1911 Encyclopædia Britannica/Hydraulics

by William Cawthorne Unwin

HYDRAULICS (Gr. ὕδωρ, water, and αὐλός, a pipe), the branch of engineering science which deals with the practical applications of the laws of hydromechanics.

I. THE DATA OF HYDRAULICS^[1]

§ 1. Properties of Fluids.—The fluids to which the laws of practical hydraulics relate are substances the parts of which possess very great mobility, or which offer a very small resistance to distortion independently of inertia. Under the general heading Hydromechanics a fluid is defined to be a substance which yields continually to the slightest tangential stress, and hence in a fluid at rest there can be no tangential stress. But, further, in fluids such as water, air, steam, &c., to which the present division of the article relates, the tangential stresses that are called into action between contiguous portions during distortion or change of figure are always small compared with the weight, inertia, pressure, &c., which produce the visible motions it is the object of hydraulics to estimate. On the other hand, while a fluid passes easily from one form to another, it opposes considerable resistance to change of volume.

It is easily deduced from the absence or smallness of the tangential stress that contiguous portions of fluid act on each other with a pressure which is exactly or very nearly normal to the interface which separates them. The stress must be a pressure, not a tension, or the parts would separate. Further, at any point in a fluid the pressure in all directions must be the same; or, in other words, the pressure on any small element of surface is independent of the orientation of the surface.

§ 2. Fluids are divided into liquids, or incompressible fluids, and gases, or compressible fluids. Very great changes of pressure change the volume of liquids only by a small amount, and if the pressure on them is reduced to zero they do not sensibly dilate. In gases or compressible fluids the volume alters sensibly for small changes of pressure, and if the pressure is indefinitely diminished they dilate without limit.

In ordinary hydraulics, liquids are treated as absolutely incompressible. In dealing with gases the changes of volume which accompany changes of pressure must be taken into account.

§ 3. Viscous fluids are those in which change of form under a continued stress proceeds gradually and increases indefinitely. A very viscous fluid opposes great resistance to change of form in a short time, and yet may be deformed considerably by a small stress acting for a long period. A block of pitch is more easily splintered than indented by a hammer, but under the action of the mere weight of its parts acting for a long enough time it flattens out and flows like a liquid.

Fig. 1.

All actual fluids are viscous. They oppose a resistance to the relative motion of their parts. This resistance diminishes with the velocity of the relative motion, and becomes zero in a fluid the parts of which are relatively at rest. When the relative motion of different parts of a fluid is small, the viscosity may be neglected without introducing important errors. On the other hand, where there is considerable relative motion, the viscosity may be expected to have an influence too great to be neglected.

Measurement of Viscosity. Coefficient of Viscosity.—Suppose the plane ab, fig. 1 of area ω, to move with the velocity V relatively to the surface cd and parallel to it. Let the space between be filled with liquid. The layers of liquid in contact with ab and cd adhere to them. The intermediate layers all offering an equal resistance to shearing or distortion, the rectangle of fluid abcd will take the form of the parallelogram a′b′cd. Further, the resistance to the motion of ab may be expressed in the form

$\mathrm {R} =\kappa \omega \mathrm {V}$,

(1)

where $\kappa$ is a coefficient the nature of which remains to be determined.

If we suppose the liquid between ab and cd divided into layers as shown in fig. 2, it will be clear that the stress R acts, at each dividing face, forwards in the direction of motion if we consider the upper layer, backwards if we consider the lower layer. Now suppose the original thickness of the layer T increased to nT; if the bounding plane in its new position has the velocity nV, the shearing at each dividing face will be exactly the same as before, and the resistance must therefore be the same. Hence,

$\mathrm {R} =\kappa '\omega \left(n\mathrm {V} \right).$

(2)

But equations (1) and (2) may both be expressed in one equation if $\kappa$ and $\kappa '$ are replaced by a constant varying inversely as the thickness of the layer. Putting ${\begin{aligned}\kappa &=\mu /\mathrm {T} ,\kappa '=\mu /n\mathrm {T} ,\\\mathrm {R} &=\mu \omega \mathrm {V} /\mathrm {T} ;\\\end{aligned}}$

or, for an indefinitely thin layer,

$\mathrm {R} =\mu \omega d\mathrm {V} /dt,$

(3)

an expression first proposed by L. M. H. Navier. The coefficient μ is termed the coefficient of viscosity.

According to J. Clerk Maxwell, the value of μ for air at θ° Fahr. in pounds, when the velocities are expressed in feet per second, is

$\mu =0.0000000256\left(461^{\circ }+\theta \right);$

that is, the coefficient of viscosity is proportional to the absolute temperature and independent of the pressure.

The value of μ for water at 77° Fahr. is, according to H. von Helmholtz and G. Piotrowski,

$\mu =0.0000188,$

the units being the same as before. For water μ decreases rapidly with increase of temperature.

Fig. 2.

§ 4. When a fluid flows in a very regular manner, as for instance when it flows in a capillary tube, the velocities vary gradually at any moment from one point of the fluid to a neighbouring point. The layer adjacent to the sides of the tube adheres to it and is at rest. The layers more interior than this slide on each other. But the resistance developed by these regular movements is very small. If in large pipes and open channels there were a similar regularity of movement, the neighbouring filaments would acquire, especially near the sides, very great relative velocities. V. J. Boussinesq has shown that the central filament in a semicircular canal of 1 metre radius, and inclined at a slope of only 0.0001, would have a velocity of 187 metres per second,^[2] the layer next the boundary remaining at rest. But before such a difference of velocity can arise, the motion of the fluid becomes much more complicated. Volumes of fluid are detached continually from the boundaries, and, revolving, form eddies traversing the fluid in all directions, and sliding with finite relative velocities against those surrounding them. These slidings develop resistances incomparably greater than the viscous resistance due to movements varying continuously from point to point. The movements which produce the phenomena commonly ascribed to fluid friction must be regarded as rapidly or even suddenly varying from one point to another. The internal resistances to the motion of the fluid do not depend merely on the general velocities of translation at different points of the fluid (or what Boussinesq terms the mean local velocities), but rather on the intensity at each point of the eddying agitation. The problems of hydraulics are therefore much more complicated than problems in which a regular motion of the fluid is assumed, hindered by the viscosity of the fluid.

Relation of Pressure, Density, and Temperature of Liquids

§ 5. Units of Volume.—In practical calculations the cubic foot and gallon are largely used, and in metric countries the litre and cubic metre (= 1000 litres). The imperial gallon is now exclusively used in England, but the United States have retained the old English wine gallon.

1 cub. ft.	= 6·236 imp. gallons	= 7·481 U.S. gallons.
1 imp. gallon	= 0·1605 cub. ft.	= 1·200 U.S. gallons.
1 U.S. gallon	= 0·1337 cub. ft.	= 0·8333 imp. gallon.
1 litre	= 0·2201 imp. gallon	= 0·2641 U.S. gallon.

Density of Water.—Water at 53° F. and ordinary pressure contains 62·4 ℔ per cub. ft., or 10 ℔ per imperial gallon at 62° F. The litre contains one kilogram of water at 4° C. or 1000 kilograms per cubic metre. River and spring water is not sensibly denser than pure water. But average sea water weighs 64 ℔ per cub. ft. at 53° F. The weight of water per cubic unit will be denoted by G. Ice free from air weighs 57·28 ℔ per cub. ft. (Leduc).

§ 6. Compressibility of Liquids.—The most accurate experiments show that liquids are sensibly compressed by very great pressures, and that up to a pressure of 65 atmospheres, or about 1000 ℔ per sq. in., the compression is proportional to the pressure. The chief results of experiment are given in the following table. Let V₁ be the volume of a liquid in cubic feet under a pressure p₁ ℔ per sq. ft., and V₂ its volume under a pressure p₂. Then the cubical compression is (V₂ − V₁)/V₁, and the ratio of the increase of pressure p₂ − p₁ to the cubical compression is sensibly constant. That is, k = (p₂ − p₁)V₁/(V₂ − V₁) is constant. This constant is termed the elasticity of volume. With the notation of the differential calculus,

k = dp / ( −	dV	) = − V	dp	.
	V		dV

Elasticity of Volume of Liquids.

	Canton.	Oersted.	Colladon and Sturm.	Regnault.
Water	45,990,000	45,900,000	42,660,000	44,000,000
Sea water	52,900,000	··	··	··
Mercury	705,300,000	··	626,100,000	604,500,000
Oil	44,090,000	··	··	··
Alcohol	32,060,000	··	23,100,000	··

According to the experiments of Grassi, the compressibility of water diminishes as the temperature increases, while that of ether, alcohol and chloroform is increased.

§ 7. Change of Volume and Density of Water with Change of Temperature.—Although the change of volume of water with change of temperature is so small that it may generally be neglected in ordinary hydraulic calculations, yet it should be noted that there is a change of volume which should be allowed for in very exact calculations. The values of ρ in the following short table, which gives data enough for hydraulic purposes, are taken from Professor Everett’s System of Units.

Density of Water at Different Temperatures.

Temperature.		ρ Density of Water.	G  Weight of  1 cub. ft. in ℔.
Cent.	Fahr.
0	32·0	·999884	62·417
1	33·8	·999941	62·420
2	35·6	·999982	62·423
3	37·4	1·000004	62·424
4	39·2	1·000013	62·425
5	41·0	1·000003	62·424
6	42·8	·999983	62·423
7	44·6	·999946	62·421
8	46·4	·999899	62·418
9	48·2	·999837	62·414
10	50·0	·999760	62·409
11	51·8	·999668	62·403
12	53·6	·999562	62·397
13	55·4	·999443	62·389
14	57·2	·999312	62·381
15	59·0	·999173	62·373
16	60·8	·999015	62·363
17	62·6	·998854	62·353
18	64·4	·998667	62·341
19	66·2	·998473	62·329
20	68·0	·998272	62·316
22	71·6	·997839	62·289
24	75·2	·997380	62·261
26	78·8	·996879	62·229
28	82·4	·996344	62·196
30	86	·995778	62·161
35	95	·99469	62·093
40	104	·99236	61·947
45	113	·99038	61·823
50	122	·98821	61·688
55	131	·98583	61·540
60	140	·98339	61·387
65	149	·98075	61·222
70	158	·97795	61·048
75	167	·97499	60·863
80	176	·97195	60·674
85	185	·96880	60·477
90	194	·96557	60·275
100	212	·95866	59·844

The weight per cubic foot has been calculated from the values of ρ, on the assumption that 1 cub. ft. of water at 39·2° Fahr. is 62·425 ℔. For ordinary calculations in hydraulics, the density of water (which will in future be designated by the symbol G) will be taken at 62·4 ℔ per cub. ft., which is its density at 53° Fahr. It may be noted also that ice at 32° Fahr. contains 57·3 ℔ per cub. ft. The values of ρ are the densities in grammes per cubic centimetre.

§ 8. Pressure Column. Free Surface Level.—Suppose a small vertical pipe introduced into a liquid at any point P (fig. 3). Then the liquid will rise in the pipe to a level OO, such that the pressure due to the column in the pipe exactly balances the pressure on its mouth. If the fluid is in motion the mouth of the pipe must be supposed accurately parallel to the direction of motion, or the impact of the liquid at the mouth of the pipe will have an influence on the height of the column. If this condition is complied with, the height h of the column is a measure of the pressure at the point P. Let ω be the area of section of the pipe, h the height of the pressure column, p the intensity of pressure at P; then

$p\omega =\mathrm {G} h\omega$℔,
$p/\mathrm {G} =h$;

that is, h is the height due to the pressure at p. The level OO will be termed the free surface level corresponding to the pressure at P.

Relation of Pressure, Temperature, and Density of Gases

     Fig. 3.

§ 9. Relation of Pressure, Volume, Temperature and Density in Compressible Fluids.—Certain problems on the flow of air and steam are so similar to those relating to the flow of water that they are conveniently treated together. It is necessary, therefore, to state as briefly as possible the properties of compressible fluids so far as knowledge of them is requisite in the solution of these problems. Air may be taken as a type of these fluids, and the numerical data here given will relate to air.

Relation of Pressure and Volume at Constant Temperature.—At constant temperature the product of the pressure p and volume V of a given quantity of air is a constant (Boyle’s law).

Let $p_{0}$ be mean atmospheric pressure (2116·8 ℔ per sq. ft.), $\mathrm {V} _{0}$ the volume of 1 ℔ of air at 32° Fahr. under the pressure $p_{0}$. Then

$p_{0}\mathrm {V} _{0}=26214$.

(1)

If $G_{0}$ is the weight per cubic foot of air in the same conditions,

$\mathrm {G} _{0}=1/\mathrm {V} _{0}=2116.8/26214=.0807$.

(2)

For any other pressure $p$, at which the volume of 1 ℔ is $\mathrm {V}$ and the weight per cubic foot is $\mathrm {G}$, the temperature being 32° Fahr.,

$p\mathrm {V} =p/\mathrm {G} =26214$; or $\mathrm {G} =p/26214$.

(3)

Change of Pressure or Volume by Change of Temperature.—Let p₀, V₀, G₀, as before be the pressure, the volume of a pound in cubic feet, and the weight of a cubic foot in pounds, at 32° Fahr. Let p, V, G be the same quantities at a temperature t (measured strictly by the air thermometer, the degrees of which differ a little from those of a mercurial thermometer). Then, by experiment,

$p\mathrm {V} =p_{0}\mathrm {V} _{0}(460.6+t)/(460.6+32)=p_{0}\mathrm {V} _{0}\tau /\tau _{0}$,

(4)

where $\tau ,\tau _{0}$ are the temperatures t and 32° reckoned from the absolute zero, which is −460·6° Fahr.;

$p/\mathrm {G} =p_{0}\tau /\mathrm {G} _{0}\tau _{0}$;

(4a)

$\mathrm {G} =p\tau _{0}\mathrm {G} _{0}/p_{0}\tau$.

(5)

If $\tau _{0}=2116.8,\mathrm {G} _{0}=.08075,\tau _{0}=460.6+32=492.6$, then

$p/\mathrm {G} =53.2\tau$.

(5a)

Or quite generally $p/\mathrm {G} =\mathrm {R} \tau$ for all gases, if $\mathrm {R}$ is a constant varying inversely as the density of the gas at 32° F. For steam $\mathrm {R}$ = 85·5.

II. KINEMATICS OF FLUIDS

§ 10. Moving fluids as commonly observed are conveniently classified thus:

(1) Streams are moving masses of indefinite length, completely or incompletely bounded laterally by solid boundaries. When the solid boundaries are complete, the flow is said to take place in a pipe. When the solid boundary is incomplete and leaves the upper surface of the fluid free, it is termed a stream bed or channel or canal.

(2) A stream bounded laterally by differently moving fluid of the same kind is termed a current.

(3) A jet is a stream bounded by fluid of a different kind.

(4) An eddy, vortex or whirlpool is a mass of fluid the particles of which are moving circularly or spirally.

(5) In a stream we may often regard the particles as flowing along definite paths in space. A chain of particles following each other along such a constant path may be termed a fluid filament or elementary stream.

§ 11. Steady and Unsteady, Uniform and Varying, Motion.—There are two quite distinct ways of treating hydrodynamical questions. We may either fix attention on a given mass of fluid and consider its changes of position and energy under the action of the stresses to which it is subjected, or we may have regard to a given fixed portion of space, and consider the volume and energy of the fluid entering and leaving that space.

Fig. 4.

If, in following a given path ab (fig. 4), a mass of water a has a constant velocity, the motion is said to be uniform. The kinetic energy of the mass a remains unchanged. If the velocity varies from point to point of the path, the motion is called varying motion. If at a given point a in space, the particles of water always arrive with the same velocity and in the same direction, during any given time, then the motion is termed steady motion. On the contrary, if at the point a the velocity or direction varies from moment to moment the motion is termed unsteady. A river which excavates its own bed is in unsteady motion so long as the slope and form of the bed is changing. It, however, tends always towards a condition in which the bed ceases to change, and it is then said to have reached a condition of permanent regime. No river probably is in absolutely permanent regime, except perhaps in rocky channels. In other cases the bed is scoured more or less during the rise of a flood, and silted again during the subsidence of the flood. But while many streams of a torrential character change the condition of their bed often and to a large extent, in others the changes are comparatively small and not easily observed.

As a stream approaches a condition of steady motion, its regime becomes permanent. Hence steady motion and permanent regime are sometimes used as meaning the same thing. The one, however, is a definite term applicable to the motion of the water, the other a less definite term applicable in strictness only to the condition of the stream bed.

§ 12. Theoretical Notions on the Motion of Water.—The actual motion of the particles of water is in most cases very complex. To simplify hydrodynamic problems, simpler modes of motion are assumed, and the results of theory so obtained are compared experimentally with the actual motions.

Fig. 5.

Motion in Plane Layers.—The simplest kind of motion in a stream is one in which the particles initially situated in any plane cross section of the stream continue to be found in plane cross sections during the subsequent motion. Thus, if the particles in a thin plane layer ab (fig. 5) are found again in a thin plane layer a′b′ after any interval of time, the motion is said to be motion in plane layers. In such motion the internal work in deforming the layer may usually be disregarded, and the resistance to the motion is confined to the circumference.

Laminar Motion.—In the case of streams having solid boundaries, it is observed that the central parts move faster than the lateral parts. To take account of these differences of velocity, the stream may be conceived to be divided into thin laminae, having cross sections somewhat similar to the solid boundary of the stream, and sliding on each other. The different laminae can then be treated as having differing velocities according to any law either observed or deduced from their mutual friction. A much closer approximation to the real motion of ordinary streams is thus obtained.

Stream Line Motion.—In the preceding hypothesis, all the particles in each lamina have the same velocity at any given cross section of the stream. If this assumption is abandoned, the cross section of the stream must be supposed divided into indefinitely small areas, each representing the section of a fluid filament. Then these filaments may have any law of variation of velocity assigned to them. If the motion is steady motion these fluid filaments (or as they are then termed stream lines) will have fixed positions in space.

Fig. 6.

Periodic Unsteady Motion.—In ordinary streams with rough boundaries, it is observed that at any given point the velocity varies from moment to moment in magnitude and direction, but that the average velocity for a sensible period (say for 5 or 10 minutes) varies very little either in magnitude or velocity. It has hence been conceived that the variations of direction and magnitude of the velocity are periodic, and that, if for each point of the stream the mean velocity and direction of motion were substituted for the actual more or less varying motions, the motion of the stream might be treated as steady stream line or steady laminar motion.

§ 13. Volume of Flow.—Let A (fig. 6) be any ideal plane surface, of area ω, in a stream, normal to the direction of motion, and let V be the velocity of the fluid. Then the volume flowing through the surface A in unit time is

Q = ωV.

(1)

Thus, if the motion is rectilinear, all the particles at any instant in the surface A will be found after one second in a similar surface A′, at a distance V, and as each particle is followed by a continuous thread of other particles, the volume of flow is the right prism AA′ having a base ω and length V.

If the direction of motion makes an angle θ with the normal to the surface, the volume of flow is represented by an oblique prism AA′ (fig. 7), and in that case

Q = ωV cos θ.

Fig. 7.

If the velocity varies at different points of the surface, let the surface be divided into very small portions, for each of which the velocity may be regarded as constant. If dω is the area and v, or v cos θ, the normal velocity for this element of the surface, the volume of flow is

Q = ∫ v dω, or ∫ v cos θ dω,

as the case may be.

§ 14. Principle of Continuity.—If we consider any completely bounded fixed space in a moving liquid initially and finally filled continuously with liquid, the inflow must be equal to the outflow. Expressing the inflow with a positive and the outflow with a negative sign, and estimating the volume of flow Q for all the boundaries,

ΣQ = 0.

In general the space will remain filled with fluid if the pressure at every point remains positive. There will be a break of continuity, if at any point the pressure becomes negative, indicating that the stress at that point is tensile. In the case of ordinary water this statement requires modification. Water contains a variable amount of air in solution, often about one-twentieth of its volume. This air is disengaged and breaks the continuity of the liquid, if the pressure falls below a point corresponding to its tension. It is for this reason that pumps will not draw water to the full height due to atmospheric pressure.

Application of the Principle of Continuity to the case of a Stream.—If A₁, A₂ are the areas of two normal cross sections of a stream, and V₁, V₂ are the velocities of the stream at those sections, then from the principle of continuity,

V₁A₁ = V₂A₂;

V₁/V₂ = A₂/A₁

(2)

that is, the normal velocities are inversely as the areas of the cross sections. This is true of the mean velocities, if at each section the velocity of the stream varies. In a river of varying slope the velocity varies with the slope. It is easy therefore to see that in parts of large cross section the slope is smaller than in parts of small cross section.

If we conceive a space in a liquid bounded by normal sections at A₁, A₂ and between A₁, A₂ by stream lines (fig. 8), then, as there is no flow across the stream lines,

V₁/V₂ = A₂/A₁,

as in a stream with rigid boundaries.

Fig. 8.

In the case of compressible fluids the variation of volume due to the difference of pressure at the two sections must be taken into account. If the motion is steady the weight of fluid between two cross sections of a stream must remain constant. Hence the weight flowing in must be the same as the weight flowing out. Let p₁, p₂ be the pressures, v₁, v₂ the velocities, G₁, G₂ the weight per cubic foot of fluid, at cross sections of a stream of areas A₁, A₂. The volumes of inflow and outflow are

A₁v₁ and A₂v₂,

and, if the weights of these are the same,

G₁A₁v₁ = G₂A₂v₂;

and hence, from (5a) § 9, if the temperature is constant,

p₁A₁v₁ = p₂A₂v₂.

(3)

Fig. 9.

Fig. 10.	Fig. 11.	Fig. 12.

Fig. 13.

§ 15. Stream Lines.—The characteristic of a perfect fluid, that is, a fluid free from viscosity, is that the pressure between any two parts into which it is divided by a plane must be normal to the plane. One consequence of this is that the particles can have no rotation impressed upon them, and the motion of such a fluid is irrotational. A stream line is the line, straight or curved, traced by a particle in a current of fluid in irrotational movement. In a steady current each stream line preserves its figure and position unchanged, and marks the track of a stream of particles forming a fluid filament or elementary stream. A current in steady irrotational movement may be conceived to be divided by insensibly thin partitions following the course of the stream lines into a number of elementary streams. If the positions of these partitions are so adjusted that the volumes of flow in all the elementary streams are equal, they represent to the mind the velocity as well as the direction of motion of the particles in different parts of the current, for the velocities are inversely proportional to the cross sections of the elementary streams. No actual fluid is devoid of viscosity, and the effect of viscosity is to render the motion of a fluid sinuous, or rotational or eddying under most ordinary conditions. At very low velocities in a tube of moderate size the motion of water may be nearly pure stream line motion. But at some velocity, smaller as the diameter of the tube is greater, the motion suddenly becomes tumultuous. The laws of simple stream line motion have hitherto been investigated theoretically, and from mathematical difficulties have only been determined for certain simple cases. Professor H. S. Hele Shaw has found means of exhibiting stream line motion in a number of very interesting cases experimentally. Generally in these experiments a thin sheet of fluid is caused to flow between two parallel plates of glass. In the earlier experiments streams of very small air bubbles introduced into the water current rendered visible the motions of the water. By the use of a lantern the image of a portion of the current can be shown on a screen or photographed. In later experiments streams of coloured liquid at regular distances were introduced into the sheet and these much more clearly marked out the forms of the stream lines. With a fluid sheet 0.02 in. thick, the stream lines were found to be stable at almost any required velocity. For certain simple cases Professor Hele Shaw has shown that the experimental stream lines of a viscous fluid are so far as can be measured identical with the calculated stream lines of a perfect fluid. Sir G. G. Stokes pointed out that in this case, either from the thinness of the stream between its glass walls, or the slowness of the motion, or the high viscosity of the liquid, or from a combination of all these, the flow is regular, and the effects of inertia disappear, the viscosity dominating everything. Glycerine gives the stream lines very satisfactorily.

Fig. 9 shows the stream lines of a sheet of fluid passing a fairly shipshape body such as a screwshaft strut. The arrow shows the direction of motion of the fluid. Fig. 10 shows the stream lines for a very thin glycerine sheet passing a non-shipshape body, the stream lines being practically perfect. Fig. 11 shows one of the earlier air-bubble experiments with a thicker sheet of water. In this case the stream lines break up behind the obstruction, forming an eddying wake. Fig. 12 shows the stream lines of a fluid passing a sudden contraction or sudden enlargement of a pipe. Lastly, fig. 13 shows the stream lines of a current passing an oblique plane. H. S. Hele Shaw, “Experiments on the Nature of the Surface Resistance in Pipes and on Ships,” Trans. Inst. Naval Arch. (1897). “Investigation of Stream Line Motion under certain Experimental Conditions,” Trans. Inst. Naval Arch. (1898); “Stream Line Motion of a Viscous Fluid,” Report of British Association (1898).

III. PHENOMENA OF THE DISCHARGE OF LIQUIDS FROM ORIFICES AS ASCERTAINABLE BY EXPERIMENTS

Fig. 14.

§ 16. When a liquid issues vertically from a small orifice, it forms a jet which rises nearly to the level of the free surface of the liquid in the vessel from which it flows. The difference of level h_r (fig. 14) is so small that it may be at once suspected to be due either to air resistance on the surface of the jet or to the viscosity of the liquid or to friction against the sides of the orifice. Neglecting for the moment this small quantity, we may infer, from the elevation of the jet, that each molecule on leaving the orifice possessed the velocity required to lift it against gravity to the height h. From ordinary dynamics, the relation between the velocity and height of projection is given by the equation

v = √2gh.

(1)

As this velocity is nearly reached in the flow from well-formed orifices, it is sometimes called the theoretical velocity of discharge. This relation was first obtained by Torricelli.

If the orifice is of a suitable conoidal form, the water issues in filaments normal to the plane of the orifice. Let ω be the area of the orifice, then the discharge per second must be, from eq. (1),

Q = ωv = ω√2gh nearly.

