam per hour, while Joule states that 100 lbs. per hour can be condensed. In practice, with the compound engine, brass condenser-tubes, 18 B.W.G. thick, 13 lbs. of steam per sq. ft. per hour, with the cooling-water at an initial temperature of 600, is considered very fair work when the temperature of the feed-water is to be maintained at 120°. It has been found that the surface in the condenser may be half the heating surface of the boiler, and under some circumstances considerably less than this. In general practice the following holds good when the temperature of sea-water is about 60°:

Terminal pres., lbs., abs.... 30 20 15 1216 10 8 6 Sq. ft. per I.H.P.. 8 2.50 2.25 2.00 1.80 1.60 1.50 For ships whose station is in the tropics the allowance should be increased by 20%, and for ships which occasionally visit the tropics 10% increase will give satisfactory results. If a ship is constantly employed in cold climates 10% less suffices.

Whitham (Steam-engine Design, p. 283, also Trans. A. S. M. E., ix. 431) gives the following: S =

WL
ck(T1-t)'

in which S

condensing-surface in sq.

ft.; T1 = temperature Fahr. of steam of the pressure indicated by the vacuum gauge; t = mean temperature of the circulating water, or the arithmetical mean of the initial and final temperatures; L = latent heat of saturated steam at temperature T1; k = perfect conductivity of 1 sq. ft. of the metal used for the condensing surface for a range of 1° F. (or 557 B.T.U. per hour for brass, according to Isherwood's experiments); c = fraction denoting the efficiency of the condensing surface; W = pounds of steam condensed per hour. From experiments by Loring and Emery, on U.S.S. Dallas. c is found to be 0.323, and ck = 180; making the equation S =

WL

180(T1-t)

Whitham recommends this formula for designing engines having independent circulating pumps. When the pump is worked by the main engine the value of S should be increased about 10%.

Taking T, at 135° F., and L = 1020, corresponding to 25 in. vacuum, and t 1020 W for summer temperatures at 75°, we have: S = 180(135-75)

17 W
180

For a mathematical discussion of the efficiency of surface condensers see a paper by T. E. Stanton in Proc. Inst. C. E., cxxxvi, June 1899, p. 321. Condenser Tubes are generally made of solid-drawn brass tubes, and tested both by hydraulic pressure and steam. They are usually made of a composition of 68% of best selected copper and 32% of best Silesian spelter.

The Admiralty, however, always specify the tubes to be made of 70% of best selected copper and to have 1% of tin in the composition, and test the tubes to a pressure of 300 lbs. per sq. in. (Seaton.)

The diameter of the condenser tubes varies from 1⁄2 inch in small condensers, when they are very short, to 1 inch in very large condensers and long tubes. In the mercantile marine the tubes are, as a rule, 34 inch diameter externally, and 18 B.W.G. thick (0.049 inch); and 16 B.W.G. (0.065), under some exceptional circumstances. In the British Navy the tubes are also, as a rule, 34 inch diameter, and 18 to 19 B. W.G. thick, tinned on both sides; when the condenser is made of brass the.Admiralty do not require the tubes to be tinned. Some of the smaller engines have tubes 5% inch diameter, and 19 B.W.G. thick. The smaller the tubes, the larger is the surface which can be got in a certain space.

In the merchant service the almost universal practice is to circulate the water through the tubes.

Whitham says the velocity of flow through the tubes should not be less than 400 nor more than 700 ft. per min.

Tube-plates are usually made of brass. Rolled-brass tube-plates should be from 1.1 to 1.5 times the diameter of tubes in thickness, depending on the method of packing. When the packings go completely through the plates the latter, but when only partly through the former, is sufficient. Hence, for 34-inch tubes the plates are usually to 1 inch thick with glands and tape-packings, and 1 to 114 inch thick with wooden ferrules.

The tube-plates should be secured to their seatings by brass studs and nuts, or brass screw-bolts; in fact there must be no wrought iron of any kind inside a condenser. When the tube-plates are of large area it is advisable to stay them by brass-rods, to prevent them from collapsing.