(2)

This is sometimes quite improperly called the theoretical discharge for any kind of orifice. Except for a well-formed conoidal orifice the result is not approximate even, so that if it is supposed to be based on a theory the theory is a false one.

Use of the term Head in Hydraulics.—The term head is an old millwright’s term, and meant primarily the height through which a mass of water descended in actuating a hydraulic machine. Since the water in fig. 14 descends through a height h to the orifice, we may say there are h ft. of head above the orifice. Still more generally any mass of liquid h ft. above a horizontal plane may be said to have h ft. of elevation head relatively to that datum plane. Further, since the pressure p at the orifice which produces outflow is connected with h by the relation p/G = h, the quantity p/G may be termed the pressure head at the orifice. Lastly, the velocity v is connected with h by the relation v²/2g = h, so that v²/2g may be termed the head due to the velocity v.

§ 17. Coefficients of Velocity and Resistance.—As the actual velocity of discharge differs from √2gh by a small quantity, let the actual velocity

= v_a = c_v √2gh,

(3)

where c_v is a coefficient to be determined by experiment, called the coefficient of velocity. This coefficient is found to be tolerably constant for different heads with well-formed simple orifices, and it very often has the value 0.97.

The difference between the velocity of discharge and the velocity due to the head may be reckoned in another way. The total height h causing outflow consists of two parts—one part h_e expended effectively in producing the velocity of outflow, another h_r in overcoming the resistances due to viscosity and friction. Let

h_r = c_rh_e,

where c_r is a coefficient determined by experiment, and called the coefficient of resistance of the orifice. It is tolerably constant for different heads with well-formed orifices. Then

v_a = √2gh_e = √ { 2gh / (1 + c_r) }.

(4)

The relation between c_v and c_r for any orifice is easily found:—

v_a = c_v√2gh = √ { 2gh / (1 + c_r) }

   c_v = √ { 1 / (1 + c_r) }

(5)

cr = 1 / cv2 − 1.

(5a)

Thus if c_v = 0.97, then c_r = 0.0628. That is, for such an orifice about 61/4% of the head is expended in overcoming frictional resistances to flow.

Fig. 15.

Coefficient of Contraction—Sharp-edged Orifices in Plane Surfaces.—When a jet issues from an aperture in a vessel, it may either spring clear from the inner edge of the orifice as at a or b (fig. 15), or it may adhere to the sides of the orifice as at c. The former condition will be found if the orifice is bevelled outwards as at a, so as to be sharp edged, and it will also occur generally for a prismatic aperture like b, provided the thickness of the plate in which the aperture is formed is less than the diameter of the jet. But if the thickness is greater the condition shown at c will occur.

When the discharge occurs as at a or b, the filaments converging towards the orifice continue to converge beyond it, so that the section of the jet where the filaments have become parallel is smaller than the section of the orifice. The inertia of the filaments opposes sudden change of direction of motion at the edge of the orifice, and the convergence continues for a distance of about half the diameter of the orifice beyond it. Let ω be the area of the orifice, and c_cω the area of the jet at the point where convergence ceases; then c_c is a coefficient to be determined experimentally for each kind of orifice, called the coefficient of contraction. When the orifice is a sharp-edged orifice in a plane surface, the value of c_c is on the average 0.64, or the section of the jet is very nearly five-eighths of the area of the orifice.

Fig. 16.

Coefficient of Discharge.—In applying the general formula Q = ωv to a stream, it is assumed that the filaments have a common velocity v normal to the section ω. But if the jet contracts, it is at the contracted section of the jet that the direction of motion is normal to a transverse section of the jet. Hence the actual discharge when contraction occurs is

Q_a = c_vv × c_cω = c_cc_vω √(2gh),

or simply, if c = c_vc_c,

Q_a = cω √(2gh),

where c is called the coefficient of discharge. Thus for a sharp-edged plane orifice c = 0.97 × 0.64 = 0.62.

§ 18. Experimental Determination of _v, c_c, and c.—The coefficient of contraction c_c is directly determined by measuring the dimensions of the jet. For this purpose fixed screws of fine pitch (fig. 16) are convenient. These are set to touch the jet, and then the distance between them can be measured at leisure.

The coefficient of velocity is determined directly by measuring the parabolic path of a horizontal jet.

Let OX, OY (fig. 17) be horizontal and vertical axes, the origin being at the orifice. Let h be the head, and x, y the coordinates of a point A on the parabolic path of the jet. If v_a is the velocity at the orifice, and t the time in which a particle moves from O to A, then

x = v_at; y = 1/2 gt².

Eliminating t,

v_a = √ (gx²/2y).

Then

c_v = v_a √ (2gh) = √ (x²/4yh).

In the case of large orifices such as weirs, the velocity can be directly determined by using a Pitot tube (§ 144).

Fig. 17.

The coefficient of discharge, which for practical purposes is the most important of the three coefficients, is best determined by tank measurement of the flow from the given orifice in a suitable time. If Q is the discharge measured in the tank per second, then

c = Q/ω √ (2gh).

Measurements of this kind though simple in principle are not free from some practical difficulties, and require much care. In fig. 18 is shown an arrangement of measuring tank. The orifice is fixed in the wall of the cistern A and discharges either into the waste channel BB, or into the measuring tank. There is a short trough on rollers C which when run under the jet directs the discharge into the tank, and when run back again allows the discharge to drop into the waste channel. D is a stilling screen to prevent agitation of the surface at the measuring point, E, and F is a discharge valve for emptying the measuring tank. The rise of level in the tank, the time of the flow and the head over the orifice at that time must be exactly observed.

Fig. 18.

For well made sharp-edged orifices, small relatively to the water surface in the supply reservoir, the coefficients under different conditions of head are pretty exactly known. Suppose the same quantity of water is made to flow in succession through such an orifice and through another orifice of which the coefficient is required, and when the rate of flow is constant the heads over each orifice are noted. Let h₁, h₂ be the heads, ω₁, ω₂ the areas of the orifices, c₁, c₂ the coefficients. Then since the flow through each orifice is the same

Q = c₁ω₁ √ (2gh₁) = c₂ω₂ √ (2gh₂).

c₂ = c₁ (ω₁/ω₂) √ (h₁/h₂).

Fig. 19.

§ 19. Coefficients for Bellmouths and Bellmouthed Orifices.—If an orifice is furnished with a mouthpiece exactly of the form of the contracted vein, then the whole of the contraction occurs within the mouthpiece, and if the area of the orifice is measured at the smaller end, c_c must be put = 1. It is often desirable to bellmouth the ends of pipes, to avoid the loss of head which occurs if this is not done; and such a bellmouth may also have the form of the contracted jet. Fig. 19 shows the proportions of such a bellmouth or bell-mouthed orifice, which approximates to the form of the contracted jet sufficiently for any practical purpose.

For such an orifice L. J. Weisbach found the following values of the coefficients with different heads.

Head over orifice, in ft. = h	·66	1·64	11·48	55·77	337·93
Coefficient of velocity = cv	·959	·967	·975	·994	·994
Coefficient of resistance = cr	·087	·069	·052	·012	·012

As there is no contraction after the jet issues from the orifice, c_c = 1, c = c_v; and therefore

Q = c_vω √ (2gh) = ω √ { 2gh / (1 + c_r }.

§ 20. Coefficients for Sharp-edged or virtually Sharp-edged Orifices.—There are a very large number of measurements of discharge from sharp-edged orifices under different conditions of head. An account of these and a very careful tabulation of the average values of the coefficients will be found in the Hydraulics of the late Hamilton Smith (Wiley & Sons, New York, 1886). The following short table abstracted from a larger one will give a fair notion of how the coefficient varies according to the most trustworthy of the experiments.

Coefficient of Discharge for Vertical Circular Orifices, Sharp-edged,
with free Discharge into the Air. Q = cω √ (2gh).

Head  measured to  Centre of Orifice.	Diameters of Orifice.
	·02	·04	·10	·20	·40	·60	1·0
	Values of C.
0·3	..	..	·621	..	..	..	..
0·4	..	·637	·618	..	..	..	..
0·6	·655	·630	·613	·601	·596	·588	..
0·8	·648	·626	·610	·601	·597	·594	·583
1·0	·644	·623	·608	·600	·598	·595	·591
2·0	·632	·614	·604	·599	·599	·597	·595
4·0	·623	·609	·602	·599	·598	·597	·596
8·0	·614	·605	·600	·598	·597	·596	·596
20·0	·601	·599	·596	·596	·596	·596	·594

At the same time it must be observed that differences of sharpness in the edge of the orifice and some other circumstances affect the results, so that the values found by different careful experimenters are not a little discrepant. When exact measurement of flow has to be made by a sharp-edged orifice it is desirable that the coefficient for the particular orifice should be directly determined.

The following results were obtained by Dr H. T. Bovey in the laboratory of McGill University.

Coefficient of Discharge for Sharp-edged Orifices.

Head in  ft.	Form of Orifice.
	Circular.	Square.		Rectangular Ratio  of Sides 4:1		Rectangular Ratio  of Sides 16:1		Tri-  angular.
		Sides Vertical.	Diagonal vertical.	Long Sides vertical.	Long Sides hori- zontal.	Long Sides vertical.	Long Sides  hori-  zontal.
1	·620	·627	·628	·642	·643	·663	·664	·636
2	·613	·620	·628	·634	·636	·650	·651	·628
4	·608	·616	·618	·628	·629	·641	·642	·623
6	·607	·614	·616	·626	·627	·637	·637	·620
8	·606	·613	·614	·623	·625	·634	·635	·619
10	·605	·612	·613	·622	·624	·632	·633	·618
12	·604	·611	·612	·622	·623	·631	·631	·618
14	·604	·610	·612	·621	·622	·630	·630	·618
16	·603	·610	·611	·620	·622	·630	·630	·617
18	·603	·610	·611	·620	·621	·630	·629	·616
20	·603	·609	·611	·620	·621	·629	·628	·616

The orifice was 0·196 sq. in. area and the reductions were made with g = 32·176 the value for Montreal. The value of the coefficient appears to increase as (perimeter) / (area) increases. It decreases as the head increases. It decreases a little as the size of the orifice is greater.

Very careful experiments by J. G. Mair (Proc. Inst. Civ. Eng. lxxxiv.) on the discharge from circular orifices gave the results shown on top of next column.

The edges of the orifices were got up with scrapers to a sharp square edge. The coefficients generally fall as the head increases and as the diameter increases. Professor W. C. Unwin found that the results agree with the formula

c = 0·6075 + 0·0098 / √ h − 0·0037d,

where h is in feet and d in inches.

Coefficients of Discharge from Circular Orifices. Temperature 51° to 55°.

Head in feet h.	Diameters of Orifices in Inches (d).
	1	11/4	11/2	13/4	2	21/4	21/2	23/4	3
	Coefficients (c).
·75	·616	·614	·616	·610	·616	·612	·607	·607	·609
1·0	·613	·612	·612	·611	·612	·611	·604	·608	·609
1·25	·613	·614	·610	·608	·612	·608	·605	·605	·606
1·50	·610	·612	·611	·606	·610	·607	·603	·607	·605
1·75	·612	·611	·611	·605	·611	·605	·604	·607	·605
2·00	·609	·613	·609	·606	·609	·606	·604	·604	·605

The following table, compiled by J. T. Fanning (Treatise on Water Supply Engineering), gives values for rectangular orifices in vertical plane surfaces, the head being measured, not immediately over the orifice, where the surface is depressed, but to the still-water surface at some distance from the orifice. The values were obtained by graphic interpolation, all the most reliable experiments being plotted and curves drawn so as to average the discrepancies.

Coefficients of Discharge for Rectangular Orifices, Sharp-edged, in Vertical Plane Surfaces.

Head to  Centre of  Orifice.	Ratio of Height to Width.
	4	2	11/2	1	3/4	1/2	1/4	1/8
Feet.	4 ft. high. 1 ft. wide.	2 ft. high. 1 ft. wide.	11/2 ft. high. 1 ft. wide.	1 ft. high. 1 ft. wide.	0·75 ft. high. 1 ft. wide.	0·50 ft. high. 1 ft. wide.	0·25 ft. high. 1 ft. wide.	0·125 ft. high.  1 ft. wide.
0·2	..	..	..	..	..	..	..	·6333
·3	..	..	..	..	..	..	·6293	·6334
·4	..	..	..	..	..	·6140	·6306	·6334
·5	..	..	..	..	·6050	·6150	·6313	·6333
·6	..	..	..	·5984	·6063	·6156	·6317	·6332
·7	..	..	..	·5994	·6074	·6162	·6319	·6328
·8	..	..	·6130	·6000	·6082	·6165	·6322	·6326
·9	..	..	·6134	·6006	·6086	·6168	·6323	·6324
1·0	..	..	·6135	·6010	·6090	·6172	·6320	·6320
1·25	..	·6188	·6140	·6018	·6095	·6173	·6317	·6312
1·50	..	·6187	·6144	·6026	·6100	·6172	·6313	·6303
1·75	..	·6186	·6145	·6033	·6103	·6168	·6307	·6296
2	..	·6183	·6144	·6036	·6104	·6166	·6302	·6291
2·25	..	·6180	·6143	·6029	·6103	·6163	·6293	·6286
2·50	·6290	·6176	·6139	·6043	·6102	·6157	·6282	·6278
2·75	·6280	·6173	·6136	·6046	·6101	·6155	·6274	·6273
3	·6273	·6170	·6132	·6048	·6100	·6153	·6267	·6267
3·5	·6250	·6160	·6123	·6050	·6094	·6146	·6254	·6254
4	·6245	·6150	·6110	·6047	·6085	·6136	·6236	·6236
4·5	·6226	·6138	·6100	·6044	·6074	·6125	·6222	·6222
5	·6208	·6124	·6088	·6038	·6063	·6114	·6202	·6202
6	·6158	·6094	·6063	·6020	·6044	·6087	·6154	·6154
7	·6124	·6064	·6038	·6011	·6032	·6058	·6110	·6114
8	·6090	·6036	·6022	·6010	·6022	·6033	·6073	·6087
9	·6060	·6020	·6014	·6010	·6015	·6020	·6045	·6070
10	·6035	·6015	·6010	·6010	·6010	·6010	·6030	·6060
15	·6040	·6018	·6010	·6011	·6012	·6013	·6033	·6066
20	·6045	·6024	·6012	·6012	·6014	·6018	·6036	·6074
25	·6048	·6028	·6014	·6012	·6016	·6022	·6040	·6083
30	·6054	·6034	·6017	·6013	·6018	·6027	·6044	·6092
35	·6060	·6039	·6021	·6014	·6022	·6032	·6049	·6103
40	·6066	·6045	·6025	·6015	·6026	·6037	·6055	·6114
45	·6054	·6052	·6029	·6016	·6030	·6043	·6062	·6125
50	·6086	·6060	·6034	·6018	·6035	·6050	·6070	·6140

§ 21. Orifices with Edges of Sensible Thickness.—When the edges of the orifice are not bevelled outwards, but have a sensible thickness, the coefficient of discharge is somewhat altered. The following table gives values of the coefficient of discharge for the arrangements of the orifice shown in vertical section at P, Q, R (fig. 20). The plan of all the orifices is shown at S. The planks forming the orifice and sluice were each 2 in. thick, and the orifices were all 24 in. wide. The heads were measured immediately over the orifice. In this case,

Q = cb (H − h) √ { 2g(H + h)/2 }.

§ 22. Partially Suppressed Contraction.—Since the contraction of the jet is due to the convergence towards the orifice of the issuing streams, it will be diminished if for any portion of the edge of the orifice the convergence is prevented. Thus, if an internal rim or border is applied to part of the edge of the orifice (fig. 21), the convergence for so much of the edge is suppressed. For such cases

G. Bidone found the following empirical formulae applicable:—

Table of Coefficients of Discharge for Rectangular Vertical Orifices in Fig. 20.

Head h above upper  edge of  Orifice in feet.	Height of Orifice, H − h, in feet.
	1.31			0.66			0.16			0.10
	P	Q	R	P	Q	R	P	Q	R	P	Q	R
0.328	0.598	0.644	0.648	0.634	0.665	0.668	0.691	0.664	0.666	0.710	0.694	0.696
.656	0.609	0.653	0.657	0.640	0.672	0.675	0.685	0.687	0.688	0.696	0.704	0.706
.787	0.612	0.655	0.659	0.641	0.674	0.677	0.684	0.690	0.692	0.694	0.706	0.708
.984	0.616	0.656	0.660	0.641	0.675	0.678	0.683	0.693	0.695	0.692	0.709	0.711
1.968	0.618	0.649	0.653	0.640	0.676	0.679	0.678	0.695	0.697	0.688	0.710	0.712
3.28	0.608	0.632	0.634	0.638	0.674	0.676	0.673	0.694	0.695	0.680	0.704	0.705
4.27	0.602	0.624	0.626	0.637	0.673	0.675	0.672	0.693	0.694	0.678	0.701	0.702
4.92	0.598	0.620	0.622	0.637	0.673	0.674	0.672	0.692	0.693	0.676	0.699	0.699
5.58	0.596	0.618	0.620	0.637	0.672	0.673	0.672	0.692	0.693	0.676	0.698	0.698
6.56	0.595	0.615	0.617	0.636	0.671	0.672	0.671	0.691	0.692	0.675	0.696	0.696
9.84	0.592	0.611	0.612	0.634	0.669	0.670	0.668	0.689	0.690	0.672	0.693	0.693

For rectangular orifices,

Cc = 0.62 (1 + 0.152 n/p);

and for circular orifices,

Cc = 0.62 (1 + 0.128 n/p);

when n is the length of the edge of the orifice over which the border extends, and p is the whole length of edge or perimeter of the orifice. The following are the values of cc, when the border extends over 1/4, 1/2, or 3/4 of the whole perimeter:—

n/p	Cc  Rectangular Orifices	Cc  Circular Orifices
0.25	0.643	.640
0.50	0.667	.660
0.75	0.691	.680

Fig. 20.	Fig. 21.

For larger values of n/p the formulae are not applicable. C. R. Bornemann has shown, however, that these formulae for suppressed contraction are not reliable.

§ 23. Imperfect Contraction.—If the sides of the vessel approach near to the edge of the orifice, they interfere with the convergence of the streams to which the contraction is due, and the contraction is then modified. It is generally stated that the influence of the sides begins to be felt if their distance from the edge of the orifice is less than 2.7 times the corresponding width of the orifice. The coefficients of contraction for this case are imperfectly known.

Fig. 22.

§ 24. Orifices Furnished with Channels of Discharge.—These external borders to an orifice also modify the contraction.

The following coefficients of discharge were obtained with openings 8 in. wide, and small in proportion to the channel of approach (fig. 22, A, B, C).

h2 − h1 in feet		h1 in feet.
		.0656	.164	.328	.656	1.640	3.28	4.92	6.56	9.84
A	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \ \end{matrix}}\right\}\,}$0.656	.480	.511	.542	.574	.599	.601	.601	.601	.601
B		.480	.510	.538	.506	.592	.600	.602	.602	.601
C		.527	.553	.574	.592	.607	.610	.610	.609	.608
A	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \ \end{matrix}}\right\}\,}$0.164	.488	.577	.624	.631	.625	.624	.619	.613	.606
B		.487	.571	.606	.617	.626	.628	.627	.623	.618
C		.585	.614	.633	.645	.652	.651	.650	.650	.649

Fig. 23.

§ 25. Inversion of the Jet.—When a jet issues from a horizontal orifice, or is of small size compared with the head, it presents no marked peculiarity of form. But if the orifice is in a vertical surface, and if its dimensions are not small compared with the head, it undergoes a series of singular changes of form after leaving the orifice. These were first investigated by G. Bidone (1781–1839); subsequently H. G. Magnus (1802–1870) measured jets from different orifices; and later Lord Rayleigh (Proc. Roy. Soc. xxix. 71) investigated them anew.

Fig. 23 shows some forms, the upper figure giving the shape of the orifices, and the others sections of the jet. The jet first contracts as described above, in consequence of the convergence of the fluid streams within the vessel, retaining, however, a form similar to that of the orifice. Afterwards it expands into sheets in planes perpendicular to the sides of the orifice. Thus the jet from a triangular orifice expands into three sheets, in planes bisecting at right angles the three sides of the triangle. Generally a jet from an orifice, in the form of a regular polygon of n sides, forms n sheets in planes perpendicular to the sides of the polygon.

Bidone explains this by reference to the simpler case of meeting streams. If two equal streams having the same axis, but moving in opposite directions, meet, they spread out into a thin disk normal to the common axis of the streams. If the directions of two streams intersect obliquely they spread into a symmetrical sheet perpendicular to the plane of the streams.

Fig. 24.

Let a₁, a₂ (fig. 24) be two points in an orifice at depths h₁, h₂ from the free surface. The filaments issuing at a₁, a₂ will have the different velocities √ 2gh₁ and √ 2gh₂. Consequently they will tend to describe parabolic paths a₁cb₁ and a₂cb₂ of different horizontal range, and intersecting in the point c. But since two filaments cannot simultaneously flow through the same point, they must exercise mutual pressure, and will be deflected out of the paths they tend to describe. It is this mutual pressure which causes the expansion of the jet into sheets.

Lord Rayleigh pointed out that, when the orifices are small and the head is not great, the expansion of the sheets in directions perpendicular to the direction of flow reaches a limit. Sections taken at greater distance from the orifice show a contraction of the sheets until a compact form is reached similar to that at the first contraction. Beyond this point, if the jet retains its coherence, sheets are thrown out again, but in directions bisecting the angles between the previous sheets. Lord Rayleigh accepts an explanation of this contraction first suggested by H. Buff (1805–1878), namely, that it is due to surface tension.

§ 26. Influence of Temperature on Discharge of Orifices.—Professor W. C. Unwin found (Phil. Mag., October 1878, p. 281) that for sharp-edged orifices temperature has a very small influence on the discharge. For an orifice 1 cm. in diameter with heads of about 1 to 11/2 ft. the coefficients were:—

Temperature F.	C.
205°	.594
62°	.598

For a conoidal or bell-mouthed orifice 1 cm. diameter the effect of temperature was greater:—

Temperature F.	C.
190°	0.987
130°	0.974
60°	0.942

an increase in velocity of discharge of 4% when the temperature increased 130°.

J. G. Mair repeated these experiments on a much larger scale (Proc. Inst. Civ. Eng. lxxxiv.). For a sharp-edged orifice 21/2 in. diameter, with a head of 1.75 ft., the coefficient was 0.604 at 57° and 0.607 at 179° F., a very small difference. With a conoidal orifice the coefficient was 0.961 at 55° and 0.981 at 170° F. The corresponding coefficients of resistance are 0.0828 and 0.0391, showing that the resistance decreases to about half at the higher temperature.

§ 27. Fire Hose Nozzles.—Experiments have been made by J. R. Freeman on the coefficient of discharge from smooth cone nozzles used for fire purposes. The coefficient was found to be 0.983 for 3/4-in. nozzle; 0.982 for 7/8 in.; 0.972 for 1 in.; 0.976 for 11/8 in.; and 0.971 for 11/4 in. The nozzles were fixed on a taper play-pipe, and the coefficient includes the resistance of this pipe (Amer. Soc. Civ. Eng. xxi., 1889). Other forms of nozzle were tried such as ring nozzles for which the coefficient was smaller.

IV. THEORY OF THE STEADY MOTION OF FLUIDS.

§ 28. The general equation of the steady motion of a fluid given under Hydrodynamics furnishes immediately three results as to the distribution of pressure in a stream which may here be assumed.

(a) If the motion is rectilinear and uniform, the variation of pressure is the same as in a fluid at rest. In a stream flowing in an open channel, for instance, when the effect of eddies produced by the roughness of the sides is neglected, the pressure at each point is simply the hydrostatic pressure due to the depth below the free surface.

(b) If the velocity of the fluid is very small, the distribution of pressure is approximately the same as in a fluid at rest.

(c) If the fluid molecules take precisely the accelerations which they would have if independent and submitted only to the external forces, the pressure is uniform. Thus in a jet falling freely in the air the pressure throughout any cross section is uniform and equal to the atmospheric pressure.

(d) In any bounded plane section traversed normally by streams which are rectilinear for a certain distance on either side of the section, the distribution of pressure is the same as in a fluid at rest.

Distribution of Energy in Incompressible Fluids.

§ 29. Application of the Principle of the Conservation of Energy to Cases of Stream Line Motion.—The external and internal work done on a mass is equal to the change of kinetic energy produced. In many hydraulic questions this principle is difficult to apply, because from the complicated nature of the motion produced it is difficult to estimate the total kinetic energy generated, and because in some cases the internal work done in overcoming frictional or viscous resistances cannot be ascertained; but in the case of stream line motion it furnishes a simple and important result known as Bernoulli’s theorem.

Fig. 25.

Let AB (fig. 25) be any one elementary stream, in a steadily moving fluid mass. Then, from the steadiness of the motion, AB is a fixed path in space through which a stream of fluid is constantly flowing. Let OO be the free surface and XX any horizontal datum line. Let ω be the area of a normal cross section, v the velocity, p the intensity of pressure, and z the elevation above XX, of the elementary stream AB at A, and ω₁, p₁, v₁, z₁ the same quantities at B. Suppose that in a short time t the mass of fluid initially occupying AB comes to A′B′. Then AA′, BB′ are equal to vt, v₁t, and the volumes of fluid AA′, BB′ are the equal inflow and outflow = Qt = ωvt = ω₁v₁t, in the given time. If we suppose the filament AB surrounded by other filaments moving with not very different velocities, the frictional or viscous resistance on its surface will be small enough to be neglected, and if the fluid is incompressible no internal work is done in change of volume. Then the work done by external forces will be equal to the kinetic energy produced in the time considered.

The normal pressures on the surface of the mass (excluding the ends A, B) are at each point normal to the direction of motion, and do no work. Hence the only external forces to be reckoned are gravity and the pressures on the ends of the stream.

The work of gravity when AB falls to A′B′ is the same as that of transferring AA′ to BB′; that is, GQt (z − z₁). The work of the pressures on the ends, reckoning that at B negative, because it is opposite to the direction of motion, is (pω × vt) − (p₁ω₁ × v₁t) = Qt(p − p₁). The change of kinetic energy in the time t is the difference of the kinetic energy originally possessed by AA′ and that finally acquired by BB′, for in the intermediate part A′B there is no change of kinetic energy, in consequence of the steadiness of the motion. But the mass of AA′ and BB′ is GQt/g, and the change of kinetic energy is therefore (GQt/g) (v₁²/2 − v²/2). Equating this to the work done on the mass AB,

GQt (z − z₁) + Qt (p − p₁) = (GQt/g) (v₁²/2 − v²/2).

Dividing by GQt and rearranging the terms,

v²/2g + p/G + z = v₁²/2g + p₁/G + z₁;

(1)

or, as A and B are any two points,

v²/2g + p/G + z = constant = H.

(2)

Now v²/2g is the head due to the velocity v, p/G is the head equivalent to the pressure, and z is the elevation above the datum (see § 16). Hence the terms on the left are the total head due to velocity, pressure, and elevation at a given cross section of the filament, z is easily seen to be the work in foot-pounds which would be done by 1 ℔ of fluid falling to the datum line, and similarly p/G and v²/2g are the quantities of work which would be done by 1 ℔ of fluid due to the pressure p and velocity v. The expression on the left of the equation is, therefore, the total energy of the stream at the section considered, per ℔ of fluid, estimated with reference to the datum line XX. Hence we see that in stream line motion, under the restrictions named above, the total energy per ℔ of fluid is uniformly distributed along the stream line. If the free surface of the fluid OO is taken as the datum, and −h, −h₁ are the depths of A and B measured down from the free surface, the equation takes the form

v²/2g + p/G − h = v₁²/2g + p₁/G − h₁;

(3)

or generally

v²/2g + p/G − h = constant.

(3a)

Fig. 26.

§ 30. Second Form of the Theorem of Bernoulli.—Suppose at the two sections A, B (fig. 26) of an elementary stream small vertical pipes are introduced, which may be termed pressure columns (§ 8), having their lower ends accurately parallel to the direction of flow. In such tubes the water will rise to heights corresponding to the pressures at A and B. Hence b = p/G, and b′ = p₁/G. Consequently the tops of the pressure columns A′ and B′ will be at total heights b + c = p/G + z and b′ + c′ = p₁/G + z₁ above the datum line XX. The difference of level of the pressure column tops, or the fall of free surface level between A and B, is therefore

ξ = (p − p₁) / G + (z − z₁);

and this by equation (1), § 29 is (v₁² − v²)/2g. That is, the fall of free, surface level between two sections is equal to the difference of the heights due to the velocities at the sections. The line A′B′ is sometimes called the line of hydraulic gradient, though this term is also used in cases where friction needs to be taken into account. It is the line the height of which above datum is the sum of the elevation and pressure head at that point, and it falls below a horizontal line A″B″ drawn at H ft. above XX by the quantities a = v²/2g and a′ = v₁²/2g, when friction is absent.

§ 31. Illustrations of the Theorem of Bernoulli. In a lecture to the mechanical section of the British Association in 1875, W. Froude gave some experimental illustrations of the principle of Bernoulli. He remarked that it was a common but erroneous impression that a fluid exercises in a contracting pipe A (fig. 27) an excess of pressure against the entire converging surface which it meets, and that, conversely, as it enters an enlargement B, a relief of pressure is experienced by the entire diverging surface of the pipe. Further it is commonly assumed that when passing through a contraction C, there is in the narrow neck an excess of pressure due to the squeezing together of the liquid at that point. These impressions are in no respect correct; the pressure is smaller as the section of the pipe is smaller and conversely.

Fig. 27.

Fig. 28 shows a pipe so formed that a contraction is followed by an enlargement, and fig. 29 one in which an enlargement is followed by a contraction. The vertical pressure columns show the decrease of pressure at the contraction and increase of pressure at the enlargement. The line abc in both figures shows the variation of free surface level, supposing the pipe frictionless. In actual pipes, however, work is expended in friction against the pipe; the total head diminishes in proceeding along the pipe, and the free surface level is a line such as ab₁c₁, falling below abc.

Froude further pointed out that, if a pipe contracts and enlarges again to the same size, the resultant pressure on the converging part exactly balances the resultant pressure on the diverging part so that there is no tendency to move the pipe bodily when water flows through it. Thus the conical part AB (fig. 30) presents the same projected surface as HI, and the pressures parallel to the axis of the pipe, normal to these projected surfaces, balance each other. Similarly the pressures on BC, CD balance those on GH, EG. In the same way, in any combination of enlargements and contractions, a balance of pressures, due to the flow of liquid parallel to the axis of the pipe, will be found, provided the sectional area and direction of the ends are the same.

Fig. 28.

Fig. 29.

The following experiment is interesting. Two cisterns provided with converging pipes were placed so that the jet from one was exactly opposite the entrance to the other. The cisterns being filled very nearly to the same level, the jet from the left-hand cistern A entered the right-hand cistern B (fig. 31), shooting across the free space between them without any waste, except that due to indirectness of aim and want of exact correspondence in the form of the orifices. In the actual experiment there was 18 in. of head in the right and 201/2 in. of head in the left-hand cistern, so that about 21/2 in. were wasted in friction. It will be seen that in the open space between the orifices there was no pressure, except the atmospheric pressure acting uniformly throughout the system.

Fig. 30.

Fig. 31.

§ 32. Venturi Meter.—An ingenious application of the variation of pressure and velocity in a converging and diverging pipe has been made by Clemens Herschel in the construction of what he terms a Venturi Meter for measuring the flow in water mains. Suppose that, as in fig. 32, a contraction is made in a water main, the change of section being gradual to avoid the production of eddies. The ratio ρ of the cross sections at A and B, that is at inlet and throat, is in actual meters 5 to 1 to 20 to 1, and is very carefully determined by the maker of the meter. Then, if v and u are the velocities at A and B, u = ρv. Let pressure pipes be introduced at A, B and C, and let H₁, H, H₂ be the pressure heads at those points. Since the velocity at B is greater than at A the pressure will be less. Neglecting friction

H₁ + v²/2g = H + u²/2g,

H₁ − H = (u² − v²) / 2g = (ρ² − 1) v² 2g.

Let h = H₁ − H be termed the Venturi head, then

u = √ { ρ².2gh / (ρ² − 1) },

from which the velocity through the throat and the discharge of the main can be calculated if the areas at A and B are known and h observed. Thus if the diameters at A and B are 4 and 12 in., the areas are 12.57 and 113.1 sq. in., and ρ = 9,

u = √ 81/80 √ (2gh) = 1.007 √ (2gh).

If the observed Venturi head is 12 ft.,

u = 28 ft. per sec.,

and the discharge of the main is

28 × 12.57 = 351 cub. ft. per sec.

Fig. 32.

Fig. 33.

Hence by a simple observation of pressure difference, the flow in the main at any moment can be determined. Notice that the pressure height at C will be the same as at A except for a small loss h_f due to friction and eddying between A and B. To get the pressure at the throat very exactly Herschel surrounds it by an annular passage communicating with the throat by several small holes, sometimes formed in vulcanite to prevent corrosion. Though constructed to prevent eddying as much as possible there is some eddy loss. The main effect of this is to cause a loss of head between A and C which may vary from a fraction of a foot to perhaps 5 ft. at the highest velocities at which a meter can be used. The eddying also affects a little the Venturi head h. Consequently an experimental coefficient must be determined for each meter by tank measurement. The range of this coefficient is, however, surprisingly small. If to allow for friction, u = k √ {ρ²/(ρ² − 1)} √(2gh), then Herschel found values of k from 0.97 to 1.0 for throat velocities varying from 8 to 28 ft. per sec. The meter is extremely convenient. At Staines reservoirs there are two meters of this type on mains 94 in. in diameter. Herschel contrived a recording arrangement which records the variation of flow from hour to hour and also the total flow in any given time. In Great Britain the meter is constructed by G. Kent, who has made improvements in the recording arrangement.

In the Deacon Waste Water Meter (fig. 33) a different principle is used. A disk D, partly counter-balanced by a weight, is suspended in the water flowing through the main in a conical chamber. The unbalanced weight of the disk is supported by the impact of the water. If the discharge of the main increases the disk rises, but as it rises its position in the chamber is such that in consequence of the larger area the velocity is less. It finds, therefore, a new position of equilibrium. A pencil P records on a drum moved by clockwork the position of the disk, and from this the variation of flow is inferred.

§ 33. Pressure, Velocity and Energy in Different Stream Lines.—The equation of Bernoulli gives the variation of pressure and velocity from point to point along a stream line, and shows that the total energy of the flow across any two sections is the same. Two other directions may be defined, one normal to the stream line and in the plane containing its radius of curvature at any point, the other normal to the stream line and the radius of curvature. For the problems most practically useful it will be sufficient to consider the stream lines as parallel to a vertical or horizontal plane. If the motion is in a vertical plane, the action of gravity must be taken into the reckoning; if the motion is in a horizontal plane, the terms expressing variation of elevation of the filament will disappear.^[3]

Fig. 34.

Let AB, CD (fig. 34) be two consecutive stream lines, at present assumed to be in a vertical plane, and PQ a normal to these lines making an angle φ with the vertical. Let P, Q be two particles moving along these lines at a distance PQ = ds, and let z be the height of Q above the horizontal plane with reference to which the energy is measured, v its velocity, and p its pressure. Then, if H is the total energy at Q per unit of weight of fluid,

H = z + p/G + v²/2g.

Differentiating, we get

dH = dz + dp/G + v dv/g,

(1)

for the increment of energy between Q and P. But

dz = PQ cos φ = ds cos φ;

∴ dH = dp/G + v dv/g + ds cos φ,

(1a)

where the last term disappears if the motion is in a horizontal plane.

Now imagine a small cylinder of section ω described round PQ as an axis. This will be in equilibrium under the action of its centrifugal force, its weight and the pressure on its ends. But its volume is ωds and its weight Gω ds. Hence, taking the components of the forces parallel to PQ—

ωdp = Gv²ω ds/gρ − Gω cos φds,

where ρ is the radius of curvature of the stream line at Q. Consequently, introducing these values in (1),

dH = v²ds/gρ + vdv/g = (v/g) (v/ρ + dv/ds) ds.

(2)

Currents

§ 34. Rectilinear Current.—Suppose the motion is in parallel straight stream lines (fig. 35) in a vertical plane. Then ρ is infinite, and from eq. (2), § 33,

dH = vdv/g.

Comparing this with (1) we see that

dz + dp/G = 0;

∴ z + p/G = constant;

(3)

Fig. 35.

or the pressure varies hydrostatically as in a fluid at rest. For two stream lines in a horizontal plane, z is constant, and therefore p is constant.

Radiating Current.—Suppose water flowing radially between horizontal parallel planes, at a distance apart = δ. Conceive two cylindrical sections of the current at radii r₁ and r₂, where the velocities are v₁ and v₂, and the pressures p₁ and p₂. Since the flow across each cylindrical section of the current is the same,

Q = 2πr₁ δv₁ = 2πr₂ δv₂

r₁v₁ = r₂v₂

 r₁/r₂ = v₂/v₁.

(4)

The velocity would be infinite at radius 0, if the current could be conceived to extend to the axis. Now, if the motion is steady,

H = p₁/G + v₁²/2g = p₂/G + v₂²/2g;

= p₂/G + r₁² + v₁²/r₂²2g;

(p₂ − p₁)/G = v₁² (1 − r₁²/r₂²)/2g;

(5)

p₂/G = H − r₁²v₁²/r₂²2g.

(6)

Hence the pressure increases from the interior outwards, in a way indicated by the pressure columns in fig. 36, the curve through the free surfaces of the pressure columns being, in a radial section, the quasi-hyperbola of the form xy² = c³. This curve is asymptotic to a horizontal line, H ft. above the line from which the pressures are measured, and to the axis of the current.

Fig. 36.

Free Circular Vortex.—A free circular vortex is a revolving mass of water, in which the stream lines are concentric circles, and in which the total head for each stream line is the same. Hence, if by any slow radial motion portions of the water strayed from one stream line to another, they would take freely the velocities proper to their new positions under the action of the existing fluid pressures only.

For such a current, the motion being horizontal, we have for all the circular elementary streams

H = p/G + v²/2g = constant;

∴ dH = dp/G + vdv/g = 0.

(7)

Consider two stream lines at radii r and r + dr (fig. 36). Then in (2), § 33, ρ = r and ds = dr,

v² dr/gr + vdv/g = 0,

dv/v = −dr/r,

v ∞ 1/r,

(8)

precisely as in a radiating current; and hence the distribution of pressure is the same, and formulae 5 and 6 are applicable to this case.

Free Spiral Vortex.—As in a radiating and circular current the equations of motion are the same, they will also apply to a vortex in which the motion is compounded of these motions in any proportions, provided the radial component of the motion varies inversely as the radius as in a radial current, and the tangential component varies inversely as the radius as in a free vortex. Then the whole velocity at any point will be inversely proportional to the radius of the point, and the fluid will describe stream lines having a constant inclination to the radius drawn to the axis of the current. That is, the stream lines will be logarithmic spirals. When water is delivered from the circumference of a centrifugal pump or turbine into a chamber, it forms a free vortex of this kind. The water flows spirally outwards, its velocity diminishing and its pressure increasing according to the law stated above, and the head along each spiral stream line is constant.

§ 35. Forced Vortex.—If the law of motion in a rotating current is different from that in a free vortex, some force must be applied to cause the variation of velocity. The simplest case is that of a rotating current in which all the particles have equal angular velocity, as for instance when they are driven round by radiating paddles revolving uniformly. Then in equation (2), § 33, considering two circular stream lines of radii r and r + dr (fig. 37), we have ρ = r, ds = dr. If the angular velocity is α, then v = αr and dv = αdr. Hence

dH = α²rdr/g + α²rdr/g = 2α²rdr/g.

Comparing this with (1), § 33, and putting dz = 0, because the motion is horizontal,

dp/G + α²rdr/g = 2α²rdr/g,

dp/G = α²rdr/g,

p/G = α²/2g + constant.

(9)

Let p₁, r₁, v₁ be the pressure, radius and velocity of one cylindrical section, p₂, r₂, v₂ those of another; then

p₁/G − α²r₁²/2g = p₂/G − α²r₂²/2g;

(p₂ − p₁)/G = α² (r₂² − r₁²)/2g = (v₂² − v₁²)/2g.

(10)

That is, the pressure increases from within outwards in a curve which in radial sections is a parabola, and surfaces of equal pressure are paraboloids of revolution (fig. 37).

Fig. 37.

Dissipation of Head in Shock

§ 36. Relation of Pressure and Velocity in a Stream in Steady Motion when the Changes of Section of the Stream are Abrupt.—When a stream changes section abruptly, rotating eddies are formed which dissipate energy. The energy absorbed in producing rotation is at once abstracted from that effective in causing the flow, and sooner or later it is wasted by frictional resistances due to the rapid relative motion of the eddying parts of the fluid. In such cases the work thus expended internally in the fluid is too important to be neglected, and the energy thus lost is commonly termed energy lost in shock. Suppose fig. 38 to represent a stream having such an abrupt change of section. Let AB, CD be normal sections at points where ordinary stream line motion has not been disturbed and where it has been re-established. Let ω, p, v be the area of section, pressure and velocity at AB, and ω₁, p₁, v₁ corresponding quantities at CD. Then if no work were expended internally, and assuming the stream horizontal, we should have

p/G + v²/2g = p₁/G + v₁²/2g.

(1)

But if work is expended in producing irregular eddying motion, the

head at the section CD will be diminished.

Suppose the mass ABCD comes in a short time t to A′B′C′D′. The resultant force parallel to the axis of the stream is

pω + p₀ (ω₁ − ω) − p₁ω₁,

where p₀ is put for the unknown pressure on the annular space between AB and EF. The impulse of that force is

{ pω + p₀ (ω₁ − ω) − p₁ω₁ } t.

Fig. 38.

The horizontal change of momentum in the same time is the difference of the momenta of CDC′D′ and ABA′B′, because the amount of momentum between A′B′ and CD remains unchanged if the motion is steady. The volume of ABA′B′ or CDC′D′, being the inflow and outflow in the time t, is Qt = ωvt = ω₁v₁t, and the momentum of these masses is (G/g) Qvt and (G/g) Qv₁t. The change of momentum is therefore (G/g) Qt (v₁ − v). Equating this to the impulse,

{ pω + p₀ (ω₁ − ω) − p₁ω₁ } t = (G/g) Qt (v₁ − v).

Assume that p₀ = p, the pressure at AB extending unchanged through the portions of fluid in contact with AE, BF which lie out of the path of the stream. Then (since Q = ω₁v₁)

(p − p₁) = (G/g) v₁ (v₁ − v);

p/G − p₁/G = v₁ (v₁ − v)/g;

(2)

p/G + v²/2g = p₁/G + v₁²/2g + (v − v₁)²/2g.

(3)

This differs from the expression (1), § 29, obtained for cases where no sensible internal work is done, by the last term on the right. That is, (v − v₁)²/2g has to be added to the total head at CD, which is p₁/G + v₁²/2g, to make it equal to the total head at AB, or (v − v₁)²/2g is the head lost in shock at the abrupt change of section. But (v − v₁) is the relative velocity of the two parts of the stream. Hence, when an abrupt change of section occurs, the head due to the relative velocity is lost in shock, or (v − v₁)²/2g foot-pounds of energy is wasted for each pound of fluid. Experiment verifies this result, so that the assumption that p₀ = p appears to be admissible.

If there is no shock,

p₁/G = p/G + (v² − v₁²)/2g.

If there is shock,

p₁/G = p/G − v₁ (v₁ − v)/g.

Hence the pressure head at CD in the second case is less than in the former by the quantity (v − v₁)²/2g, or, putting ω₁v₁ = ωv, by the quantity

(v²/2g) (1 − ω/ω₁)².

(4)

V. THEORY OF THE DISCHARGE FROM ORIFICES AND MOUTHPIECES

Fig. 39.

§ 37. Minimum Coefficient of Contraction. Re-entrant Mouthpiece of Borda.—In one special case the coefficient of contraction can be determined theoretically, and, as it is the case where the convergence of the streams approaching the orifice takes place through the greatest possible angle, the coefficient thus determined is the minimum coefficient.

Let fig. 39 represent a vessel with vertical sides, OO being the free water surface, at which the pressure is p_a. Suppose the liquid issues by a horizontal mouthpiece, which is re-entrant and of the greatest length which permits the jet to spring clear from the inner end of the orifice, without adhering to its sides. With such an orifice the velocity near the points CD is negligible, and the pressure at those points may be taken equal to the hydrostatic pressure due to the depth from the free surface. Let Ω be the area of the mouthpiece AB, ω that of the contracted jet aa Suppose that in a short time t, the mass OOaa comes to the position O′O′ a′a′; the impulse of the horizontal external forces acting on the mass during that time is equal to the horizontal change of momentum.

The pressure on the side OC of the mass will be balanced by the pressure on the opposite side OE, and so for all other portions of the vertical surfaces of the mass, excepting the portion EF opposite the mouthpiece and the surface AaaB of the jet. On EF the pressure is simply the hydrostatic pressure due to the depth, that is, (p_a + Gh). On the surface and section AaaB of the jet, the horizontal resultant of the pressure is equal to the atmospheric pressure p_a acting on the vertical projection AB of the jet; that is, the resultant pressure is −p_aΩ. Hence the resultant horizontal force for the whole mass OOaa is (p_a + Gh) Ω − p_aΩ = GhΩ. Its impulse in the time t is GhΩt. Since the motion is steady there is no change of momentum between O′O′ and aa. The change of horizontal momentum is, therefore, the difference of the horizontal momentum lost in the space OOO′O′ and gained in the space aaa′a′. In the former space there is no horizontal momentum.

The volume of the space aaa′a′ is ωvt; the mass of liquid in that space is (G/g)ωvt; its momentum is (G/g)ωv²t. Equating impulse to momentum gained,

GhΩt = (G/g) ωv²t;

∴ ω/Ω = gh/v²

But

v² = 2gh, and ω/Ω = c_c;

∴ ω/Ω = 1/2 = c_c;

a result confirmed by experiment with mouthpieces of this kind. A similar theoretical investigation is not possible for orifices in plane surfaces, because the velocity along the sides of the vessel in the neighbourhood of the orifice is not so small that it can be neglected. The resultant horizontal pressure is therefore greater than GhΩ, and the contraction is less. The experimental values of the coefficient of discharge for a re-entrant mouthpiece are 0.5149 (Borda), 0.5547 (Bidone), 0.5324 (Weisbach), values which differ little from the theoretical value, 0.5, given above.

Fig. 40.	Fig. 41.

§ 38. Velocity of Filaments issuing in a Jet.—A jet is composed of fluid filaments or elementary streams, which start into motion at some point in the interior of the vessel from which the fluid is discharged, and gradually acquire the velocity of the jet. Let Mm, fig. 40 be such a filament, the point M being taken where the velocity is insensibly small, and m at the most contracted section of the jet, where the filaments have become parallel and exercise uniform mutual pressure. Take the free surface AB for datum line, and let p₁, v₁, h₁, be the pressure, velocity and depth below datum at M; p, v, h, the corresponding quantities at m. Then § 29, eq. (3a),

v₁²/2g + p₁/G − h₁ = v²/2g + p/G − h

(1)

But at M, since the velocity is insensible, the pressure is the hydrostatic pressure due to the depth; that is v₁ = 0, p₁ = p_a + Gh₁. At m, p = p_a, the atmospheric pressure round the jet. Hence, inserting these values,

0 + p_a/G + h₁ − h₁ = v²/2g + p_a/G − h;

v2/2g = h;

(2)

or

v = √ (2gh) = 8.025V √ h

(2a)

That is, neglecting the viscosity of the fluid, the velocity of filaments at the contracted section of the jet is simply the velocity due to the difference of level of the free surface in the reservoir and the orifice. If the orifice is small in dimensions compared with h, the filaments will all have nearly the same velocity, and if h is measured to the centre of the orifice, the equation above gives the mean velocity of the jet.

Case of a Submerged Orifice.—Let the orifice discharge below the level of the tail water. Then using the notation shown in fig. 41, we have at M, v₁ = 0, p₁ = Gh; + p_a at m, p = Gh₃ + p_a. Inserting these values in (3), § 29,

0 + h₁ + p_a/G − h₁ = v²/2g + h₃ − h₂2 + p_a/G;

v²/2g = h₂ − h₃ = h,

(3)

where h is the difference of level of the head and tail water, and may be termed the effective head producing flow.

Fig. 42.

Case where the Pressures are different on the Free Surface and at the Orifice.—Let the fluid flow from a vessel in which the pressure is p₀ into a vessel in which the pressure is p, fig. 42. The pressure p₀ will produce the same effect as a layer of fluid of thickness p₀/G added to the head water; and the pressure p, will produce the same effect as a layer of thickness p/G added to the tail water. Hence the effective difference of level, or effective head producing flow, will be

h = h₀ + p₀/G − p/G;

and the velocity of discharge will be

v = √ [ 2g { h₀ + (p₀ − p) / G } ].

(4)

We may express this result by saying that differences of pressure at the free surface and at the orifice are to be reckoned as part of the effective head.

Hence in all cases thus far treated the velocity of the jet is the velocity due to the effective head, and the discharge, allowing for contraction of the jet, is

Q = cωv = cω√ (2gh),

(5)

where ω is the area of the orifice, cω the area of the contracted section of the jet, and h the effective head measured to the centre of the orifice. If h and ω are taken in feet, Q is in cubic feet per second.

It is obvious, however, that this formula assumes that all the filaments have sensibly the same velocity. That will be true for horizontal orifices, and very approximately true in other cases, if the dimensions of the orifice are not large compared with the head h. In large orifices in say a vertical surface, the value of h is different for different filaments, and then the velocity of different filaments is not sensibly the same.

Simple Orifices—Head Constant

Fig. 43.

§ 39. Large Rectangular Jets from Orifices in Vertical Plane Surfaces.—Let an orifice in a vertical plane surface be so formed that it produces a jet having a rectangular contracted section with vertical and horizontal sides. Let b (fig. 43) be the breadth of the jet, h₁ and h₂ the depths below the free surface of its upper and lower surfaces. Consider a lamina of the jet between the depths h and h + dh. Its normal section is bdh, and the velocity of discharge √2gh. The discharge per second in this lamina is therefore b√2gh dh, and that of the whole jet is therefore

${\mbox{Q}}=\int _{h_{1}}^{h_{2}}b{\sqrt {(}}2gh)dh$

= 2/3 b √2g { h₂^3/2 − h₁^3/2 },

(6)

where the first factor on the right is a coefficient depending on the form of the orifice.

Now an orifice producing a rectangular jet must itself be very approximately rectangular. Let B be the breadth, H₁, H₂, the depths to the upper and lower edges of the orifice. Put

b (h ₂^3/2 − h₁^3/2) / B (H₂^3/2 − H₁^3/2) = c.

(7)

Then the discharge, in terms of the dimensions of the orifice, instead of those of the jet, is

Q = 2/3 cB √2g (H₂^3/2 − H₁^3/2),

(8)

the formula commonly given for the discharge of rectangular orifices. The coefficient c is not, however, simply the coefficient of contraction, the value of which is

b (h ₂ − h₁) / B (H₂ − H₁),

and not that given in (7). It cannot be assumed, therefore, that c in equation (8) is constant, and in fact it is found to vary for different values of B/H₂ and B/H₁, and must be ascertained experimentally.