Spacing of Tubes, etc.-The holes for ferrules, glands, or indiarubber are usually 4 inch larger in diameter than the tubes; but when absolutely necessary the wood ferrules may be only 3/32 inch thick.

The pitch of tubes when packed with wood ferrules is usually 4 inch more than the diameter of the ferrule-hole. For example, the tubes are generally arranged zigzag, and the number which may be fitted into & square foot of plate is as follows:

Quantity of Cooling Water.-The quantity depends chiefly upon its initial temperature, which in Atlantic practice may vary from 40° in the winter of temperate zone to 80° in subtropical seas. To raise the temperature to 100° in the condenser will require three times as many thermal units in the former case as in the latter, and therefore only one third as much cooling-water will be required in the former case as in the latter.

T1 temperature of steam entering the condenser;

"circulating-water entering the condenser;
leaving the condenser;
"water condensed from the steam;

Q quantity of circulating water in lbs. =

11140.3(T-T3).
T2- To

It is usual to provide pumping power sufficient to supply 40 times the weight of steam for general traders, and as much as 50 times for ships stationed in subtropical seas, when the engines are compound. If the circulating-pump is double-acting, its capacity may be 1/53 in the former and 1/42 in the latter case of the capacity of the low-pressure cylinder.

Air-pump.-The air-pump in all condensers abstracts the water condensed and the air originally contained in the water when it entered the boiler. In the case of jet-condensers it also pumps out the water of condensation and the air which it contained. The size of the pump is calculated from these conditions, making allowance for efficiency of the pump.

Ordinary sea-water contains, mechanically mixed with it, 1/20 of its volume of air when under the atmospheric pressure. Suppose the pressure in the condenser to be 2 lbs. and the atmospheric pressure 15 lbs., neglecting the effect of temperature, the air on entering the condenser will be expanded to 15/2 times its original volume; so that a cubic foot of sea-water, when it has entered the condenser, is represented by 19/20 of a cubic foot of water and 15/40 of a cubic foot of air.

Let q be the volume of water condensed per minute, and Q the volume of sea water required to condense it; and let T be the temperature of the condenser, and T, that of the sea-water.

Then 19/20 (q+Q) will be the volume of water to be pumped from the condenser per minute,

If the temperature of the condenser be taken at 120°, and that of seawater at 60°, the quantity of air will then be .418(g + Q), so that the total volume to be abstracted will be

.95(g+Q)+.418(q + Q) = 1.68(g + Q).

If the average quantity of injection-water be taken at 26 times that condensed, q+Q will equal 27q. Therefore, volume to be pumped from the condenser per minute = 37q, nearly.

In surface condensation allowance must be made for the water occasion. ally admitted to the boilers to make up for waste, and the air contained in it, also for slight leak in the joints and glands, so that the air-pump is made about half as large as for jet-condensation.

The efficiency of a single-acting air-pump is generally taken at 0.5, and that of a double-acting pump at 0.35. When the temperatur of the sea is 60°, and that of the (jet) condenser is 120°, Q being the volume of the cooling water and q the volume of the condensed water in cubic feet, and n the number of strokes per minute,

The volume of the single-acting pump = 2.74(+9).

n

The volume of the double-acting pump = 4 4(2 + 2).

n

The following table gives the ratio of capacity of cylinder or cylinders to that of the air-pump; in the case of the compound engine, the low-pressure cylinder capacity only is taken.

The Area through Valve-seats and past the valves should not be less than will admit the full quantity of water for condensation at a velocity not exceeding 400 ft. per ininute. In practice the area is generally in excess of this.

Area through foot-valves = D2 X S1000 square inches.
Area through head-valves D2 XS 800 square inches.

Diameter of discharge-pipe = DX VS+ 35 inches.

D= diam. of air-pump in inches, S = its speed in ft. per min.

James Tribe (Am. Much., Oct. 8, 1891) gives the following rule for air.