Relation between the Expressions (5) and (8).—For a rectangular orifice the area of the orifice is ω = B(H₂ − H₁), and the depth measured to its centre is 1/2 (H₂ + H₁). Putting these values in (5),

Q₁ = cB (H₂ − H₁) √ {g (H₂ + H₁) }.

From (8) the discharge is

Q₂ = 2/3 cB √2g (H₂^3/2 − H₁^3/2).

Hence, for the same value of c in the two cases,

Q₂/Q₁ = 2/3 (H₂^3/2 − H₁^3/2) / [ (H₂ − H₁) √ { (H₂ + H₁)/2} ].

Let H₁/H₂ = σ, then

Q₂/Q₁ = 0.9427 (1 − σ^3/2) / {1 − σ √ (1 + σ) }.

(9)

If H₁ varies from 0 to ∞, σ( = H₁/H₂) varies from 0 to 1. The following table gives values of the two estimates of the discharge for different values of σ:—

H1/H2 = σ.	Q2/Q1.	H1/H2 = σ.	Q2/Q1.
0.0	.943	0.8	.999
0.2	.979	0.9	.999
0.5	.995	1.0	1.000
0.7	.998

Hence it is obvious that, except for very small values of σ, the simpler equation (5) gives values sensibly identical with those of (8). When σ < 0.5 it is better to use equation (8) with values of c determined experimentally for the particular proportions of orifice which are in question.

Fig. 44.

§ 40. Large Jets having a Circular Section from Orifices in a Vertical Plane Surface.—Let fig. 44 represent the section of the jet, OO being the free surface level in the reservoir. The discharge through the horizontal strip aabb, of breadth aa = b, between the depths h₁ + y and h₁ + y + dy, is

dQ = b √ {2g (h₁ + y) } dy.

The whole discharge of the jet is

Q = $\int _{0}^{d}$ b √ { 2g (h ₁ + y) } dy.

But b = d sin φ; y = 1/2d (1 − cos φ); dy = 1/2d sin φ dφ. Let ε = d/(2h₁ + d), then

Q = 1/2d ² √ { 2g (h ₁ + d/2) } $\int _{0}^{\pi }\sin ^{2}\phi {\sqrt {1-\epsilon \cos \phi }}\ d\phi .$

From eq. (5), putting ω = πd ²/4, h = h₁ + d/2, c = 1 when d is the diameter of the jet and not that of the orifice,

Q₁ = 1/4πd ² √ {2g(h₁ + d/2) },

Q/Q₁ = 2/π $\int _{0}^{\pi }$ sin² φ √ {1 − ε cos φ} dφ.

For

h₁ = ∞, ε = 0 and Q/Q₁ = 1;

and for

h₁ = 0, ε = 1 and Q/Q₁ = 0.96.

So that in this case also the difference between the simple formula (5) and the formula above, in which the variation of head at different parts of the orifice is taken into account, is very small.

Notches and Weirs

§ 41. Notches, Weirs and Byewashes.—A notch is an orifice extending up to the free surface level in the reservoir from which the discharge takes place. A weir is a structure over which the water flows, the discharge being in the same conditions as for a notch. The formula of discharge for an orifice of this kind is ordinarily deduced by putting H₁ = 0 in the formula for the corresponding orifice, obtained as in the preceding section. Thus for a rectangular notch, put H₁ = 0 in (8). Then

Q = 2/3 cB √(2g) H^3/2,

(11)

where H is put for the depth to the crest of the weir or the bottom of the notch. Fig. 45 shows the mode in which the discharge occurs in the case of a rectangular notch or weir with a level crest. As, the free surface level falls very sensibly near the notch, the head H should be measured at some distance back from the notch, at a point where the velocity of the water is very small.

Since the area of the notch opening is BH, the above formula is of the form

Q = c × BH × k √(2gH),

where k is a factor depending on the form of the notch and expressing the ratio of the mean velocity of discharge to the velocity due to the depth H.

§ 42. Francis’s Formula for Rectangular Notches.—The jet discharged through a rectangular notch has a section smaller than BH, (a) because of the fall of the water surface from the point where H is measured towards the weir, (b) in consequence of the crest contraction, (c) in consequence of the end contractions. It may be pointed out that while the diminution of the section of the jet due to the surface fall and to the crest contraction is proportional to the length of the weir, the end contractions have nearly the same effect whether the weir is wide or narrow.

Fig. 45.

J. B. Francis’s experiments showed that a perfect end contraction, when the heads varied from 3 to 24 in., and the length of the weir was not less than three times the head, diminished the effective length of the weir by an amount approximately equal to one-tenth of the head. Hence, if l is the length of the notch or weir, and H the head measured behind the weir where the water is nearly still, then the width of the jet passing through the notch would be l − 0.2H, allowing for two end contractions. In a weir divided by posts there may be more than two end contractions. Hence, generally, the width of the jet is l − 0.1nH, where n is the number of end contractions of the stream. The contractions due to the fall of surface and to the crest contraction are proportional to the width of the jet. Hence, if cH is the thickness of the stream over the weir, measured at the contracted section, the section of the jet will be c(l − 0.1nH)H and (§ 41) the mean velocity will be 2/3 √(2gH). Consequently the discharge will be given by an equation of the form

Q = 2/3 c (l − 0.1nH) H √2gH

= 5.35c (l − 0.1nH) H^3/2.

This is Francis’s formula, in which the coefficient of discharge c is much more nearly constant for different values of l and h than in the ordinary formula. Francis found for c the mean value 0.622, the weir being sharp-edged.

§ 43. Triangular Notch (fig. 46).—Consider a lamina issuing between the depths h and h + dh. Its area, neglecting contraction, will be bdh, and the velocity at that depth is √(2gh). Hence the discharge for this lamina is

b√2gh dh.

But

B/b = H / (H − h); b = B (H − h) / H.

Hence discharge of lamina

= B(H − h) √(2gh) dh/H;

and total discharge of notch

= Q = B √(2g) ∫H0 (H − h) h^1/2 dh/H

= 4/15 B √(2g) H^3/2.

or, introducing a coefficient to allow for contraction,

Q = 4/15 cB √(2g) H^1/2,

Fig. 46.

When a notch is used to gauge a stream of varying flow, the ratio B/H varies if the notch is rectangular, but is constant if the notch is triangular. This led Professor James Thomson to suspect that the coefficient of discharge, c, would be much more constant with different values of H in a triangular than in a rectangular notch, and this has been experimentally shown to be the case. Hence a triangular notch is more suitable for accurate gaugings than a rectangular notch. For a sharp-edged triangular notch Professor J. Thomson found c = 0.617. It will be seen, as in § 41, that since 1/2BH is the area of section of the stream through the notch, the formula is again of the form

Q = c × 1/2BH × k √(2gH),

where k = 8/15 is the ratio of the mean velocity in the notch to the velocity at the depth H. It may easily be shown that for all notches the discharge can be expressed in this form.

Coefficients for the Discharge over Weirs, derived from the Experiments of T. E. Blackwell. When more than one experiment was made with the

same head, and the results were pretty uniform, the resulting coefficients are marked with an (*). The effect of the converging wing-boards is very strongly marked.

Heads in inches measured from still Water in Reservoir.	Sharp Edge.		Planks 2 in. thick, square on Crest.				Crests 3 ft. wide.
	3 ft. long.	10 ft. long.	3 ft. long.	6 ft. long.	10 ft. long.	10 ft. long, wing-boards making an angle of 60°.	3 ft. long. level.	3 ft. long, fall 1 in 18.	3 ft. long, fall 1 in 12.	6 ft. long. level.	10 ft. long. level.	10 ft. long, fall 1 in 18.
1	.677	.809	.467	.459	.435[4]	.754	.452	.545	.467	..	.381	.467
2	.675	.803	.509*	.561	.585*	.675	.482	.546	.533	..	.479*	.495*
3	.630	.642*	.563*	.597*	.569*	..	.441	.537	.539	.492*	..	..
4	.617	.656	.549	.575	.602*	.656	.419	.431	.455	.497*	..	.515
5	.602	.650*	.588	.601*	.609*	.671	.479	.516	..	..	.518	..
6	.593	..	.593*	.608*	.576*	..	.501*	..	.531	.507	.513	.543
7	..	..	.617*	.608*	.576*	..	.488	.513	.527	.497	..	..
8	..	.581	.606*	.590*	.548*	..	.470	.491	..	..	.468	.507
9	..	.530	.600	.569*	.558*	..	.476	.492*	.498	.480*	.486	..
10	..	..	.614*	.539	.534*	..	..	..	..	.465*	.455	..
12	..	..	..	.525	.534*	..	..	..	..	.467*	..	..
14	..	..	..	.549*	..	..	..	..	..	..	..	..

Fig. 47.

§ 44. Weir with a Broad Sloping Crest.—Suppose a weir formed with a broad crest so sloped that the streams flowing over it have a movement sensibly rectilinear and uniform (fig. 47). Let the inner edge be so rounded as to prevent a crest contraction. Consider a filament aa′, the point a being so far back from the weir that the velocity of approach is negligible. Let OO be the surface level in the reservoir, and let a be at a height h″ below OO, and h′ above a′. Let h be the distance from OO to the weir crest and e the thickness of the stream upon it. Neglecting atmospheric pressure, which has no influence, the pressure at a is Gh″; at a′ it is Gz. If v be the velocity at a′,

v²/2g = h′ + h″ − z = h − e;

Q = be √2g (h − e).

Theory does not furnish a value for e, but Q = 0 for e = 0 and for e = h. Q has therefore a maximum for a value of e between 0 and h, obtained by equating dQ/de to zero. This gives e = 2/3h, and, inserting this value,

Q = 0.385 bh √2gh,

as a maximum value of the discharge with the conditions assigned. Experiment shows that the actual discharge is very approximately equal to this maximum, and the formula is more legitimately applicable to the discharge over broad-crested weirs and to cases such as the discharge with free upper surface through large masonry sluice openings than the ordinary weir formula for sharp-edged weirs. It should be remembered, however, that the friction on the sides and crest of the weir has been neglected, and that this tends to reduce a little the discharge. The formula is equivalent to the ordinary weir formula with c = 0.577.

Special Cases of Discharge from Orifices

§ 45. Cases in which the Velocity of Approach needs to be taken into Account. Rectangular Orifices and Notches.—In finding the velocity at the orifice in the preceding investigations, it has been assumed that the head h has been measured from the free surface of still water above the orifice. In many cases which occur in practice the channel of approach to an orifice or notch is not so large, relatively to the stream through the orifice or notch, that the velocity in it can be disregarded.

Fig. 48.

Let h₁, h₂ (fig. 48) be the heads measured from the free surface to the top and bottom edges of a rectangular orifice, at a point in the channel of approach where the velocity is u. It is obvious that a fall of the free surface,

ɧ = u²/2g

has been somewhere expended in producing the velocity u, and hence the true heads measured in still water would have been h₁ + ɧ and h₂ + ɧ. Consequently the discharge, allowing for the velocity of approach, is

Q = 2/3 cb √2g { (h ₂ + ɧ)^3/2 − (h ₁ + ɧ)^3/2 }.

(1)

And for a rectangular notch for which h₁ = 0, the discharge is

Q = 2/3 cb √2g { (h ₂ + ɧ)^3/2 − ɧ^3/2 }.

(2)

In cases where u can be directly determined, these formulae give the discharge quite simply. When, however, u is only known as a function of the section of the stream in the channel of approach, they become complicated. Let Ω be the sectional area of the channel where h₁ and h₂ are measured. Then u = Q/Ω and ɧ = Q²/2g Ω².

This value introduced in the equations above would render them excessively cumbrous. In cases therefore where Ω only is known, it is best to proceed by approximation. Calculate an approximate value Q′ of Q by the equation

Q′ = 2/3 cb √2g {h₂^3/2 − h₁^3/2 }.

Then ɧ = Q′²/2gΩ² nearly. This value of ɧ introduced in the equations above will give a second and much more approximate value of Q.

Fig. 49.

§ 46. Partially Submerged Rectangular Orifices and Notches.—When the tail water is above the lower but below the upper edge of the orifice, the flow in the two parts of the orifice, into which it is divided by the surface of the tail water, takes place under different conditions. A filament M₁m₁ (fig. 49) in the upper part of the orifice issues with a head h′ which may have any value between h₁ and h. But a filament M₂m₂ issuing in the lower part of the orifice has a velocity due to h″ − h″′, or h, simply. In the upper part of the orifice the head is variable, in the lower constant. If Q₁, Q₂ are the discharges from the upper and lower parts of the orifice, b the width of the orifice, then

Q₁ = 2/3 cb √2g { h^3/2 − h₁^3/2 }

Q₂ = cb (h ₂ − h) √2gh.

(3)

In the case of a rectangular notch or weir, h₁ = 0. Inserting this value, and adding the two portions of the discharge together, we get for a drowned weir

Q = cb √2gh (h ₂ − h/3),

(4)

where h is the difference of level of the head and tail water, and h₂ is the head from the free surface above the weir to the weir crest (fig. 50).

From some experiments by Messrs A. Fteley and F.P. Stearns (Trans. Am. Soc. C.E., 1883, p. 102) some values of the coefficient c can be reduced

h3/h2	c	h3/h2	c
0.1	0.629	0.7	0.578
0.2	0.614	0.8	0.583
0.3	0.600	0.9	0.596
0.4	0.590	0.95	0.607
0.5	0.582	1.00	0.628
0.6	0.578

If velocity of approach is taken into account, let ɧ be the head due to that velocity; then, adding ɧ to each of the heads in the equations (3), and reducing, we get for a weir

Q = cb √2g [ (h ₂ + ɧ) (h + ɧ)^1/2 − 1/3 (h + ɧ)^3/2 − 2/3 ɧ^3/2 ];

(5)

an equation which may be useful in estimating flood discharges.

Fig. 50.

Bridge Piers and other Obstructions in Streams.—When the piers of a bridge are erected in a stream they create an obstruction to the flow of the stream, which causes a difference of surface-level above and below the pier (fig. 51). If it is necessary to estimate this difference of level, the flow between the piers may be treated as if it occurred over a drowned weir. But the value of c in this case is imperfectly known.

§ 47. Bazin’s Researches on Weirs.—H. Bazin has executed a long series of researches on the flow over weirs, so systematic and complete that they almost supersede other observations. The account of them is contained in a series of papers in the Annales des Ponts et Chaussées (October 1888, January 1890, November 1891, February 1894, December 1896, 2nd trimestre 1898). Only a very abbreviated account can be given here. The general plan of the experiments was to establish first the coefficients of discharge for a standard weir without end contractions; next to establish weirs of other types in series with the standard weir on a channel with steady flow, to compare the observed heads on the different weirs and to determine their coefficients from the discharge computed at the standard weir. A channel was constructed parallel to the Canal de Bourgogne, taking water from it through three sluices 0.3 × 1.0 metres. The water enters a masonry chamber 15 metres long by 4 metres wide where it is stilled and passes into the canal at the end of which is the standard weir. The canal has a length of 15 metres, a width of 2 metres and a depth of 0.6 metres. From this extends a channel 200 metres in length with a slope of 1 mm. per metre. The channel is 2 metres wide with vertical sides. The channels were constructed of concrete rendered with cement. The water levels were taken in chambers constructed near the canal, by floats actuating an index on a dial. Hook gauges were used in determining the heads on the weirs.

Fig. 51.

Standard Weir.—The weir crest was 3.72 ft. above the bottom of the canal and formed by a plate 1/4 in. thick. It was sharp-edged with free overfall. It was as wide as the canal so that end contractions were suppressed, and enlargements were formed below the crest to admit air under the water sheet. The channel below the weir was used as a gauging tank. Gaugings were made with the weir 2 metres in length and afterwards with the weir reduced to 1 metre and 0.5 metre in length, the end contractions being suppressed in all cases. Assuming the general formula

Q = mlh √(2gh),

(1)

Bazin arrives at the following values of m:—

Coefficients of Discharge of Standard Weir.

Head h metres.	Head h feet.	m
0.05	.164	0.4485
0.10	.328	0.4336
0.15	.492	0.4284
0.20	.656	0.4262
0.25	.820	0.4259
0.30	.984	0.4266
0.35	1.148	0.4275
0.40	1.312	0.4286
0.45	1.476	0.4299
0.50	1.640	0.4313
0.55	1.804	0.4327
0.60	1.968	0.4341

Bazin compares his results with those of Fteley and Stearns in 1877 and 1879, correcting for a different velocity of approach, and finds a close agreement.

Influence of Velocity of Approach.—To take account of the velocity of approach u it is usual to replace h in the formula by h + au²/2g where α is a coefficient not very well ascertained. Then

Q = μl (h + αu²/2g) √ { 2g (h + αu²/2g) }

= μlh √(2gh) (1 + αu²/2gh)^3/2.

(2)

The original simple equation can be used if

m = μ (1 + αu²/2gh)^3/2

or very approximately, since u²/2gh is small,

m = μ (1 + 3/2αu²/2gh).

(3)

Fig. 52.

Now if p is the height of the weir crest above the bottom of the canal (fig. 52), u = Q/l(p + h). Replacing Q by its value in (1)

u²/2gh = Q² / {2ghl²(p + h)²} = m² {h/(p + h) }²,

(4)

so that (3) may be written

m = μ [1 + k {h/(p + h)}² ].

(5)

Gaugings were made with weirs of 0.75, 0.50, 0.35, and 0.24 metres height above the canal bottom and the results compared with those of the standard weir taken at the same time. The discussion of the results leads to the following values of m in the general equation (1):—

m = μ (1 + 2.5u²/2gh)

= μ [1 + 0.55 {h/(p + h)}² ].

Values of μ—

Head h metres.	Head h feet.	μ
0.05	.164	0.4481
0.10	.328	0.4322
0.20	.656	0.4215
0.30	.984	0.4174
0.40	1.312	0.4144
0.50	1.640	0.4118
0.60	1.968	0.4092

An approximate formula for μ is:

μ = 0.405 + 0.003/h (h in metres)
μ = 0.405 + 0.01/h (h in feet).  

Inclined Weirs.—-Experiments were made in which the plank weir was inclined up or down stream, the crest being sharp and the end contraction suppressed. The following are coefficients by which the discharge of a vertical weir should be multiplied to obtain the discharge of the inclined weir.

			Coefficient.
Inclination up stream		1 to 1	0.93
”	”	3 to 2	0.94
”	”	3 to 1	0.96
Vertical weir		..	1.00
Inclination	down stream	3 to 1	1.04
”	”	3 to 2	1.07
”	”	1 to 1	1.10
”	”	1 to 2	1.12
”	”	1 to 4	1.09

The coefficient varies appreciably, if h/p approaches unity, which case should be avoided.

Fig. 53.

Fig. 54.

In all the preceding cases the sheet passing over the weir is detached completely from the weir and its under-surface is subject to atmospheric pressure. These conditions permit the most exact determination of the coefficient of discharge. If the sides of the canal below the weir are not so arranged as to permit the access of air under the sheet, the phenomena are more complicated. So long as the head does not exceed a certain limit the sheet is detached from the weir, but encloses a volume of air which is at less than atmospheric pressure, and the tail water rises under the sheet. The discharge is a little greater than for free overfall. At greater head the air disappears from below the sheet and the sheet is said to be “drowned.” The drowned sheet may be independent of the tail water level or influenced by it. In the former case the fall is followed by a rapid, terminating in a standing wave. In the latter case when the foot of the sheet is drowned the level of the tail water influences the discharge even if it is below the weir crest.

Weirs with Flat Crests.—The water sheet may spring clear from the upstream edge or may adhere to the flat crest falling free beyond the down-stream edge. In the former case the condition is that of a sharp-edged weir and it is realized when the head is at least double the width of crest. It may arise if the head is at least 11/2 the width of crest. Between these limits the condition of the sheet is unstable. When the sheet is adherent the coefficient m depends on the ratio of the head h to the width of crest c (fig. 53), and is given by the equation m = m₁ [0.70 + 0.185h/c], where m₁ is the coefficient for a sharp-edged weir in similar conditions. Rounding the upstream edge even to a small extent modifies the discharge. If R is the radius of the rounding the coefficient m is increased in the ratio 1 to 1 + R/h nearly. The results are limited to R less than 1/2 in.

Drowned Weirs.—Let h (fig. 54) be the height of head water and h₁ that of tail water above the weir crest. Then Bazin obtains as the approximate formula for the coefficient of discharge

m = 1.05m₁ [1 + 1/5 h₁/p] ³√ { (h − h₁) / h },

Fig. 55.

where as before m₁ is the coefficient for a sharp-edged weir in similar conditions, that is, when the sheet is free and the weir of the same height.

§ 48. Separating Weirs.—Many towns derive their water-supply from streams in high moorland districts, in which the flow is extremely variable. The water is collected in large storage reservoirs, from which an uniform supply can be sent to the town. In such cases it is desirable to separate the coloured water which comes down the streams in high floods from the purer water of ordinary flow. The latter is sent into the reservoirs; the former is allowed to flow away down the original stream channel, or is stored in separate reservoirs and used as compensation water. To accomplish the separation of the flood and ordinary water, advantage is taken of the different horizontal range of the parabolic path of the water falling over a weir, as the depth on the weir and, consequently, the velocity change. Fig. 55 shows one of these separating weirs in the form in which they were first introduced on the Manchester Waterworks; fig. 56 a more modern weir of the same kind designed by Sir A. Binnie for the Bradford Waterworks. When the quantity of water coming down the stream is not excessive, it drops over the weir into a transverse channel leading to the reservoirs. In flood, the water springs over the mouth of this channel and is led into a waste channel.

Fig. 56.

It may be assumed, probably with accuracy enough for practical purposes, that the particles describe the parabolas due to the mean velocity of the water passing over the weir, that is, to a velocity

2/3 √(2gh),

where h is the head above the crest of the weir.

Let cb = x be the width of the orifice and ac = y the difference of level of its edges (fig. 57). Then, if a particle passes from a to b in t seconds,

y = 1/2 gt², x = 2/3 √(2gh)t;

∴ y = 9/16 x²/h,

which gives the width x for any given difference of level y and head h, which the jet will just pass over the orifice. Set off ad vertically and equal to 1/2g on any scale; af horizontally and equal to 2/3 √(gh). Divide af, fe into an equal number of equal parts. Join a with the divisions on ef. The intersections of these lines with verticals from the divisions on af give the parabolic path of the jet.

Fig. 57.

Mouthpieces—Head Constant

§ 49. Cylindrical Mouthpieces.—When water issues from a short cylindrical pipe or mouthpiece of a length at least equal to 11/2 times its smallest transverse dimension, the stream, after contraction within the mouthpiece, expands to fill it and issues full bore, or without contraction, at the point of discharge. The discharge is found to be about one-third greater than that from a simple orifice of the same size. On the other hand, the energy of the fluid per unit of weight is less than that of the stream from a simple orifice with the same head, because part of the energy is wasted in eddies produced at the point where the stream expands to fill the mouthpiece, the action being something like that which occurs at an abrupt change of section.

Let fig. 58 represent a vessel discharging through a cylindrical mouthpiece at the depth h from the free surface, and let the axis of the jet XX be taken as the datum with reference to which the head is estimated. Let Ω be the area of the mouthpiece, ω the area of the stream at the contracted section EF. Let v, p be the velocity and pressure at EF, and v₁, p₁ the same quantities at GH. If the discharge is into the air, p₁ is equal to the atmospheric pressure p_a.

The total head of any filament which goes to form the jet, taken at a point where its velocity is sensibly zero, is h + p_a/G; at EF the total head is v²/2g + p/G; at GH it is v₁²/2g + p₁/G.

Between EF and GH there is a loss of head due to abrupt change of velocity, which from eq. (3), § 36, may have the value

(v − v₁)²/2g.

Adding this head lost to the head at GH, before equating it to the heads at EF and at the point where the filaments start into motion,—

h + p_a/G = v²/2g + p/G = v₁²/2g + p₁/G + (v − v₁)²/2g.

But ωv = Ωv₁, and ω = c_cΩ, if c_c is the coefficient of contraction within the mouthpiece. Hence

v = Ωv₁/ω = v₁/c_c.

Supposing the discharge into the air, so that p₁ = p_a,

h + p_a/G = v₁²/2g + p_a/G + (v₁²/2g) (1/c_c − 1)²;

(v₁/2g) {1 + (1/c_c − 1)²} = h;

∴ v₁ = √(2gh) / √ {1 + (1/c_c − 1)² };

(1)

Fig. 58.

where the coefficient on the right is evidently the coefficient of velocity for the cylindrical mouthpiece in terms of the coefficient of contraction at EF. Let c_c = 0.64, the value for simple orifices, then the coefficient of velocity is

c_v = 1/√ {1 + (1/c_c − 1)² } = 0.87

(2)

The actual value of c_v, found by experiment is 0.82, which does not differ more from the theoretical value than might be expected if the friction of the mouthpiece is allowed for. Hence, for mouthpieces of this kind, and for the section at GH,

c_v = 0.82   c_c = 1.00   c = 0.82,

Q = 0.82Ω √(2gh).

It is easy to see from the equations that the pressure p at EF is less than atmospheric pressure. Eliminating v₁, we get

(p_a − p)/G = 3/4 h nearly;

(3)

or

p = p_a − 3/4 Gh ℔ per sq. ft.

If a pipe connected with a reservoir on a lower level is introduced into the mouthpiece at the part where the contraction is formed (fig. 59), the water will rise in this pipe to a height

KL = (p_a − p) / G = 3/4 h nearly.

If the distance X is less than this, the water from the lower reservoir will be forced continuously into the jet by the atmospheric pressure, and discharged with it. This is the crudest form of a kind of pump known as the jet pump.

§ 50. Convergent Mouthpieces.—With convergent mouthpieces there is a contraction within the mouthpiece causing a loss of head, and a diminution of the velocity of discharge, as with cylindrical mouthpieces. There is also a second contraction of the stream outside the mouthpiece. Hence the discharge is given by an equation of the form

Q = c_vc_cΩ √(2gh),

(4)

where Ω is the area of the external end of the mouthpiece, and c_cΩ the section of the contracted jet beyond the mouthpiece.

Convergent Mouthpieces (Castel’s Experiments).—Smallest diameter of
orifice = 0.05085 ft. Length of mouthpiece = 2.6 Diameters.

Angle of Convergence.	Coefficient of Contraction, cc	Coefficient of Velocity, cv	Coefficient of Discharge, c
0°  0′	.999	.830	.829
1° 36′	1.000	.866	.866
3° 10′	1.001	.894	.895
4° 10′	1.002	.910	.912
5° 26′	1.004	.920	.924
7° 52′	.998	.931	.929
8° 58′	.992	.942	.934
10° 20′	.987	.950	.938
12°  4′	.986	.955	.942
13° 24′	.983	.962	.946
14° 28′	.979	.966	.941
16° 36′	.969	.971	.938
19° 28′	.953	.970	.924
21°  0′	.945	.971	.918
23°  0′	.937	.974	.913
29° 58′	.919	.975	.896
40° 20′	.887	.980	.869
48° 50′	.861	.984	.847

The maximum coefficient of discharge is that for a mouthpiece with a convergence of 13°24′.

Fig. 59.	Fig. 60.

The values of c_v and c_c must here be determined by experiment. The above table gives values sufficient for practical purposes. Since the contraction beyond the mouthpiece increases with the convergence, or, what is the same thing, c_c diminishes, and on the other hand the loss of energy diminishes, so that c_v increases with the convergence, there is an angle for which the product c_c c_v, and consequently the discharge, is a maximum.

§ 51. Divergent Conoidal Mouthpiece.—Suppose a mouthpiece so designed that there is no abrupt change in the section or velocity of the stream passing through it. It may have a form at the inner end approximately the same as that of a simple contracted vein, and may then enlarge gradually, as shown in fig. 60. Suppose that at EF it becomes cylindrical, so that the jet may be taken to be of the diameter EF. Let ω, v, p be the section, velocity and pressure at CD, and Ω, v₁, p₁ the same quantities at EF, p_a being as usual the atmospheric pressure, or pressure on the free surface AB. Then, since there is no loss of energy, except the small frictional resistance of the surface of the mouthpiece,

h + p_a/G = v²/2g + p/G
= v₁²/2g + p₁/G.

If the jet discharges into the air, p₁ = p_a; and

v₁²/2g = h;

v₁ = √(2gh);

or, if a coefficient is introduced to allow for friction,

v₁ = c_v √(2gh);

where c_v is about 0.97 if the mouthpiece is smooth and well formed.

Q = Ωv₁ = c_v Ω √(2gh).

Fig. 61.

Hence the discharge depends on the area of the stream at EF, and not at all on that at CD, and the latter may be made as small as we please without affecting the amount of water discharged.

There is, however, a limit to this. As the velocity at CD is greater than at EF the pressure is less, and therefore less than atmospheric pressure, if the discharge is into the air. If CD is so contracted that p = 0, the continuity of flow is impossible. In fact the stream disengages itself from the mouthpiece for some value of p greater than 0 (fig. 61).

From the equations,

p/G = p_a/G − (v² − v₁²) / 2g.

Let Ω/ω = m. Then

v = v₁m;

p/G = p_a/G − v₁² (m² − 1) / 2g

= p_a/G − (m² − 1)h;

whence we find that p/G will become zero or negative if

Ω/ω ≧ √ {(h + p_a/G) / h }
= √ {1 + p_a/Gh};

or, putting p_a/G = 34 ft., if

Ω/ω ≧ √ { (h + 34)/h}.

In practice there will be an interruption of the full bore flow with a less ratio of Ω/ω, because of the disengagement of air from the water. But, supposing this does not occur, the maximum discharge of a mouthpiece of this kind is

Q = ω √ {2g (h + p_a/G) };

that is, the discharge is the same as for a well-bell-mouthed mouthpiece of area ω, and without the expanding part, discharging into a vacuum.

§ 52. Jet Pump.—A divergent mouthpiece may be arranged to act as a pump, as shown in fig. 62. The water which supplies the energy required for pumping enters at A. The water to be pumped enters at B. The streams combine at DD where the velocity is greatest and the pressure least. Beyond DD the stream enlarges in section, and its pressure increases, till it is sufficient to balance the head due to the height of the lift, and the water flows away by the discharge pipe C.

Fig. 62.

Fig. 63 shows the whole arrangement in a diagrammatic way. A is the reservoir which supplies the water that effects the pumping; B is the reservoir of water to be pumped; C is the reservoir into which the water is pumped.

Fig. 63.

Discharge with Varying Head

§ 53. Flow from a Vessel when the Effective Head varies with the Time.—Various useful problems arise relating to the time of emptying and filling vessels, reservoirs, lock chambers, &c., where the flow is dependent on a head which increases or diminishes during the operation. The simplest of these problems is the case of filling or emptying a vessel of constant horizontal section.

Fig. 64.

Time of Emptying or Filling a Vertical-sided Lock Chamber.—Suppose the lock chamber, which has a water surface of Ω square ft., is emptied through a sluice in the tail gates, of area ω, placed below the tail-water level. Then the effective head producing flow through the sluice is the difference of level in the chamber and tail bay. Let H (fig. 64) be the initial difference of level, h the difference of level after t seconds. Let −dh be the fall of level in the chamber during an interval dt. Then in the time dt the volume in the chamber is altered by the amount −Ωdh, and the outflow from the sluice in the same time is cω √(2gh)dt. Hence the differential equation connecting h and t is

cω √(2gh)dt + Ωh = 0.

For the time t, during which the initial head H diminishes to any other value h,

−{Ω/(cω √2g) } $\int _{H}^{h}$ dh/√h = $\int _{0}^{t}$ dt.

∴ t = 2Ω (√H − √h) / {cω √(2g)}

= (Ω/cω) {√(2H/g) − √(2h/g) }.

For the whole time of emptying, during which h diminishes from H to 0,

T = (Ω/cω) √(2H/g).

Comparing this with the equation for flow under a constant head, it will be seen that the time is double that required for the discharge of an equal volume under a constant head.

The time of filling the lock through a sluice in the head gates is exactly the same, if the sluice is below the tail-water level. But if the sluice is above the tail-water level, then the head is constant till the level of the sluice is reached, and afterwards it diminishes with the time.

Practical Use of Orifices in Gauging Water

§ 54. If the water to be measured is passed through a known orifice under an arrangement by which the constancy of the head is ensured, the amount which passes in a given time can be ascertained by the formulae already given. It will obviously be best to make the orifices of the forms for which the coefficients are most accurately determined; hence sharp-edged orifices or notches are most commonly used.

Water Inch.—For measuring small quantities of water circular sharp-edged orifices have been used. The discharge from a circular orifice one French inch in diameter, with a head of one line above the top edge, was termed by the older hydraulic writers a water-inch. A common estimate of its value was 14 pints per minute, or 677 English cub. ft. in 24 hours. An experiment by C. Bossut gave 634 cub. ft. in 24 hours (see Navier’s edition of Belidor’s Arch. Hydr., p. 212).

L. J. Weisbach points out that measurements of this kind would be made more accurately with a greater head over the orifice, and he proposes that the head should be equal to the diameter of the orifice. Several equal orifices may be used for larger discharges.

Fig. 65.

Pin Ferrules or Measuring Cocks.—To give a tolerably definite supply of water to houses, without the expense of a meter, a ferrule with an orifice of a definite size, or a cock, is introduced in the service-pipe. If the head in the water main is constant, then a definite quantity of water would be delivered in a given time. The arrangement is not a very satisfactory one, and acts chiefly as a check on extravagant use of water. It is interesting here chiefly as an example of regulation of discharge by means of an orifice. Fig. 65 shows a cock of this kind used at Zurich. It consists of three cocks, the middle one having the orifice of the predetermined size in a small circular plate, protected by wire gauze from stoppage by impurities in the water. The cock on the right hand can be used by the consumer for emptying the pipes. The one on the left and the measuring cock are connected by a key which can be locked by a padlock, which is under the control of the water company.

§ 55. Measurement of the Flow in Streams.—To determine the quantity of water flowing off the ground in small streams, which is available for water supply or for obtaining water power, small temporary weirs are often used. These may be formed of planks supported by piles and puddled to prevent leakage. The measurement of the head may be made by a thin-edged scale at a short distance behind the weir, where the water surface has not begun to slope down to the weir and where the velocity of approach is not high. The measurements are conveniently made from a short pile driven into the bed of the river, accurately level with the crest of the weir (fig. 66). Then if at any moment the head is h, the discharge is, for a rectangular notch of breadth b,

Q = 2/3 cbh√2gh

where c = 0.62; or, better, the formula in § 42 may be used.

Gauging weirs are most commonly in the form of rectangular notches; and care should be taken that the crest is accurately horizontal, and that the weir is normal to the direction of flow of the stream. If the planks are thick, they should be bevelled (fig. 67), and then the edge may be protected by a metal plate about 1/10th in. thick to secure the requisite accuracy of form and sharpness of edge. In permanent gauging weirs, a cast steel plate is sometimes used to form the edge of the weir crest. The weir should be large enough to discharge the maximum volume flowing in the stream, and at the same time it is desirable that the minimum head should not be too small (say half a foot) to decrease the effects of errors of measurement. The section of the jet over the weir should not exceed one-fifth the section of the stream behind the weir, or the velocity of approach will need to be taken into account. A triangular notch is very suitable for measurements of this kind.

Fig. 66.

If the flow is variable, the head h must be recorded at equidistant intervals of time, say twice daily, and then for each 12-hour period the discharge must be calculated for the mean of the heads at the beginning and end of the time. As this involves a good deal of troublesome calculation, E. Sang proposed to use a scale so graduated as to read off the discharge in cubic feet per second. The lengths of the principal graduations of such a scale are easily calculated by putting Q = 1, 2, 3 . . . in the ordinary formulae for notches; the intermediate graduations may be taken accurately enough by subdividing equally the distances between the principal graduations.

Fig. 67.

Fig. 68.

The accurate measurement of the discharge of a stream by means of a weir is, however, in practice, rather more difficult than might be inferred from the simplicity of the principle of the operation. Apart from the difficulty of selecting a suitable coefficient of discharge, which need not be serious if the form of the weir and the nature of its crest are properly attended to, other difficulties of measurement arise. The length of the weir should be very accurately determined, and if the weir is rectangular its deviations from exactness of level should be tested. Then the agitation of the water, the ripple on its surface, and the adhesion of the water to the scale on which the head is measured, are liable to introduce errors. Upon a weir 10 ft. long, with 1 ft. depth of water flowing over, an error of 1-1000th of a foot in measuring the head, or an error of 1-100th of a foot in measuring the length of the weir, would cause an error in computing the discharge of 2 cub. ft. per minute.

Hook Gauge.—For the determination of the surface level of water, the most accurate instrument is the hook gauge used first by U. Boyden of Boston, in 1840. It consists of a fixed frame with scale and vernier. In the instrument in fig. 68 the vernier is fixed to the frame, and the scale slides vertically. The scale carries at its lower end a hook with a fine point, and the scale can be raised or lowered by a fine pitched screw. If the hook is depressed below the water surface and then raised by the screw, the moment of its reaching the water surface will be very distinctly marked, by the reflection from a small capillary elevation of the water surface over the point of the hook. In ordinary light, differences of level of the water of .001 of a foot are easily detected by the hook gauge. If such a gauge is used to determine the heads at a weir, the hook should first be set accurately level with the weir crest, and a reading taken. Then the difference of the reading at the water surface and that for the weir crest will be the head at the weir.

§ 56. Modules used in Irrigation.—In distributing water for irrigation, the charge for the water may be simply assessed on the area of the land irrigated for each consumer, a method followed in India; or a regulated quantity of water may be given to each consumer, and the charge may be made proportional to the quantity of water supplied, a method employed for a long time in Italy and other parts of Europe. To deliver a regulated quantity of water from the irrigation channel, arrangements termed modules are used. These are constructions intended to maintain a constant or approximately constant head above an orifice of fixed size, or to regulate the size of the orifice so as to give a constant discharge, notwithstanding the variation of level in the irrigating channel.

Fig. 69.

§ 57. Italian Module.—The Italian modules are masonry constructions, consisting of a regulating chamber, to which water is admitted by an adjustable sluice from the canal. At the other end of the chamber is an orifice in a thin flagstone of fixed size. By means of the adjustable sluice a tolerably constant head above the fixed orifice is maintained, and therefore there is a nearly constant discharge of ascertainable amount through the orifice, into the channel leading to the fields which are to be irrigated.

Fig. 70.—Scale 1/100.

In fig. 69, A is the adjustable sluice by which water is admitted to the regulating chamber, B is the fixed orifice through which the water is discharged. The sluice A is adjusted from time to time by the canal officers, so as to bring the level of the water in the regulating chamber to a fixed level marked on the wall of the chamber. When adjusted it is locked. Let ω₁ be the area of the orifice through the sluice at A, and ω₂ that of the fixed orifice at B; let h₁ be the difference of level between the surface of the water in the canal and regulating chamber; h₂ the head above the centre of the discharging orifice, when the sluice has been adjusted and the flow has become steady; Q the normal discharge in cubic feet per second. Then, since the flow through the orifices at A and B is the same,

Q = c₁ω₁ √(2gh₁) = c₂ω₂ √(2gh₂),

where c₁ and c₂ are the coefficients of discharge suitable for the two orifices. Hence

c₁ω₁ / c₂ω₂ = √(h ₂/h₁).

If the orifice at B opened directly into the canal without any intermediate regulating chamber, the discharge would increase for a given change of level in the canal in exactly the same ratio. Consequently the Italian module in no way moderates the fluctuations of discharge, except so far as it affords means of easy adjustment from time to time. It has further the advantage that the cultivator, by observing the level of the water in the chamber, can always see whether or not he is receiving the proper quantity of water.

On each canal the orifices are of the same height, and intended to work with the same normal head, the width of the orifices being varied to suit the demand for water. The unit of discharge varies on different canals, being fixed in each case by legal arrangements. Thus on the Canal Lodi the unit of discharge or one module of water is the discharge through an orifice 1.12 ft. high, 0.12416 ft. wide, with a head of 0.32 ft. above the top edge of the orifice, or .88 ft. above the centre. This corresponds to a discharge of about 0.6165 cub. ft. per second.

Fig. 71.

In the most elaborate Italian modules the regulating chamber is arched over, and its dimensions are very exactly prescribed. Thus in the modules of the Naviglio Grande of Milan, shown in fig. 70, the measuring orifice is cut in a thin stone slab, and so placed that the discharge is into the air with free contraction on all sides. The adjusting sluice is placed with its sill flush with the bottom of the canal, and is provided with a rack and lever and locking arrangement. The covered regulating chamber is about 20 ft. long, with a breadth 1.64 ft. greater than that of the discharging orifice. At precisely the normal level of the water in the regulating chamber, there is a ceiling of planks intended to still the agitation of the water. A block of stone serves to indicate the normal level of the water in the chamber. The water is discharged into an open channel 0.655 ft. wider than the orifice, splaying out till it is 1.637 ft. wider than the orifice, and about 18 ft. in length.

§ 58. Spanish Module.—On the canal of Isabella II., which supplies water to Madrid, a module much more perfect in principle than the Italian module is employed. Part of the water is supplied for irrigation, and as it is very valuable its strict measurement is essential. The module (fig. 72) consists of two chambers one above the other, the upper chamber being in free communication with the irrigation canal, and the lower chamber discharging by a culvert to the fields. In the arched roof between the chambers there is a circular sharp-edged orifice in a bronze plate. Hanging in this there is a bronze plug of variable diameter suspended from a hollow brass float. If the water level in the canal lowers, the plug descends and gives an enlarged opening, and conversely. Thus a perfectly constant discharge with a varying head can be obtained, provided no clogging or silting of the chambers prevents the free discharge of the water or the rise and fall of the float. The theory of the module is very simple. Let R (fig. 71) be the radius of the fixed opening, r the radius of the plug at a distance h from the plane of flotation of the float, and Q the required discharge of the module. Then

Q = cπ (R² − r ²) √(2gh).

Taking c = 0.63,

Q = 15.88 (R² − r ²) √h;

r = √ {R² − Q/15.88 √h}.

Choosing a value for R, successive values of r can be found for different values of h, and from these the curve of the plug can be drawn. The module shown in fig. 72 will discharge 1 cubic metre per second. The fixed opening is 0.2 metre diameter, and the greatest head above the fixed orifice is 1 metre. The use of this module involves a great sacrifice of level between the canal and the fields. The module is described in Sir C. Scott-Moncrieff’s Irrigation in Southern Europe.

§ 59. Reservoir Gauging Basins.—In obtaining the power to store the water of streams in reservoirs, it is usual to concede to riparian owners below the reservoirs a right to a regulated supply throughout the year. This compensation water requires to be measured in such a way that the millowners and others interested in the matter can assure themselves that they are receiving a proper quantity, and they are generally allowed a certain amount of control as to the times during which the daily supply is discharged into the stream.

Fig. 72.

Fig. 74 shows an arrangement designed for the Manchester water works. The water enters from the reservoir at chamber A, the object of which is to still the irregular motion of the water. The admission is regulated by sluices at b, b, b. The water is discharged by orifices or notches at a, a, over which a tolerably constant head is maintained by adjusting the sluices at b, b, b. At any time the millowners can see whether the discharge is given and whether the proper head is maintained over the orifices. To test at any time the discharge of the orifices, a gauging basin B is provided. The water ordinarily flows over this, without entering it, on a floor of cast-iron plates. If the discharge is to be tested, the water is turned for a definite time into the gauging basin, by suddenly opening and closing a sluice at c. The volume of flow can be ascertained from the depth in the gauging chamber. A mechanical arrangement (fig. 73) was designed for securing an absolutely constant head over the orifices at a, a. The orifices were formed in a cast-iron plate capable of sliding up and down, without sensible leakage, on the face of the wall of the chamber. The orifice plate was attached by a link to a lever, one end of which rested on the wall and the other on floats f in the chamber A. The floats rose and fell with the changes of level in the chamber, and raised and lowered the orifice plate at the same time. This mechanical arrangement was not finally adopted, careful watching of the sluices at b, b, b, being sufficient to secure a regular discharge. The arrangement is then equivalent to an Italian module, but on a large scale.

Fig. 73.—Scale 1/120.

Fig. 74.—Scale 1/500.

§ 60. Professor Fleeming Jenkin’s Constant Flow Valve.—In the modules thus far described constant discharge is obtained by varying the area of the orifice through which the water flows. Professor F. Jenkin has contrived a valve in which a constant pressure head is obtained, so that the orifice need not be varied (Roy. Scot. Society of Arts, 1876). Fig. 75 shows a valve of this kind suitable for a 6-in. water main. The water arriving by the main C passes through an equilibrium valve D into the chamber A, and thence through a sluice O, which can be set for any required area of opening, into the discharging main B. The object of the arrangement is to secure a constant difference of pressure between the chambers A and B, so that a constant discharge flows through the stop valve O. The equilibrium valve D is rigidly connected with a plunger P loosely fitted in a diaphragm, separating A from a chamber B₂ connected by a pipe B₁ with the discharging main B. Any increase of the difference of pressure in A and B will drive the plunger up and close the equilibrium valve, and conversely a decrease of the difference of pressure will cause the descent of the plunger and open the equilibrium valve wider. Thus a constant difference of pressure is obtained in the chambers A and B. Let ω be the area of the plunger in square feet, p the difference of pressure in the chambers A and B in pounds per square foot, w the weight of the plunger and valve. Then if at any moment pω exceeds w the plunger will rise, and if it is less than w the plunger will descend. Apart from friction, and assuming the valve D to be strictly an equilibrium valve, since ω and w are constant, p must be constant also, and equal to w/ω. By making w small and ω large, the difference of pressure required to ensure the working of the apparatus may be made very small. Valves working with a difference of pressure of 1/2 in. of water have been constructed.

Fig. 75.—Scale 1/24.

VI. STEADY FLOW OF COMPRESSIBLE FLUIDS.

 

Fig. 76.

§ 61. External Work during the Expansion of Air.—If air expands without doing any external work, its temperature remains constant. This result was first experimentally demonstrated by J. P. Joule. It leads to the conclusion that, however air changes its state, the internal work done is proportional to the change of temperature. When, in expanding, air does work against an external resistance, either heat must be supplied or the temperature falls.

To fix the conditions, suppose 1 ℔ of air confined behind a piston of 1 sq. ft. area (fig. 76). Let the initial pressure be p₁ and the volume of the air v₁, and suppose this to expand to the pressure p₂ and volume v₂. If p and v are the corresponding pressure and volume at any intermediate point in the expansion, the work done on the piston during the expansion from v to v + dv is pdv, and the whole work during the expansion from v₁ to v₂, represented by the area abcd, is

$\int _{v_{1}}^{v_{2}}pdv.$

Amongst possible cases two may be selected.

Case 1.—So much heat is supplied to the air during expansion that the temperature remains constant. Hyperbolic expansion.

Then

$pv=p_{1}v_{1}$.

Work done during expansion per pound of air

= $\int _{v_{1}}^{v_{2}}pdv=p_{1}v_{1}\int _{v_{1}}^{v_{2}}dv/v$

= p1v1 logε v2/v1 = p1v1 logε p1/p2.

(1)

Since the weight per cubic foot is the reciprocal of the volume per pound, this may be written

(p1/G1) logε G1/G2.

(1a)

Then the expansion curve ab is a common hyperbola.

Case 2.—No heat is supplied to the air during expansion. Then the air loses an amount of heat equivalent to the external work done and the temperature falls. Adiabatic expansion.

In this case it can be shown that

pv^γ = p₁v₁^γ,

where γ is the ratio of the specific heats of air at constant pressure and volume. Its value for air is 1.408, and for dry steam 1.135.

Work done during expansion per pound of air.

= $\int _{v_{1}}^{v_{2}}$pdv = p₁v₁^γ $\int _{v_{1}}^{v_{2}}$ dv/v^γ

= −{p₁v₁^γ / (γ − 1)} {1/v₂^γ−1 − 1/v₁^γ−1}

= {p₁v₁^γ / (γ − 1)} {1/v₁^γ−1 − 1/v₂^γ−1}

={p1v1 / (γ − 1)} {1 − (v1/v2) γ−1}.

(2)

The value of p₁v₁ for any given temperature can be found from the data already given.

As before, substituting the weights G₁, G₂ per cubic foot for the volumes per pound, we get for the work of expansion

(p1/G1) {1/(γ − 1)} {1 − (G2/G1) γ−1},

(2a)

= p1v1 {1/(γ − 1)} {1 − (p2/p1) γ−1/γ}.

(2b)

 

Fig. 77.

§ 62. Modification of the Theorem of Bernoulli for the Case of a Compressible Fluid.—In the application of the principle of work to a filament of compressible fluid, the internal work done by the expansion of the fluid, or absorbed in its compression, must be taken into account. Suppose, as before, that AB (fig. 77) comes to A′B′ in a short time t. Let p₁, ω₁, v₁, G₁ be the pressure, sectional area of stream, velocity and weight of a cubic foot at A, and p₂, ω₂, v₂, G₂ the same quantities at B. Then, from the steadiness of motion, the weight of fluid passing A in any given time must be equal to the weight passing B:

G₁ω₁v₁t = G₂ω₂v₂t.

Let z₁, z₂ be the heights of the sections A and B above any given datum. Then the work of gravity on the mass AB in t seconds is

G₁ω₁v₁t (z₁ − z₂) = W (z₁ − z₂) t,

where W is the weight of gas passing A or B per second. As in the case of an incompressible fluid, the work of the pressures on the ends of the mass AB is

p₁ω₁v₁t − p₂ω₂v₂t,

= (p₁/G₁ − p₂/G₂) Wt.

The work done by expansion of Wt ℔ of fluid between A and B is Wt$\int _{v_{1}}^{v_{2}}$pdv The change of kinetic energy as before is (W/2g) (v₂² − v₁²) t. Hence, equating work to change of kinetic energy,

W(z₁ − z₂) t + (p₁/G₁ − p₂/G₂)Wt + Wt $\int _{v_{1}}^{v_{2}}$ pdv = (W/2g) (v₂² − v₁²) t;

∴ z1 + p1/G1 + v12/2g = z2 + p2/G2 + v22/2g − $\int _{v_{1}}^{v_{2}}$pdv.

(1)

Now the work of expansion per pound of fluid has already been given. If the temperature is constant, we get (eq. 1a, § 61)

Z₁ + P₁/G₁ + v₁²/2g = z₂ + p²/G₂ + v₂²/2g − (p₁/G₁) log_ε (G₁/G₂).

But at constant temperature p₁/G₁ = p₂/G₂;

∴ z1 + v12/2g = z2 + v22/2g − (p1/G1) logε (p1/p2),

(2)

or, neglecting the difference of level,

(v22 − v12) / 2g = (p1/G1) logε (p1/p2).

(2a)

Similarly, if the expansion is adiabatic (eq. 2a, § 61),

z1 + p1/G1 + v12/2g = z2 + p2/G2 + v22/2g − (p1/G1) {1/(γ − 1) } {1 − (p2/p1)(γ−1)/γ};

(3)

or, neglecting the difference of level,

(v22 − v12)/2g = (p1/G1) [1 + 1/(γ − 1) {1 − (p2/p1)(γ−1)/γ)} ] − p2/G2.

(3a)

It will be seen hereafter that there is a limit in the ratio p₁/p₂ beyond which these expressions cease to be true.

§ 63. Discharge of Air from an Orifice.—The form of the equation of work for a steady stream of compressible fluid is

z₁ + p₁/G₁ + v₁²/2g = z₂ + p₂/G₂ + v₂²/2g−(p₁/G₁) {1/(γ − 1)} {1 − (p₂/p₁^(γ−1)/γ},

the expansion being adiabatic, because in the flow of the streams of air through an orifice no sensible amount of heat can be communicated from outside.

Suppose the air flows from a vessel, where the pressure is p₁ and the velocity sensibly zero, through an orifice, into a space where the pressure is p₂. Let v₂ be the velocity of the jet at a point where the convergence of the streams has ceased, so that the pressure in the jet is also p₂. As air is light, the work of gravity will be small compared with that of the pressures and expansion, so that z₁z₂ may be neglected. Putting these values in the equation above—

p₁/G₁ = p₂/G₂ + v₂²/2g − (p₁/G₁) {1/(γ − 1)} {1 − (p₂/p₁)^(γ−1)/γ;

v₂²/2g = p₁/G₁ − p₂/G₂ + (p₁/G₁) {1/(γ − 1)} {1 − (p₂/p₁)^(γ−1)/γ}

= (p₁/G₁) {γ/(γ − 1) − (p₂/p₁)^{γ−1 /γ} / (γ − 1)} − p₂/G₂.

v22/2g = (p1/G1) {γ/(γ − 1)} {1 − (p2/p1)(γ−1)/γ};

(1)

an equation commonly ascribed to L. J. Weisbach (Civilingenieur, 1856), though it appears to have been given earlier by A. J. C. Barre de Saint Venant and L. Wantzel.

It has already (§ 9, eq. 4a) been seen that

p₁/G₁ = (p₀/G₀) (τ₁/τ₀)

where for air p₀ = 2116.8, G₀ = .08075 and τ₀ = 492.6.

v22/2g = {p0τ1γ / G0τ0 (γ − 1)} {1 − (p2/p1)(γ−1)/γ};

(2)

or, inserting numerical values,

v22/2g = 183.6τ1 {1 − (p2/p1)0.29};

(2a)

which gives the velocity of discharge v₂ in terms of the pressure and absolute temperature, p₁, τ₁, in the vessel from which the air flows, and the pressure p₂ in the vessel into which it flows.

Proceeding now as for liquids, and putting ω for the area of the orifice and c for the coefficient of discharge, the volume of air discharged per second at the pressure p₂ and temperature τ₂ is

Q₂ = cωv₂ = cω √ [(2gγp₁ / (γ − 1) G₁) (1 − (p₂/p₁)^(γ−1)/γ)]

= 108.7cω √ [τ1 {1 − (p2/p1)0.29}].

(3)

If the volume discharged is measured at the pressure p₁ and absolute temperature τ₁ in the vessel from which the air flows, let Q₁ be that volume; then

p₁Q₁^γ = p₂Q₂^γ;

Q₁ = (p₂/p₁)^1/γ Q₂;

Q₁ = cω √ [ {2gγp₁ / (γ − 1) G₁} {(p₂/p₁)^2/γ − (p₂/p₁)^(γ+1)/γ}].

Q₁ = cω √ [2gγp₁ψ / (γ − 1) G₁]

= 108.7cω √ (τ1ψ).

(4)

The weight of air at pressure p₁ and temperature τ₁ is

G₁ = p₁/53.2τ₁ ℔ per cubic foot.

Hence the weight of air discharged is

W = G₁Q₁ = cω √ [2gγp₁G₁ψ / (γ − 1)]

= 2.043cωp1 √ (ψ/τ1).

(5)

Weisbach found the following values of the coefficient of discharge c:—

Conoidal mouthpieces of the form of the	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \ \end{matrix}}\right\}\,}$
contracted vein with effective pressures		c =
of .23 to 1.1 atmosphere		0.97	to	0.99
Circular sharp-edged orifices		0.563	”	0.788
Short cylindrical mouthpieces		0.81	”	0.84
The same rounded at the inner end		0.92	”	0.93
Conical converging mouthpieces		0.90	”	0.99

§ 64. Limit to the Application of the above Formulae.—In the formulae above it is assumed that the fluid issuing from the orifice expands from the pressure p₁ to the pressure p₂, while passing from the vessel to the section of the jet considered in estimating the area ω. Hence p₂ is strictly the pressure in the jet at the plane of the external orifice in the case of mouthpieces, or at the plane of the contracted section in the case of simple orifices. Till recently it was tacitly assumed that this pressure p₂ was identical with the general pressure external to the orifice. R. D. Napier first discovered that, when the ratio p₂/p₁ exceeded a value which does not greatly differ from 0.5, this was no longer true. In that case the expansion of the fluid down to the external pressure is not completed at the time it reaches the plane of the contracted section, and the pressure there is greater than the general external pressure; or, what amounts to the same thing, the section of the jet where the expansion is completed is a section which is greater than the area c_cω of the contracted section of the jet, and may be greater than the area ω of the orifice. Napier made experiments with steam which showed that, so long as p₂/p₁ > 0.5, the formulae above were trustworthy, when p₂ was taken to be the general external pressure, but that, if p₂/p₁ < 0.5, then the pressure at the contracted section was independent of the external pressure and equal to 0.5p₁. Hence in such cases the constant value 0.5 should be substituted in the formulae for the ratio of the internal and external pressures p₂/p₁.

It is easily deduced from Weisbach’s theory that, if the pressure external to an orifice is gradually diminished, the weight of air discharged per second increases to a maximum for a value of the ratio

p₂/p₁ = {2/(γ + 1)}^γ−1/γ
= 0.527 for air
= 0.58 for dry steam.

For a further decrease of external pressure the discharge diminishes,—a result no doubt improbable. The new view of Weisbach’s formula is that from the point where the maximum is reached, or not greatly differing from it, the pressure at the contracted section ceases to diminish.

A. F. Fliegner showed (Civilingenieur xx., 1874) that for air flowing from well-rounded mouthpieces there is no discontinuity of the law of flow, as Napier’s hypothesis implies, but the curve of flow bends so sharply that Napier’s rule may be taken to be a good approximation to the true law. The limiting value of the ratio p₂/p₁, for which Weisbach’s formula, as originally understood, ceases to apply, is for air 0.5767; and this is the number to be substituted for p₂/p₁ in the formulae when p₂/p₁ falls below that value. For later researches on the flow of air, reference may be made to G. A. Zeuner’s paper (Civilingenieur, 1871), and Fliegner’s papers (ibid., 1877, 1878).

VII. FRICTION OF LIQUIDS.

§ 65. When a stream of fluid flows over a solid surface, or conversely when a solid moves in still fluid, a resistance to the motion is generated, commonly termed fluid friction. It is due to the viscosity of the fluid, but generally the laws of fluid friction are very different from those of simple viscous resistance. It would appear that at all speeds, except the slowest, rotating eddies are formed by the roughness of the solid surface, or by abrupt changes of velocity distributed throughout the fluid; and the energy expended in producing these eddying motions is gradually lost in overcoming the viscosity of the fluid in regions more or less distant from that where they are first produced.

The laws of fluid friction are generally stated thus:—

1. The frictional resistance is independent of the pressure between the fluid and the solid against which it flows. This may be verified by a simple direct experiment. C. H. Coulomb, for instance, oscillated a disk under water, first with atmospheric pressure acting on the water surface, afterwards with the atmospheric pressure removed. No difference in the rate of decrease of the oscillations was observed. The chief proof that the friction is independent of the pressure is that no difference of resistance has been observed in water mains and in other cases, where water flows over solid surfaces under widely different pressures.

2. The frictional resistance of large surfaces is proportional to the area of the surface.

3. At low velocities of not more than 1 in. per second for water, the frictional resistance increases directly as the relative velocity of the fluid and the surface against which it flows. At velocities of 1/2 ft. per second and greater velocities, the frictional resistance is more nearly proportional to the square of the relative velocity.

In many treatises on hydraulics it is stated that the frictional resistance is independent of the nature of the solid surface. The explanation of this was supposed to be that a film of fluid remained attached to the solid surface, the resistance being generated between this fluid layer and layers more distant from the surface. At extremely low velocities the solid surface does not seem to have much influence on the friction. In Coulomb’s experiments a metal surface covered with tallow, and oscillated in water, had exactly the same resistance as a clean metal surface, and when sand was scattered over the tallow the resistance was only very slightly increased. The earlier calculations of the resistance of water at higher velocities in iron and wood pipes and earthen channels seemed to give a similar result. These, however, were erroneous, and it is now well understood that differences of roughness of the solid surface very greatly influence the friction, at such velocities as are common in engineering practice. H. P. G. Darcy’s experiments, for instance, showed that in old and incrusted water mains the resistance was twice or sometimes thrice as great as in new and clean mains.

§ 66. Ordinary Expressions for Fluid Friction at Velocities not Extremely Small.—Let f be the frictional resistance estimated in pounds per square foot of surface at a velocity of 1 ft. per second; ω the area of the surface in square feet; and v its velocity in feet per second relatively to the water in which it is immersed. Then, in accordance with the laws stated above, the total resistance of the surface is

R = fωv2

(1)

where f is a quantity approximately constant for any given surface. If

ξ = 2gf/G,

R = ξGωv2/2g,

(2)

where ξ is, like f, nearly constant for a given surface, and is termed the coefficient of friction.

The following are average values of the coefficient of friction for water, obtained from experiments on large plane surfaces, moved in an indefinitely large mass of water.

	Coefficient  of Friction, ξ	Frictional  Resistance in  ℔ per sq. ft. f
New well-painted iron plate	.00489	.00473
Painted and planed plank (Beaufoy)	.00350	.00339
Surface of iron ships (Rankine)	.00362	.00351
Varnished surface (Froude)	.00258	.00250
Fine sand surface (Froude)	.00418	.00405
Coarser sand surface (Froude)	.00503	.00488

The distance through which the frictional resistance is overcome is v ft. per second. The work expended in fluid friction is therefore given by the equation—

Work expended = fωv3 foot-pounds per second = ξGωv3/2g    〃    〃$\scriptstyle {\left.{\begin{matrix}\ \\\ \end{matrix}}\right\}\,}$

(3).

The coefficient of friction and the friction per square foot of surface can be indirectly obtained from observations of the discharge of pipes and canals. In obtaining them, however, some assumptions as to the motion of the water must be made, and it will be better therefore to discuss these values in connexion with the cases to which they are related.

Many attempts have been made to express the coefficient of friction in a form applicable to low as well as high velocities. The older hydraulic writers considered the resistance termed fluid friction to be made up of two parts,—a part due directly to the distortion of the mass of water and proportional to the velocity of the water relatively to the solid surface, and another part due to kinetic energy imparted to the water striking the roughnesses of the solid surface and proportional to the square of the velocity. Hence they proposed to take

ξ = α + β/v

in which expression the second term is of greatest importance at very low velocities, and of comparatively little importance at velocities over about 1/2 ft. per second. Values of ξ expressed in this and similar forms will be given in connexion with pipes and canals.

All these expressions must at present be regarded as merely empirical expressions serving practical purposes.

The frictional resistance will be seen to vary through wider limits than these expressions allow, and to depend on circumstances of which they do not take account.

§ 67. Coulomb’s Experiments.—The first direct experiments on fluid friction were made by Coulomb, who employed a circular disk suspended by a thin brass wire and oscillated in its own plane. His experiments were chiefly made at very low velocities. When the disk is rotated to any given angle, it oscillates under the action of its inertia and the torsion of the wire. The oscillations diminish gradually in consequence of the work done in overcoming the friction of the disk. The diminution furnishes a means of determining the friction.

 

Fig. 78.

Fig. 78 shows Coulomb’s apparatus. LK supports the wire and disk: ag is the brass wire, the torsion of which causes the oscillations; DS is a graduated disk serving to measure the angles through which the apparatus oscillates. To this the friction disk is rigidly attached hanging in a vessel of water. The friction disks were from 4.7 to 7.7 in. diameter, and they generally made one oscillation in from 20 to 30 seconds, through angles varying from 360° to 6°. When the velocity of the circumference of the disk was less than 6 in. per second, the resistance was sensibly proportional to the velocity.

Beaufoy’s Experiments.—Towards the end of the 18th century Colonel Mark Beaufoy (1764–1827) made an immense mass of experiments on the resistance of bodies moved through water (Nautical and Hydraulic Experiments, London, 1834). Of these the only ones directly bearing on surface friction were some made in 1796 and 1798. Smooth painted planks were drawn through water and the resistance measured. For two planks differing in area by 46 sq. ft., at a velocity of 10 ft. per second, the difference of resistance, measured on the difference of area, was 0.339 ℔ per square foot. Also the resistance varied as the 1.949th power of the velocity.

§ 68. Froude’s Experiments.—The most important direct experiments on fluid friction at ordinary velocities are those made by William Froude (1810–1879) at Torquay. The method adopted in these experiments was to tow a board in a still water canal, the velocity and the resistance being registered by very ingenious recording arrangements. The general arrangement of the apparatus is shown in fig. 79. AA is the board the resistance of which is to be determined. B is a cutwater giving a fine entrance to the plane surfaces of the board. CC is a bar to which the board AA is attached, and which is suspended by a parallel motion from a carriage running on rails above the still water canal. G is a link by which the resistance of the board is transmitted to a spiral spring H. A bar I rigidly connects the other end of the spring to the carriage. The dotted lines K, L indicate the position of a couple of levers by which the extension of the spring is caused to move a pen M, which records the extension on a greatly increased scale, by a line drawn on the paper cylinder N. This cylinder revolves at a speed proportionate to that of the carriage, its motion being obtained from the axle of the carriage wheels. A second pen O, receiving jerks at every second and a quarter from a clock P, records time on the paper cylinder. The scale for the line of resistance is ascertained by stretching the spiral spring by known weights. The boards used for the experiment were 3/16 in. thick, 19 in. deep, and from 1 to 50 ft. in length, cutwater included. A lead keel counteracted the buoyancy of the board. The boards were covered with various substances, such as paint, varnish, Hay’s composition, tinfoil, &c., so as to try the effect of different degrees of roughness of surface. The results obtained by Froude may be summarized as follows:—

Fig. 79.

1. The friction per square foot of surface varies very greatly for different surfaces, being generally greater as the sensible roughness of the surface is greater. Thus, when the surface of the board was covered as mentioned below, the resistance for boards 50 ft. long, at 10 ft. per second, was—

Tinfoil or varnish	0.25	℔ per sq. ft.
Calico	0.47	〃   〃
Fine sand	0.405	〃   〃
Coarser sand	0.488	〃   〃

2. The power of the velocity to which the friction is proportional varies for different surfaces. Thus, with short boards 2 ft. long,

For tinfoil the resistance varied as v^2.16.
For other surfaces the resistance varied as v^2.00.

With boards 50 ft. long,

For varnish or tinfoil the resistance varied as v^1.83.
For sand the resistance varied as v^2.00.

3. The average resistance per square foot of surface was much greater for short than for long boards; or, what is the same thing, the resistance per square foot at the forward part of the board was greater than the friction per square foot of portions more sternward. Thus,

			Mean Resistance in ℔ per sq. ft.
Varnished surface	2	ft. long	0.41
	50	〃	0.25
Fine sand surface	2	〃	0.81
	50	〃	0.405

This remarkable result is explained thus by Froude: “The portion of surface that goes first in the line of motion, in experiencing resistance from the water, must in turn communicate motion to the water, in the direction in which it is itself travelling. Consequently the portion of surface which succeeds the first will be rubbing, not against stationary water, but against water partially moving in its own direction, and cannot therefore experience so much resistance from it.”

§ 69. The following table gives a general statement of Froude’s results. In all the experiments in this table, the boards had a fine cutwater and a fine stern end or run, so that the resistance was entirely due to the surface. The table gives the resistances per square foot in pounds, at the standard speed of 600 feet per minute, and the power of the speed to which the friction is proportional, so that the resistance at other speeds is easily calculated.

	Length of Surface, or Distance from Cutwater, in feet.
	2 ft.			8 ft.			20 ft.			50 ft.
	A	B	C	A	B	C	A	B	C	A	B	C
Varnish	2.00	.41	.390	1.85	.325	.264	1.85	.278	.240	1.83	.250	.226
Paraffin	..	.38	.370	1.94	.314	.260	1.93	.271	.237	..	..	..
Tinfoil	2.16	.30	.295	1.99	.278	.263	1.90	.262	.244	1.83	.246	.232
Calico	1.93	.87	.725	1.92	.626	.504	1.89	.531	.447	1.87	.474	.423
Fine sand	2.00	.81	.690	2.00	.583	.450	2.00	.480	.384	2.06	.405	.337
Medium sand	2.00	.90	.730	2.00	.625	.488	2.00	.534	.465	2.00	.488	.456
Coarse sand	2.00	1.10	.880	2.00	.714	.520	2.00	.588	.490	..	..	..

Columns A give the power of the speed to which the resistance is approximately proportional.

Columns B give the mean resistance per square foot of the whole surface of a board of the lengths stated in the table.

Columns C give the resistance in pounds of a square foot of surface at the distance sternward from the cutwater stated in the heading.

Although these experiments do not directly deal with surfaces of greater length than 50 ft., they indicate what would be the resistances of longer surfaces. For at 50 ft. the decrease of resistance for an increase of length is so small that it will make no very great difference in the estimate of the friction whether we suppose it to continue to diminish at the same rate or not to diminish at all. For a varnished surface the friction at 10 ft. per second diminishes from 0.41 to 0.32 ℔ per square foot when the length is increased from 2 to 8 ft., but it only diminishes from 0.278 to 0.250 ℔ per square foot for an increase from 20 ft. to 50 ft.

If the decrease of friction sternwards is due to the generation of a current accompanying the moving plane, there is not at first sight any reason why the decrease should not be greater than that shown by the experiments. The current accompanying the board might be assumed to gain in volume and velocity sternwards, till the velocity was nearly the same as that of the moving plane and the friction per square foot nearly zero. That this does not happen appears to be due to the mixing up of the current with the still water surrounding it. Part of the water in contact with the board at any point, and receiving energy of motion from it, passes afterwards to distant regions of still water, and portions of still water are fed in towards the board to take its place. In the forward part of the board more kinetic energy is given to the current than is diffused into surrounding space, and the current gains in velocity. At a greater distance back there is an approximate balance between the energy communicated to the water and that diffused. The velocity of the current accompanying the board becomes constant or nearly constant, and the friction per square foot is therefore nearly constant also.

§ 70. Friction of Rotating Disks.—A rotating disk is virtually a surface of unlimited extent and it is convenient for experiments on friction with different surfaces at different speeds. Experiments carried out by Professor W. C. Unwin (Proc. Inst. Civ. Eng. lxxx.) are useful both as illustrating the laws of fluid friction and as giving data for calculating the resistance of the disks of turbines and centrifugal pumps. Disks of 10, 15 and 20 in. diameter fixed on a vertical shaft were rotated by a belt driven by an engine. They were enclosed in a cistern of water between parallel top and bottom fixed surfaces. The cistern was suspended by three fine wires. The friction of the disk is equal to the tendency of the cistern to rotate, and this was measured by balancing the cistern by a fine silk cord passing over a pulley and carrying a scale pan in which weights could be placed.

If ω is an element of area on the disk moving with the velocity v, the friction on this element is fωvⁿ, where f and n are constant for any given kind of surface. Let α be the angular velocity of rotation, R the radius of the disk. Consider a ring of the surface between r and r + dr. Its area is 2πrdr, its velocity αr and the friction of this ring is f 2πrdrαⁿr ⁿ. The moment of the friction about the axis of rotation is 2παⁿfr ⁿ⁺² dr, and the total moment of friction for the two sides of the disk is

M = 4πα_nf $\int _{0}^{\text{R}}$ r ⁿ⁺² dr = {4παⁿ/(n + 3) } f Rⁿ⁺³.

If N is the number of revolutions per sec.,

M = {2ⁿ⁺² πⁿ⁺¹ Nⁿ/(n + 3) } f Rⁿ⁺³,

and the work expended in rotating the disk is

Mα = {2ⁿ⁺³ πⁿ⁺² Nⁿ⁺¹/(n + 3) } f Rⁿ⁺³ foot ℔ per sec.

The experiments give directly the values of M for the disks corresponding to any speed N. From these the values of f and n can be deduced, f being the friction per square foot at unit velocity. For comparison with Froude’s results it is convenient to calculate the resistance at 10 ft. per second, which is F = f10ⁿ.

The disks were rotated in chambers 22 in. diameter and 3, 6 and 12 in. deep. In all cases the friction of the disks increased a little as the chamber was made larger. This is probably due to the stilling of the eddies against the surface of the chamber and the feeding back of the stilled water to the disk. Hence the friction depends not only on the surface of the disk but to some extent on the surface of the chamber in which it rotates. If the surface of the chamber is made rougher by covering with coarse sand there is also an increase of resistance.

For the smoother surfaces the friction varied as the 1.85th power of the velocity. For the rougher surfaces the power of the velocity to which the resistance was proportional varied from 1.9 to 2.1. This is in agreement with Froude’s results.

Experiments with a bright brass disk showed that the friction decreased with increase of temperature. The diminution between 41° and 130° F. amounted to 18%. In the general equation M = cNⁿ for any given disk,

c_t = 0.1328 (1 − 0.0021t),

where c_t is the value of c for a bright brass disk 0.85 ft. in diameter at a temperature t° F.

The disks used were either polished or made rougher by varnish or by varnish and sand. The following table gives a comparison of the results obtained with the disks and Froude’s results on planks 50 ft. long. The values given are the resistances per square foot at 10 ft. per sec.

Froude’s Experiments.		Disk Experiments.
Tinfoil surface	0.232	Bright brass	0.202 to 0.229
Varnish	0.226	Varnish	0.220 to 0.233
Fine sand	0.337	Fine sand	0.339
Medium sand	0.456	Very coarse sand	0.587 to 0.715

VIII. STEADY FLOW OF WATER IN PIPES OF UNIFORM SECTION.

§ 71. The ordinary theory of the flow of water in pipes, on which all practical formulae are based, assumes that the variation of velocity at different points of any cross section may be neglected. The water is considered as moving in plane layers, which are driven through the pipe against the frictional resistance, by the difference of pressure at or elevation of the ends of the pipe. If the motion is steady the velocity at each cross section remains the same from moment to moment, and if the cross sectional area is constant the velocity at all sections must be the same. Hence the motion is uniform. The most important resistance to the motion of the water is the surface friction of the pipe, and it is convenient to estimate this independently of some smaller resistances which will be accounted for presently.

 

Fig. 80.

In any portion of a uniform pipe, excluding for the present the ends of the pipe, the water enters and leaves at the same velocity. For that portion therefore the work of the external forces and of the surface friction must be equal. Let fig. 80 represent a very short portion of the pipe, of length dl, between cross sections at z and z + dz ft. above any horizontal datum line xx, the pressures at the cross sections being p and p + dp ℔ per square foot. Further, let Q be the volume of flow or discharge of the pipe per second, Ω the area of a normal cross section, and χ the perimeter of the pipe. The Q cubic feet, which flow through the space considered per second, weigh GQ ℔, and fall through a height −dz ft. The work done by gravity is then

−GQ dz;

a positive quantity if dz is negative, and vice versa. The resultant pressure parallel to the axis of the pipe is p − (p + dp) = −dp ℔ per square foot of the cross section. The work of this pressure on the volume Q is

−Q dp.

The only remaining force doing work on the system is the friction against the surface of the pipe. The area of that surface is χdl.

The work expended in overcoming the frictional resistance per second is (see § 66, eq. 3)

−ζGχ dl v³/2g,

or, since Q = Ωv,

−ζG (χ/Ω) Q (v²/2g)dl;

the negative sign being taken because the work is done against a resistance. Adding all these portions of work, and equating the result to zero, since the motion is uniform,—

−GQ dz − Q dp − ζG (χ/Ω) Q (v ²/2g)dl = 0.

Dividing by GQ,

dz + dp/G + ζ (χ/Ω) (v ²/2g)dl = 0.

Integrating,

z + p/G + ζ (χ/Ω) (v 2/2g)l = constant.

(1)

§ 72. Let A and B (fig. 81) be any two sections of the pipe for which p, z, l have the values p₁, z₁, l₁, and p₂, z₂, l₂, respectively. Then

z₁ + p₁/G + ζ (χ/Ω) (v ²/2g)l₁ = z₂ + p₂/G + ζ (χ/Ω) (v ²/2g)l₂;

or, if l₂ − l₁ = L, rearranging the terms,

ζv 2/2g = (1/L) {(z1 + p1/G) − (z2 + p2/G)} Ω/χ.

(2)

Fig. 81.

Suppose pressure columns introduced at A and B. The water will rise in those columns to the heights p₁/G and p₂/G due to the pressures p₁ and p₂ at A and B. Hence (z₁ + p₁/G) − (z₂ + p₂/G) is the quantity represented in the figure by DE, the fall of level of the pressure columns, or virtual fall of the pipe. If there were no friction in the pipe, then by Bernoulli’s equation there would be no fall of level of the pressure columns, the velocity being the same at A and B. Hence DE or h is the head lost in friction in the distance AB. The quantity DE/AB = h/L is termed the virtual slope of the pipe or virtual fall per foot of length. It is sometimes termed very conveniently the relative fall. It will be denoted by the symbol i.

The quantity Ω/χ which appears in many hydraulic equations is called the hydraulic mean radius of the pipe. It will be denoted by m.

Introducing these values,

ζv 2/2g = mh/L = mi.

(3)

For pipes of circular section, and diameter d,

m = Ω/χ = 1/4πd ²/πd = 1/4d.

which shows that the head lost in friction is proportional to the head due to the velocity, and is found by multiplying that head by the coefficient 4ζL/d. It is assumed above that the atmospheric pressure at C and D is the same, and this is usually nearly the case. But if C and D are at greatly different levels the excess of barometric pressure at C, in feet of water, must be added to p₂/G.

§ 73. Hydraulic Gradient or Line of Virtual Slope.—Join CD. Since the head lost in friction is proportional to L, any intermediate pressure column between A and B will have its free surface on the line CD, and the vertical distance between CD and the pipe at any point measures the pressure, exclusive of atmospheric pressure, in the pipe at that point. If the pipe were laid along the line CD instead of AB, the water would flow at the same velocity by gravity without any change of pressure from section to section. Hence CD is termed the virtual slope or hydraulic gradient of the pipe. It is the line of free surface level for each point of the pipe.

If an ordinary pipe, connecting reservoirs open to the air, rises at any joint above the line of virtual slope, the pressure at that point is less than the atmospheric pressure transmitted through the pipe. At such a point there is a liability that air may be disengaged from the water, and the flow stopped or impeded by the accumulation of air. If the pipe rises more than 34 ft. above the line of virtual slope, the pressure is negative. But as this is impossible, the continuity of the flow will be broken.

If the pipe is not straight, the line of virtual slope becomes a curved line, but since in actual pipes the vertical alterations of level are generally small, compared with the length of the pipe, distances measured along the pipe are sensibly proportional to distances measured along the horizontal projection of the pipe. Hence the line of hydraulic gradient may be taken to be a straight line without error of practical importance.

Fig. 82.

§ 74. Case of a Uniform Pipe connecting two Reservoirs, when all the Resistances are taken into account.—Let h (fig. 82) be the difference of level of the reservoirs, and v the velocity, in a pipe of length L and diameter d. The whole work done per second is virtually the removal of Q cub. ft. of water from the surface of the upper reservoir to the surface of the lower reservoir, that is GQh foot-pounds. This is expended in three ways. (1) The head v ²/2g, corresponding to an expenditure of GQv ²/2g foot-pounds of work, is employed in giving energy of motion to the water. This is ultimately wasted in eddying motions in the lower reservoir. (2) A portion of head, which experience shows may be expressed in the form ζ₀v ²/2g, corresponding to an expenditure of GQζ₀v ²/2g foot-pounds of work, is employed in overcoming the resistance at the entrance to the pipe. (3) As already shown the head expended in overcoming the surface friction of the pipe is ζ(4L/d) (v ²/2g) corresponding to GQζ (4L/d) (v ²/2g) foot-pounds of work. Hence

GQh = GQv ²/2g + GQζ₀v ²/2g + GQζ·4L·v ²/d·2g;

h = (1 + ζ0 + ζ·4L/d)v2/2g.       v = 8.025 √ [hd / {(1 + ζ0)d + 4ζL}].$\scriptstyle {\left.{\begin{matrix}\ \\\ \end{matrix}}\right\}\,}$

(5)

If the pipe is bell-mouthed, ζ₀ is about = .08. If the entrance to the pipe is cylindrical, ζ₀ = 0.505. Hence 1 + ζ₀ = 1.08 to 1.505. In general this is so small compared with ζ4L/d that, for practical calculations, it may be neglected; that is, the losses of head other than the loss in surface friction are left out of the reckoning. It is only in short pipes and at high velocities that it is necessary to take account of the first two terms in the bracket, as well as the third. For instance, in pipes for the supply of turbines, v is usually limited to 2 ft. per second, and the pipe is bellmouthed. Then 1.08v ²/2g = 0.067 ft. In pipes for towns’ supply v may range from 2 to 41/2 ft. per second, and then 1.5v ²/2g = 0.1 to 0.5 ft. In either case this amount of head is small compared with the whole virtual fall in the cases which most commonly occur.

When d and v or d and h are given, the equations above are solved quite simply. When v and h are given and d is required, it is better to proceed by approximation. Find an approximate value of d by assuming a probable value for ζ as mentioned below. Then from that value of d find a corrected value for ζ and repeat the calculation.

The equation above may be put in the form

h = (4ζ/d) [{ (1 + ζ0)d/4ζ} + L] v 2/2g;

(6)

from which it is clear that the head expended at the mouthpiece is equivalent to that of a length

(1 + ζ₀)d/4ζ

of the pipe. Putting 1 + ζ₀ = 1.505 and ζ = 0.01, the length of pipe equivalent to the mouthpiece is 37.6d nearly. This may be added to the actual length of the pipe to allow for mouthpiece resistance in approximate calculations.

§ 75. Coefficient of Friction for Pipes discharging Water.—From the average of a large number of experiments, the value of ζ for ordinary iron pipes is

ζ = 0.007567.

(7)

But practical experience shows that no single value can be taken applicable to very different cases. The earlier hydraulicians occupied themselves chiefly with the dependence of ζ on the velocity. Having regard to the difference of the law of resistance at very low and at ordinary velocities, they assumed that ζ might be expressed in the form

ζ = a + β/v.

The following are the best numerical values obtained for ζ so expressed:—

	α	β
R. de Prony (from 51 experiments)	0.006836	0.001116
J. F. d’Aubuisson de Voisins	0.00673	0.001211
J. A. Eytelwein	0.005493	0.00143

Weisbach proposed the formula

4ζ = α + β/√v = 0.003598 + 0.004289/√v.

(8)

§ 76. Darcy’s Experiments on Friction in Pipes.—All previous experiments on the resistance of pipes were superseded by the remarkable researches carried out by H. P. G. Darcy (1803–1858), the Inspector-General of the Paris water works. His experiments were carried out on a scale, under a variation of conditions, and with a degree of accuracy which leaves little to be desired, and the results obtained are of very great practical importance. These results may be stated thus:—

1. For new and clean pipes the friction varies considerably with the nature and polish of the surface of the pipe. For clean cast iron it is about 11/2 times as great as for cast iron covered with pitch.

2. The nature of the surface has less influence when the pipes are old and incrusted with deposits, due to the action of the water. Thus old and incrusted pipes give twice as great a frictional resistance as new and clean pipes. Darcy’s coefficients were chiefly determined from experiments on new pipes. He doubles these coefficients for old and incrusted pipes, in accordance with the results of a very limited number of experiments on pipes containing incrustations and deposits.

3. The coefficient of friction may be expressed in the form ζ = α + β/v; but in pipes which have been some time in use it is sufficiently accurate to take ζ = α₁ simply, where α₁ depends on the diameter of the pipe alone, but α and β on the other hand depend both on the diameter of the pipe and the nature of its surface. The following are the values of the constants.

For pipes which have been some time in use, neglecting the term depending on the velocity;

ζ = α (1 + β/d).

(9)

	α	β
For drawn wrought-iron or smooth cast-iron pipes	.004973	.084
For pipes altered by light incrustations	.00996	.084

These coefficients may be put in the following very simple form, without sensibly altering their value:—

For clean pipes	ζ = .005 (1 + 1/12d)	$\scriptstyle {\left.{\begin{matrix}\ \\\ \end{matrix}}\right\}\,}$
For slightly incrusted pipes	ζ = .01 (1 + 1/12d)

(9a)

Darcy’s Value of the Coefficient of Friction ζ for Velocities not less
than 4 in. per second.

Diameter of Pipe  in Inches.	ζ		Diameter of Pipe  in Inches.	ζ
	New Pipes.	Incrusted  Pipes.		New Pipes.	Incrusted  Pipes.
2	0.00750	0.01500	18	.00528	.01056
3	.00667	.01333	21	.00524	.01048
4	.00625	.01250	24	.00521	.01042
5	.00600	.01200	27	.00519	.01037
6	.00583	.01167	30	.00517	.01033
7	.00571	.01143	36	.00514	.01028
8	.00563	.01125	42	.00512	.01024
9	.00556	.01111	48	.00510	.01021
12	.00542	.01083	54	.00509	.01019
15	.00533	.01067

These values of ζ are, however, not exact for widely differing velocities. To embrace all cases Darcy proposed the expression

ζ = (α + α₁/d) + (β + β₁/d ²) /v ;

(10)

which is a modification of Coulomb’s, including terms expressing the influence of the diameter and of the velocity. For clean pipes Darcy found these values

α	= .004346
α1	= .0003992
β	= .0010182
β1	= .000005205.

It has become not uncommon to calculate the discharge of pipes by the formula of E. Ganguillet and W. R. Kutter, which will be discussed under the head of channels. For the value of c in the relation v = c √(mi), Ganguillet and Kutter take

c =	41.6 + 1.811/n + .00281/i
	1 + [ (41.6 + .00281/i) (n/ √m) ]

where n is a coefficient depending only on the roughness of the pipe. For pipes uncoated as ordinarily laid n = 0.013. The formula is very cumbrous, its form is not rationally justifiable and it is not at all clear that it gives more accurate values of the discharge than simpler formulae.

§ 77. Later Investigations on Flow in Pipes.—The foregoing statement gives the theory of flow in pipes so far as it can be put in a simple rational form. But the conditions of flow are really more complicated than can be expressed in any rational form. Taking even selected experiments the values of the empirical coefficient ζ range from 0.16 to 0.0028 in different cases. Hence means of discriminating the probable value of ζ are necessary in using the equations for practical purposes. To a certain extent the knowledge that ζ decreases with the size of the pipe and increases very much with the roughness of its surface is a guide, and Darcy’s method of dealing with these causes of variation is very helpful. But a further difficulty arises from the discordance of the results of different experiments. For instance F. P. Stearns and J. M. Gale both experimented on clean asphalted cast-iron pipes, 4 ft. in diameter. According to one set of gaugings ζ = .0051, and according to the other ζ = .0031. It is impossible in such cases not to suspect some error in the observations or some difference in the condition of the pipes not noticed by the observers.

It is not likely that any formula can be found which will give exactly the discharge of any given pipe. For one of the chief factors in any such formula must express the exact roughness of the pipe surface, and there is no scientific measure of roughness. The most that can be done is to limit the choice of the coefficient for a pipe within certain comparatively narrow limits. The experiments on fluid friction show that the power of the velocity to which the resistance is proportional is not exactly the square. Also in determining the form of his equation for ζ Darcy used only eight out of his seventeen series of experiments, and there is reason to think that some of these were exceptional. Barré de Saint-Venant was the first to propose a formula with two constants,

dh/4l = mVⁿ,

where m and n are experimental constants. If this is written in the form

log m + n log v = log (dh/4l),

we have, as Saint-Venant pointed out, the equation to a straight line, of which m is the ordinate at the origin and n the ratio of the slope. If a series of experimental values are plotted logarithmically the determination of the constants is reduced to finding the straight line which most nearly passes through the plotted points. Saint-Venant found for n the value of 1.71. In a memoir on the influence of temperature on the movement of water in pipes (Berlin, 1854) by G. H. L. Hagen (1797–1884) another modification of the Saint-Venant formula was given. This is h/l = mv ⁿ/d ^x, which involves three experimental coefficients. Hagen found n = 1.75; x = 1.25; and m was then nearly independent of variations of v and d. But the range of cases examined was small. In a remarkable paper in the Trans. Roy. Soc., 1883, Professor Osborne Reynolds made much clearer the change from regular stream line motion at low velocities to the eddying motion, which occurs in almost all the cases with which the engineer has to deal. Partly by reasoning, partly by induction from the form of logarithmically plotted curves of experimental results, he arrived at the general equation h/l = c (vⁿ/d ³⁻ⁿ) P²⁻ⁿ, where n = l for low velocities and n = 1.7 to 2 for ordinary velocities. P is a function of the temperature. Neglecting variations of temperature Reynold’s formula is identical with Hagen’s if x = 3−n. For practical purposes Hagen’s form is the more convenient.

Values of Index of Velocity.

Surface of Pipe.	Authority.	Diameter of Pipe  in Metres.	Values of n.
Tin plate	Bossut	.036	1.697	$\scriptstyle {\left.{\begin{matrix}\ \\\ \end{matrix}}\right\}\,}$1.72
		.054	1.730
Wrought iron (gas pipe)	Hamilton Smith	.0159	1.756	$\scriptstyle {\left.{\begin{matrix}\ \\\ \end{matrix}}\right\}\,}$1.75
		.0267	1.770
Lead	Darcy	.014	1.866	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \ \end{matrix}}\right\}\,}$1.77
		.027	1.755
		.041	1.760
Clean brass	Mair	.036	1.795	1.795
Asphalted	Hamilton Smith	.0266	1.760	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \\\ \ \end{matrix}}\right\}\,}$1.85
	Lampe.	.4185	1.850
	W. W. Bonn	.306	1.582
	Stearns	1.219	1.880
Riveted wrought iron	Hamilton Smith	.2776	1.804	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \ \end{matrix}}\right\}\,}$1.87
		.3219	1.892
		.3749	1.852
Wrought iron (gas pipe)	Darcy	.0122	1.900	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \ \end{matrix}}\right\}\,}$1.87
		.0266	1.899
		.0395	1.838
New cast iron	Darcy	.0819	1.950	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \\\ \ \end{matrix}}\right\}\,}$1.95
		.137	1.923
		.188	1.957
		.50	1.950
Cleaned cast iron	Darcy	.0364	1.835	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \\\ \ \end{matrix}}\right\}\,}$2.00
		.0801	2.000
		.2447	2.000
		.397	2.07
Incrusted cast iron	Darcy	.0359	1.980	$\scriptstyle {\left.{\begin{matrix}\ \\\\\ \ \end{matrix}}\right\}\,}$2.00
		.0795	1.990
		.2432	1.990

Fig. 83.

In 1886, Professor W. C. Unwin plotted logarithmically all the most trustworthy experiments on flow in pipes then available.^[5] Fig. 83 gives one such plotting. The results of measuring the slopes of the lines drawn through the plotted points are given in the table.

It will be seen that the values of the index n range from 1.72 for the smoothest and cleanest surface, to 2.00 for the roughest. The numbers after the brackets are rounded off numbers.

The value of n having been thus determined, values of m/d^x were next found and averaged for each pipe. These were again plotted logarithmically in order to find a value for x. The lines were not very regular, but in all cases the slope was greater than 1 to 1, so that the value of x must be greater than unity. The following table gives the results and a comparison of the value of x and Reynolds’s value 3−n.

Kind of Pipe.	n	3−n	x
Tin plate	1.72	1.28	1.100
Wrought iron (Smith)	1.75	1.25	1.210
Asphalted pipes	1.85	1.15	1.127
Wrought iron (Darcy)	1.87	1.13	1.680
Riveted wrought iron	1.87	1.13	1.390
New cast iron	1.95	1.05	1.168
Cleaned cast iron	2.00	1.00	1.168
Incrusted cast iron	2.00	1.00	1.160

With the exception of the anomalous values for Darcy’s wrought-iron pipes, there is no great discrepancy between the values of x and 3−n, but there is no appearance of relation in the two quantities. For the present it appears preferable to assume that x is independent of n.

It is now possible to obtain values of the third constant m, using the values found for n and x. The following table gives the results, the values of m being for metric measures.

Here, considering the great range of diameters and velocities in the experiments, the constancy of m is very satisfactorily close. The asphalted pipes give rather variable values. But, as some of these were new and some old, the variation is, perhaps, not surprising. The incrusted pipes give a value of m quite double that for new pipes but that is perfectly consistent with what is known of fluid friction in other cases.

Kind of Pipe.	Diameter  in Metres.	Value of m.	Mean Value of m.	Authority.
Tin plate	0.036	.01697	.01686	Bossut
	0.054	.01676
Wrought iron	0.016	.01302	.01310	Hamilton Smith
	0.027	.01319
Asphalted pipes	0.027	.01749	.01831	Hamilton Smith
	0.306	.02058		W. W. Bonn
	0.306	.02107		W. W. Bonn
	0.419	.01650		Lampe
	1.219	.01317		Stearns
	1.219	.02107		Gale
Riveted wrought iron	0.278	.01370	.01403	Hamilton Smith
	0.322	.01440
	0.375	.01390
	0.432	.01368
	0.657	.01448
New cast iron	0.082	.01725	.01658	Darcy
	0.137	.01427
	0.188	.01734
	0.500	.01745
Cleaned cast iron	0.080	.01979	.01994	Darcy
	0.245	.02091
	0.297	.01913
Incrusted cast iron	0.036	.03693	.03643	Darcy
	0.080	.03530
	0.243	.03706

General Mean Values of Constants.

The general formula (Hagen’s)—h/l = mvⁿ/d^x·2g—can therefore be taken to fit the results with convenient closeness, if the following mean values of the coefficients are taken, the unit being a metre:—

Kind of Pipe.	m	x	n
Tin plate	.0169	1.10	1.72
Wrought iron	.0131	1.21	1.75
Asphalted iron	.0183	1.127	1.85
Riveted wrought iron	.0140	1.390	1.87
New cast iron	.0166	1.168	1.95
Cleaned cast iron	.0199	1.168	2.0
Incrusted cast iron	.0364	1.160	2.0

The variation of each of these coefficients is within a comparatively narrow range, and the selection of the proper coefficient for any given case presents no difficulty, if the character of the surface of the pipe is known.

It only remains to give the values of these coefficients when the quantities are expressed in English feet. For English measures the following are the values of the coefficients:—

Kind of Pipe.	m	x	n
Tin plate	.0265	1.10	1.72
Wrought iron	.0226	1.21	1.75
Asphalted iron	.0254	1.127	1.85
Riveted wrought iron	.0260	1.390	1.87
New cast iron	.0215	1.168	1.95
Cleaned cast iron	.0243	1.168	2.0
Incrusted cast iron	.0440	1.160	2.0

§ 78. Distribution of Velocity in the Cross Section of a Pipe.—Darcy made experiments with a Pitot tube in 1850 on the velocity at different points in the cross section of a pipe. He deduced the relation

V − v = 11.3 (r^3/2/R) √i,

where V is the velocity at the centre and v the velocity at radius r in a pipe of radius R with a hydraulic gradient i. Later Bazin repeated the experiments and extended them (Mém. de l’Académie des Sciences, xxxii. No. 6). The most important result was the ratio of mean to central velocity. Let b = Ri/U², where U is the mean velocity in the pipe; then V/U = 1 + 9.03 √b. A very useful result for practical purposes is that at 0.74 of the radius of the pipe the velocity is equal to the mean velocity. Fig. 84 gives the velocities at different radii as determined by Bazin.

Fig. 84.

§ 79. Influence of Temperature on the Flow through Pipes.—Very careful experiments on the flow through a pipe 0.1236 ft. in diameter and 25 ft. long, with water at different temperatures, have been made by J. G. Mair (Proc. Inst. Civ. Eng. lxxxiv.). The loss of head was measured from a point 1 ft. from the inlet, so that the loss at entry was eliminated. The 11/2 in. pipe was made smooth inside and to gauge, by drawing a mandril through it. Plotting the results logarithmically, it was found that the resistance for all temperatures varied very exactly as v1.795, the index being less than 2 as in other experiments with very smooth surfaces. Taking the ordinary equation of flow h = ζ (4L/D) (v²/2g), then for heads varying from 1 ft. to nearly 4 ft., and velocities in the pipe varying from 4 ft. to 9 ft. per second, the values of ζ were as follows:—

Temp. F.	ζ	Temp. F.	ζ
57	.0044 to .0052	100	.0039 to .0042
70	.0042 to .0045	110	.0037 to .0041
80	.0041 to .0045	120	.0037 to .0041
90	.0040 to .0045	130	.0035 to .0039
		160	.0035 to .0038

This shows a marked decrease of resistance as the temperature rises. If Professor Osborne Reynolds’s equation is assumed h = mLVⁿ/d³⁻ⁿ, and n is taken 1.795, then values of m at each temperature are practically constant—

Temp. F.	m.	Temp. F.	m.
57	0.000276	100	0.000244
70	0.000263	110	0.000235
80	0.000257	120	0.000229
90	0.000250	130	0.000225
		160	0.000206

where again a regular decrease of the coefficient occurs as the temperature rises. In experiments on the friction of disks at different temperatures Professor W. C. Unwin found that the resistance was proportional to constant × (1 − 0.0021t) and the values of m given above are expressed almost exactly by the relation

m = 0.000311 (1 − 0.00215 t).

In tank experiments on ship models for small ordinary variations of temperature, it is usual to allow a decrease of 3% of resistance for 10° F. increase of temperature.

§ 80. Influence of Deposits in Pipes on the Discharge. Scraping Water Mains.—The influence of the condition of the surface of a pipe on the friction is shown by various facts known to the engineers of waterworks. In pipes which convey certain kinds of water, oxidation proceeds rapidly and the discharge is considerably diminished. A main laid at Torquay in 1858, 14 m. in length, consists of 10-in., 9-in. and 8-in. pipes. It was not protected from corrosion by any coating. But it was found to the surprise of the engineer that in eight years the discharge had diminished to 51% of the original discharge. J. G. Appold suggested an apparatus for scraping the interior of the pipe, and this was constructed and used under the direction of William Froude (see “Incrustation of Iron Pipes,” by W. Ingham, Proc. Inst. Mech. Eng., 1899). It was found that by scraping the interior of the pipe the discharge was increased 56%. The scraping requires to be repeated at intervals. After each scraping the discharge diminishes rather rapidly to 10% and afterwards more slowly, the diminution in a year being about 25%.

Fig. 85 shows a scraper for water mains, similar to Appold’s but modified in details, as constructed by the Glenfield Company, at Kilmarnock. A is a longitudinal section of the pipe, showing the scraper in place; B is an end view of the plungers, and C, D sections of the boxes placed at intervals on the main for introducing or withdrawing the scraper. The apparatus consists of two plungers, packed with leather so as to fit the main pretty closely. On the spindle of these plungers are fixed eight steel scraping blades, with curved scraping edges fitting the surface of the main. The apparatus is placed in the main by removing the cover from one of the boxes shown at C, D. The cover is then replaced, water pressure is admitted behind the plungers, and the apparatus driven through the main. At Lancaster after twice scraping the discharge was increased 561/2%, at Oswestry 541/2%. The increased discharge is due to the diminution of the friction of the pipe by removing the roughnesses due to oxidation. The scraper can be easily followed when the mains are about 3 ft. deep by the noise it makes. The average speed of the scraper at Torquay is 21/3 m. per hour. At Torquay 49% of the deposit is iron rust, the rest being silica, lime and organic matter.

Fig. 85. Scale 1/25.

In the opinion of some engineers it is inadvisable to use the scraper. The incrustation is only temporarily removed, and if the use of the scraper is continued the life of the pipe is reduced. The only treatment effective in preventing or retarding the incrustation due to corrosion is to coat the pipes when hot with a smooth and perfect layer of pitch. With certain waters such as those derived from the chalk the incrustation is of a different character, consisting of nearly pure calcium carbonate. A deposit of another character which has led to trouble in some mains is a black slime containing a good deal of iron not derived from the pipes. It appears to be an organic growth. Filtration of the water appears to prevent the growth of the slime, and its temporary removal may be effected by a kind of brush scraper devised by G. F. Deacon (see “Deposits in Pipes,” by Professor J. C. Campbell Brown, Proc. Inst. Civ. Eng., 1903–1904).

§ 81. Flow of Water through Fire Hose.—The hose pipes used for fire purposes are of very varied character, and the roughness of the surface varies. Very careful experiments have been made by J. R. Freeman (Am. Soc. Civ. Eng. xxi., 1889). It was noted that under pressure the diameter of the hose increased sufficiently to have a marked influence on the discharge. In reducing the results the true diameter has been taken. Let v = mean velocity in ft. per sec.; r = hydraulic mean radius or one-fourth the diameter in feet; i = hydraulic gradient. Then v = n √(ri).

	Diameter  in Inches.	Gallons (United States)  per min.	i	v	n
Solid rubber hose $\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	2.65	215	.1863	12.50	123.3
	”	344	.4714	20.00	124.0
Woven cotton, rubber lined $\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	2.47	200	.2464	13.40	119.1
	”	299	.5269	20.00	121.5
Woven cotton, rubber lined $\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	2.49	200	.2427	13.20	117.7
	”	319	.5708	21.00	122.1
Knit cotton, rubber lined$\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	2.68	132	.0809	7.50	111.6
	”	299	.3931	17.00	114.8
Knit cotton, rubber lined$\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	2.69	204	.2357	11.50	100.1
	”	319	.5165	18.00	105.8
Woven cotton, rubber lined $\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	2.12	154	.3448	14.00	113.4
	”	240	.7673	21.81	118.4
Woven cotton, rubber lined $\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	2.53	54.8	.0261	3.50	94.3
	”	298	.8264	19.00	91.0
Unlined linen hose$\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	2.60	57.9	.0414	3.50	73.9
	”	331	1.1624	20.00	79.6

§ 82. Reduction of a Long Pipe of Varying Diameter to an Equivalent Pipe of Uniform Diameter. Dupuit’s Equation.—Water mains for the supply of towns often consist of a series of lengths, the diameter being the same for each length, but differing from length to length. In approximate calculations of the head lost in such mains, it is generally accurate enough to neglect the smaller losses of head and to have regard to the pipe friction only, and then the calculations may be facilitated by reducing the main to a main of uniform diameter, in which there would be the same loss of head. Such a uniform main will be termed an equivalent main.

Fig. 86.

In fig. 86 let A be the main of variable diameter, and B the equivalent uniform main. In the given main of variable diameter A, let

l1, l2	be the lengths,
d1, d2	the diameters,
v1, v2	the velocities,
i1, i2	the slopes,

for the successive portions, and let l, d, v and i be corresponding quantities for the equivalent uniform main B. The total loss of head in A due to friction is

h = i₁l₁ + i₂l₂ + . . .
= ζ (v₁² · 4l₁/2gd₁) + ζ (v₂² · 4l₂/2gd₂) + . . .

and in the uniform main

il = ζ (v² · 4l/2gd).

If the mains are equivalent, as defined above,

ζ (v² · 4l/2gd) = ζ (v₁² · 4l₁/2gd₁) + ζ (v₂² · 4l₂/2gd₂) + . . .

But, since the discharge is the same for all portions,

1/4πd ²v = 1/4πd₁²v₁ = 1/4πd₂²v₂ = . . .
v₁ = vd ²/d₁²; v₂ = vd ²/d₂² . . .

Also suppose that ζ may be treated as constant for all the pipes. Then

l/d = (d ⁴/d₁⁴) (l₁/d₁) + (d ⁴/d₂⁴) (l₂/d₂) + . . .
l = (d ⁵/d₁⁵) l₁ + (d ⁵/d₂⁵) l₂ + . . .

which gives the length of the equivalent uniform main which would have the same total loss of head for any given discharge.

 

Fig. 87.

§ 83. Other Losses of Head in Pipes.—Most of the losses of head in pipes, other than that due to surface friction against the pipe, are due to abrupt changes in the velocity of the stream producing eddies. The kinetic energy of these is deducted from the general energy of translation, and practically wasted.

Sudden Enlargement of Section.—Suppose a pipe enlarges in section from an area ω₀ to an area ω₁ (fig. 87); then

v₁/v₀ = ω₀/ω₁;

or, if the section is circular,

v₁/v₀ = (d₀/d₁)².

The head lost at the abrupt change of velocity has already been shown to be the head due to the relative velocity of the two parts of the stream. Hence head lost

ɧ_e = (v₀ − v₁)²/2g = (ω₁/ω₀ − 1)² v₁²/2g = {(d₁/d₀)² − 1}² v₁²/2g

if ζ_e is put for the expression in brackets.

ω1/ω0 =	1.1	1.2	1.5	1.7	1.8	1.9	2.0	2.5	3.0	3.5	4.0	5.0	6.0	7.0	8.0
d1/d0 =	1.05	1.10	1.22	1.30	1.34	1.38	1.41	1.58	1.73	1.87	2.00	2.24	2.45	2.65	2.83
ζe =	.01	.04	.25	.49	.64	.81	1.00	2.25	4.00	6.25	9.00	16.00	25.00	36.0	49.0

Fig. 88.	Fig. 89.

Abrupt Contraction of Section.—When water passes from a larger to a smaller section, as in figs. 88, 89, a contraction is formed, and the contracted stream abruptly expands to fill the section of the pipe. Let ω be the section and v the velocity of the stream at bb. At aa the section will be c_cω, and the velocity (ω/c_cω) v = v/c₁, where c_c is the coefficient of contraction. Then the head lost is

ɧ_m = (v/c_c − v)² / 2g = (1/c_c − 1)² v²/2g;

and, if c_c is taken 0.64,

ɧ_m = 0.316 v²/2g.

(2)

The value of the coefficient of contraction for this case is, however, not well ascertained, and the result is somewhat modified by friction. For water entering a cylindrical, not bell-mouthed, pipe from a reservoir of indefinitely large size, experiment gives

ɧ_a = 0.505 v²/2g.

(3)

If there is a diaphragm at the mouth of the pipe as in fig. 89, let ω₁ be the area of this orifice. Then the area of the contracted stream is c_cω₁, and the head lost is

ɧ_c = {(ω/c_cω₁) − 1}² v²/2g
  = ζ_cv² / 2g

(4)

if ζ, is put for {(ω/c_cω₁) − 1}². Weisbach has found experimentally the following values of the coefficient, when the stream approaching the orifice was considerably larger than the orifice:—

ω1/ω =	0.1	0.2	0.3	0.4	0.5	0.6	0.7	0.8	0.9	1.0
cc =	.616	.614	.612	.610	.617	.605	.603	.601	.598	.596
ζc =	231.7	50.99	19.78	9.612	5.256	3.077	1.876	1.169	0.734	0.480

Fig. 90.

When a diaphragm was placed in a tube of uniform section (fig. 90) the following values were obtained, ω₁ being the area of the orifice and ω that of the pipe:—

ω1/ω =	0.1	0.2	0.3	0.4	0.5	0.6	0.7	0.8	0.9	1.0
ce =	.624	.632	.643	.659	.681	.712	.755	.813	.892	1.00
ξc =	225.9	47.77	30.83	7.801	1.753	1.796	.797	.290	.060	.000

Elbows.—Weisbach considers the loss of head at elbows (fig. 91) to be due to a contraction formed by the stream. From experiments with a pipe 11/4 in. diameter, he found the loss of head

ɧe = ζεv2 / 2g;

(5)

ζ_e = 0.9457 sin² 1/2φ + 2.047 sin⁴ 1/2φ.

φ =	20°	40°	60°	80°	90°	100°	110°	120°	130°	140°
ζε =	0.046	0.139	0.364	0.740	0.984	1.260	1.556	1.861	2.158	2.431

Hence at a right-angled elbow the whole head due to the velocity very nearly is lost.

 

Fig. 91.

 

Fig. 92.

Bends.—Weisbach traces the loss of head at curved bends to a similar cause to that at elbows, but the coefficients for bends are not very satisfactorily ascertained. Weisbach obtained for the loss of head at a bend in a pipe of circular section

ɧ_b = ζ_bv² / 2g;

(6)

ζ_b = 0.131 + 1.847 (d/2ρ)^7/2,

where d is the diameter of the pipe and ρ the radius of curvature of the bend. The resistance at bends is small and at present very ill determined.

Valves, Cocks and Sluices.—These produce a contraction of the water-stream, similar to that for an abrupt diminution of section already discussed. The loss of head may be taken as before to be

ɧ_v = ζ_vv² / 2g;

(7)

where v is the velocity in the pipe beyond the valve and ζ_v a coefficient determined by experiment. The following are Weisbach’s results.

Sluice in Pipe of Rectangular Section (fig. 92). Section at sluice = ω₁ in pipe = ω.

ω1/ω =	1.0	0.9	0.8	0.7	0.6	0.5	0.4	0.3	0.2	0.1
ζv =	0.00	.09	.39	.95	2.08	4.02	8.12	17.8	44.5	193

Sluice in Cylindrical Pipe (fig. 93).

Ratio of height of opening to diameter of pipe${\Big \}}$	1.0	7/8	3/4	5/8	1/2	3/8	1/4	1/5
ω1/ω =	1.00	0.948	.856	.740	.609	.466	.315	.159
ζv =	0.00	0.07	0.26	0.81	2.06	5.52	17.0	97.8

Fig. 93.	Fig. 94.

Cock in a Cylindrical Pipe (fig. 94). Angle through which cock is turned = θ.

θ =	5°	10°	15°	20°	25°	30°	35°
Ratio of cross sections${\Big \}}$	.926	.850	.772	.692	.613	.535	.458
ζv =	.05	.29	.75	1.56	3.10	5.47	9.68

θ =	40°	45°	50°	55°	60°	65°	82°
Ratio of cross sections${\Big \}}$	.385	.315	.250	.190	.137	.091	0
ζv =	17.3	31.2	52.6	106	206	486	∞

Throttle Valve in a Cylindrical Pipe (fig. 95)

θ =	5°	10°	15°	20°	25°	30°	35°	40°
ζv =	.24	.52	.90	1.54	2.51	3.91	6.22	10.8

θ =	45°	50°	55°	60°	65°	70°	90°
ζv =	18.7	32.6	58.8	118	256	751	∞

Fig. 95.

§ 84. Practical Calculations on the Flow of Water in Pipes.—In the following explanations it will be assumed that the pipe is of so great a length that only the loss of head in friction against the surface of the pipe needs to be considered. In general it is one of the four quantities d, i, v or Q which requires to be determined. For since the loss of head h is given by the relation h = il, this need not be separately considered.

There are then three equations (see eq. 4, § 72, and 9a, § 76) for the solution of such problems as arise:—

ζ = α (1 + 1/12d);

(1)

where α = 0.005 for new and = 0.01 for incrusted pipes.

ζv² / 2g = 1/4di.

(2)

Q = 1/4πd ²v.

(3)

Problem 1. Given the diameter of the pipe and its virtual slope, to find the discharge and velocity of flow. Here d and i are given, and Q and v are required. Find ζ from (1); then v from (2); lastly Q from (3). This case presents no difficulty.

By combining equations (1) and (2), v is obtained directly:—

v = √ (gdi /2ζ) = √ (g/2α) √ [di / {1 + 1/12d}].

(4)

For new pipes	√ (g/2α) = 56.72
For incrusted pipes	= 40.13

For pipes not less than 1, or more than 4 ft. in diameter, the mean values of ζ are

For new pipes	0.00526
For incrusted pipes	0.01052.

Using these values we get the very simple expressions—

v = 55.31 √ (di) for new pipes
= 39.11 √ (di) for incrusted pipes.

(4a)

Within the limits stated, these are accurate enough for practical purposes, especially as the precise value of the coefficient ζ cannot be known for each special case.

Problem 2. Given the diameter of a pipe and the velocity of flow, to find the virtual slope and discharge. The discharge is given by (3); the proper value of ζ by (1); and the virtual slope by (2). This also presents no special difficulty.

Problem 3. Given the diameter of the pipe and the discharge, to find the virtual slope and velocity. Find v from (3); ζ from (1); lastly i from (2). If we combine (1) and (2) we get

i = ζ (v²/2g) (4/d) = 2a {1 + 1/12d} v²/gd;

(5)

and, taking the mean values of ζ for pipes from 1 to 4 ft. diameter, given above, the approximate formulae are

i = 0.0003268 v²/d for new pipes
= 0.0006536 v²/d for incrusted pipes.

(5a)

Problem 4. Given the virtual slope and the velocity, to find the diameter of the pipe and the discharge. The diameter is obtained from equations (2) and (1), which give the quadratic expression

d ² − d (2αv²/gi) − αv²/6gi = 0.

∴ d = αv²/gi + √ {(αv²/gi) (αv²/gi + 1/6)}.

(6)

For practical purposes, the approximate equations

d = 2αv²/gi + 1/12
  = 0.00031 v²/i + .083 for new pipes
  = 0.00062 v²/i + .083 for incrusted pipes

(6a)

are sufficiently accurate.

Problem 5. Given the virtual slope and the discharge, to find the diameter of the pipe and velocity of flow. This case, which often occurs in designing, is the one which is least easy of direct solution. From equations (2) and (3) we get—

d 5 = 32ζQ2 / gπ2i.

(7)

If now the value of ζ in (1) is introduced, the equation becomes very cumbrous. Various approximate methods of meeting the difficulty may be used.

(a) Taking the mean values of ζ given above for pipes of 1 to 4 ft. diameter we get

d = 5√ (32ζ/gπ2) 5√ (Q2/i)

(8)

= 0.2216 ⁵√ (Q²/i) for new pipes
= 0.2541 ⁵√ (Q²/i) for incrusted pipes;

equations which are interesting as showing that when the value of ζ is doubled the diameter of pipe for a given discharge is only increased by 13%.

(b) A second method is to obtain a rough value of d by assuming ζ = α. This value is

d ′ = ⁵√ (32Q²/gπ²i) ⁵√ α = 0.6319 ⁵√ (Q²/i) ⁵√ α.

Then a very approximate value of ζ is

ζ′ = α (1 + 1/12d ′);

and a revised value of d, not sensibly differing from the exact value, is

d ″ = ⁵√ (32Q²/gπ²i) ⁵√ ζ′ = 0.6319 ⁵√ (Q²/i) ⁵√ ζ′.

(c) Equation 7 may be put in the form

d = ⁵√ (32αQ²/gπ²i) ⁵√ (1 + 1/12d).

(9)

Expanding the term in brackets,

⁵√ (1 + 1/12d) = 1 + 1/60d − 1/1800d ² . . .

Neglecting the terms after the second,

d = 5√ (32α/gπ2) 5√ (Q2/i) · {1 + 1/60d } = 5√ (32α/gπ2) 5√ (Q2/i) + 0.01667;

(9a)

and

⁵√ (32α/gπ²) = 0.219 for new pipes
= 0.252 for incrusted pipes.

Fig. 96.

Fig. 97.

§ 85. Arrangement of Water Mains for Towns’ Supply.—Town mains are usually supplied oy gravitation from a service reservoir, which in turn is supplied by gravitation from a storage reservoir or by pumping from a lower level. The service reservoir should contain three days’ supply or in important cases much more. Its elevation should be such that water is delivered at a pressure of at least about 100 ft. to the highest parts of the district. The greatest pressure in the mains is usually about 200 ft., the pressure for which ordinary pipes and fittings are designed. Hence if the district supplied has great variations of level it must be divided into zones of higher and lower pressure. Fig. 96 shows a district of two zones each with its service reservoir and a range of pressure in the lower district from 100 to 200 ft. The total supply required is in England about 25 gallons per head per day. But in many towns, and especially in America, the supply is considerably greater, but also in many cases a good deal of the supply is lost by leakage of the mains. The supply through the branch mains of a distributing system is calculated from the population supplied. But in determining the capacity of the mains the fluctuation of the demand must be allowed for. It is usual to take the maximum demand at twice the average demand. Hence if the average demand is 25 gallons per head per day, the mains should be calculated for 50 gallons per head per day.

Fig. 98.

§ 86. Determination of the Diameters of Different Parts of a Water Main.—When the plan of the arrangement of mains is determined upon, and the supply to each locality and the pressure required is ascertained, it remains to determine the diameters of the pipes. Let fig. 97 show an elevation of a main ABCD . . ., R being the reservoir from which the supply is derived. Let NN be the datum line of the levelling operations, and H_a, H_b . . . the heights of the main above the datum line, H_r being the height of the water surface in the reservoir from the same datum. Set up next heights AA₁, BB₁, . . . representing the minimum pressure height necessary for the adequate supply of each locality. Then A₁B₁C₁D₁ . . . is a line which should form a lower limit to the line of virtual slope. Then if heights ɧ_a, ɧ_b, ɧ_c . . . are taken representing the actual losses of head in each length l_a, l_b, l_c. . . of the main, A₀B₀C₀ will be the line of virtual slope, and it will be obvious at what points such as D₀ and E₀, the pressure is deficient, and a different choice of diameter of main is required. For any point z in the length of the main, we have

Pressure height = H_r − H_z − (ɧ_a + ɧ_b + . . . ɧ_z).

Where no other circumstance limits the loss of head to be assigned to a given length of main, a consideration of the safety of the main from fracture by hydraulic shock leads to a limitation of the velocity of flow. Generally the velocity in water mains lies between 11/2 and 41/2 ft. per second. Occasionally the velocity in pipes reaches 10 ft. per second, and in hydraulic machinery working under enormous pressures even 20 ft. per second. Usually the velocity diminishes along the main as the discharge diminishes, so as to reduce somewhat the total loss of head which is liable to render the pressure insufficient at the end of the main.

J. T. Fanning gives the following velocities as suitable in pipes for towns’ supply:—

Diameter in inches	4	8	12	18	24	30	36
Velocity in feet per sec.	2.5	3.0	3.5	4.5	5.3	6.2	7.0

§ 87. Branched Pipe connecting Reservoirs at Different Levels.—Let A, B, C (fig. 98) be three reservoirs connected by the arrangement of pipes shown,—l₁, d₁, Q₁, v₁; l₂, d₂, Q₂, v₂; l₃, d₃, Q₃, v₃ being the length, diameter, discharge and velocity in the three portions of the main pipe. Suppose the dimensions and positions of the pipes known and the discharges required.

If a pressure column is introduced at X, the water will rise to a height XR, measuring the pressure at X, and aR, Rb, Rc will be the lines of virtual slope. If the free surface level at R is above b, the reservoir A supplies B and C, and if R is below b, A and B supply C. Consequently there are three cases:—

I.	R above b; Q1 = Q2 + Q3.
II.	R level with b; Q1 = Q3; Q2 = 0
III.	R below b; Q1 + Q2 = Q3.

To determine which case has to be dealt with in the given conditions, suppose the pipe from X to B closed by a sluice. Then there is a simple main, and the height of free surface h′ at X can be determined. For this condition

h_a − h′ = ζ (v₁²/2g) (4l₁/d₁)
= 32ζQ′²l₁/gπ²d₁⁵;

h′ − h_c = ζ (v₃²/2g) (4l₃/d₃)
= 32ζQ′²l₃/gπ²d₃⁵;

where Q′ is the common discharge of the two portions of the pipe. Hence

(h_a − h′)/(h′ − h_c) = l₁d₃⁵/l₃d₁⁵,

from which h′ is easily obtained. If then h′ is greater than hb, opening the sluice between X and B will allow flow towards B, and the case in hand is case I. If h′ is less than h_b, opening the sluice will allow flow from B, and the case is case III. If h′ = h_b, the case is case II., and is already completely solved.

The true value of h must lie between h′ and h_b. Choose a new value of h, and recalculate Q₁, Q₂, Q₃. Then if

Q₁ > Q₂ + Q₃ in case I.,

or

Q₁ + Q₂ > Q₃ in case III.,

the value chosen for h is too small, and a new value must be chosen.

If

Q₁ < Q₂ + Q₃ in case I.,

or

Q₁ + Q₂ < Q₃ in case III.,

the value of h is too great.

Since the limits between which h can vary are in practical cases not very distant, it is easy to approximate to values sufficiently accurate.

§ 88. Water Hammer.—If in a pipe through which water is flowing a sluice is suddenly closed so as to arrest the forward movement of the water, there is a rise of pressure which in some cases is serious enough to burst the pipe. This action is termed water hammer or water ram. The fluctuation of pressure is an oscillating one and gradually dies out. Care is usually taken that sluices should only be closed gradually and then the effect is inappreciable. Very careful experiments on water hammer were made by N. J. Joukowsky at Moscow in 1898 (Stoss in Wasserleitungen, St Petersburg, 1900), and the results are generally confirmed by experiments made by E. B. Weston and R. C. Carpenter in America. Joukowsky used pipes, 2, 4 and 6 in. diameter, from 1000 to 2500 ft. in length. The sluice closed in 0.03 second, and the fluctuations of pressure were automatically registered. The maximum excess pressure due to water-hammer action was as follows:—

Pipe 4-in. diameter.		Pipe 6-in. diameter.
Velocity  ft. per sec.	Excess Pressure.  ℔ per sq. in.	Velocity  ft. per sec.	Excess Pressure.  ℔ per sq. in.
0.5	31	0.6	43
2.9	168	3.0	173
4.1	232	5.6	369
9.2	519	7.5	426

In some cases, in fixing the thickness of water mains, 100 ℔ per sq. in. excess pressure is allowed to cover the effect of water hammer. With the velocities usual in water mains, especially as no valves can be quite suddenly closed, this appears to be a reasonable allowance (see also Carpenter, Am. Soc. Mech. Eng., 1893).

IX. FLOW OF COMPRESSIBLE FLUIDS IN PIPES

§ 89. Flow of Air in Long Pipes.—When air flows through a long pipe, by far the greater part of the work expended is used in overcoming frictional resistances due to the surface of the pipe. The work expended in friction generates heat, which for the most part must be developed in and given back to the air. Some heat may be transmitted through the sides of the pipe to surrounding materials, but in experiments hitherto made the amount so conducted away appears to be very small, and if no heat is transmitted the air in the tube must remain sensibly at the same temperature during expansion. In other words, the expansion may be regarded as isothermal expansion, the heat generated by friction exactly neutralizing the cooling due to the work done. Experiments on the pneumatic tubes used for the transmission of messages, by R. S. Culley and R. Sabine (Proc. Inst. Civ. Eng. xliii.), show that the change of temperature of the air flowing along the tube is much less than it would be in adiabatic expansion.

§ 90. Differential Equation of the Steady Motion of Air Flowing in a Long Pipe of Uniform Section.—When air expands at a constant absolute temperature τ, the relation between the pressure p in pounds per square foot and the density or weight per cubic foot G is given by the equation

p/G = cτ,

(1)

where c = 53.15. Taking τ = 521, corresponding to a temperature of 60° Fahr.,

cτ = 27690 foot-pounds.

(2)

 

Fig. 99.

The equation of continuity, which expresses the condition that in steady motion the same weight of fluid, W, must pass through each cross section of the stream in the unit of time, is

GΩu = W = constant,

(3)

where Ω is the section of the pipe and u the velocity of the air. Combining (1) and (3),

Ωup/W = cτ = constant.

(3a)

Since the work done by gravity on the air during its flow through a pipe due to variations of its level is generally small compared with the work done by changes of pressure, the former may in many cases be neglected.

Consider a short length dl of the pipe limited by sections A₀, A₁ at a distance dl (fig. 99). Let p, u be the pressure and velocity at A₀, p + dp and u + du those at A₁. Further, suppose that in a very short time dt the mass of air between A₀A₁ comes to A′₀A′₁ so that A₀A′₀ = udt and A₁A′₁ = (u + du) dt₁. Let Ω be the section, and m the hydraulic mean radius of the pipe, and W the weight of air flowing through the pipe per second.

From the steadiness of the motion the weight of air between the sections A₀A′₀, and A₁A′₁ is the same. That is,

Wdt = GΩudt = GΩ (u+du)dt.

By analogy with liquids the head lost in friction is, for the length dl (see § 72, eq. 3), ζ (u²/2g) (dl/m). Let H = u²/2g. Then the head lost is ζ(H/m)dl; and, since Wdt ℔ of air flow through the pipe in the time considered, the work expended in friction is −ζ (H/m)W dl dt. The change of kinetic energy in dt seconds is the difference of the kinetic energy of A₀A′₀ and A₁A′₁, that is,

(W/g)dt{(u+du)²−u²}/2 = (W/g) u du dt = WdHdt.

The work of expansion when Ωudt cub. ft. of air at a pressure p expand to Ω(u + du) dt cub. ft. is Ωp du dt. But from (3a) u = cτW/Ωp, and therefore

du / dp = −cτW / Ωp².

And the work done by expansion is −(cτW/p) dp dt.

The work done by gravity on the mass between A₀ and A₁ is zero if the pipe is horizontal, and may in other cases be neglected without great error. The work of the pressures at the sections A₀A₁ is

pΩudt−(p+dp) Ω (u + du) dt
= −(pdu + udp) Ωdt

But from (3a)

pu = constant,
pdu+udp = 0,

and the work of the pressures is zero. Adding together the quantities of work, and equating them to the change of kinetic energy,

WdHdt = −(cτW/p) dp dt − ζ (H/m) W dl dt
dH + (cτ/p) dp + ζ (H/m) dl = 0,
dH/H + (cτ/Hp) dp + ζdl / m = 0

(4)

But

u = cτW / Ωp,

and

H = u²/2g = c²τ²W² / 2gΩ²p²,

∴ dH/H + (2gΩ²p / cτW²) dp + ζdl / m = 0.

(4a)

For tubes of uniform section m is constant; for steady motion W is constant; and for isothermal expansion τ is constant. Integrating,

log H + gΩ²p² / W²cτ + ζ l / m = constant;

(5)

for

l = 0, let H = H₀, and p = p₀;

and for

l = l, let H = H₁, and p = p₁.

log (H₁/H₀) + (gΩ² / W²cτ) (p₁² − p₀²) + ζ l / m = 0.

(5a)

where p₀ is the greater pressure and p₁ the less, and the flow is from A₀ towards A₁.

By replacing W and H,

log (p₀/p₁) + (gcτ / u₀²p₀²) (p₁² − p₀² + ζ l/m = 0

(6)

Hence the initial velocity in the pipe is

u₀ = √ [{gcτ (p₀² − p₁²)} / {p₀² (ζ l/m + log (p₀ / p₁) }].

(7)

When l is great, log p₀/p₁ is comparatively small, and then

u₀ = √ [ (gcτm/ζ l) {(p₀² − p₁²) / p₀²} ],

(7a)

a very simple and easily used expression. For pipes of circular section m = d/4, where d is the diameter:—

u₀ = √ [ (gcτd / 4ζ l) {(p₀² − p₁²) / p₀²} ];

(7b)

or approximately

u₀ = (1.1319 − 0.7264 p₁/p₀) √ (gcτd / 4ζl).

(7c)

§ 91. Coefficient of Friction for Air.—A discussion by Professor Unwin of the experiments by Culley and Sabine on the rate of transmission of light carriers through pneumatic tubes, in which there is steady flow of air not sensibly affected by any resistances other than surface friction, furnished the value ζ = .007. The pipes were lead pipes, slightly moist, 21/4 in. (0.187 ft.) in diameter, and in lengths of 2000 to nearly 6000 ft.

In some experiments on the flow of air through cast-iron pipes A. Arson found the coefficient of friction to vary with the velocity and diameter of the pipe. Putting

ζ = α/v + β,

(8)

he obtained the following values—

Diameter of Pipe  in feet.	α	β	ζ for 100 ft.  per second.
1.64	.00129	.00483	.00484
1.07	.00972	.00640	.00650
.83	.01525	.00704	.00719
.338	.03604	.00941	.00977
.266	.03790	.00959	.00997
.164	.04518	.01167	.01212

It is worth while to try if these numbers can be expressed in the form proposed by Darcy for water. For a velocity of 100 ft. per second, and without much error for higher velocities, these numbers agree fairly with the formula

ζ = 0.005 (1 + 3/10d),

(9)

which only differs from Darcy’s value for water in that the second term, which is always small except for very small pipes, is larger.

Some later experiments on a very large scale, by E. Stockalper at the St Gotthard Tunnel, agree better with the value

ζ＝0.0028 (1 + 3/10d).

These pipes were probably less rough than Arson’s.

When the variation of pressure is very small, it is no longer safe to neglect the variation of level of the pipe. For that case we may neglect the work done by expansion, and then

z0 − z1 − p0/G0 − p1/G1 − ζ (v2/2g) (l/m)＝0,

(10)

precisely equivalent to the equation for the flow of water, z₀ and z₁ being the elevations of the two ends of the pipe above any datum, p₀ and p₁ the pressures, G₀ and G₁ the densities, and v the mean velocity in the pipe. This equation may be used for the flow of coal gas.

§ 92. Distribution of Pressure in a Pipe in which Air is Flowing.—From equation (7a) it results that the pressure p, at l ft. from that end of the pipe where the pressure is p₀, is

p＝p0 √{1 − ζlu02 / mgcτ};

(11)

which is of the form

p＝√ (al+b)

for any given pipe with given end pressures. The curve of free surface level for the pipe is, therefore, a parabola with horizontal axis. Fig. 100 shows calculated curves of pressure for two of Sabine’s experiments, in one of which the pressure was greater than atmospheric pressure, and in the other less than atmospheric pressure. The observed pressures are given in brackets and the calculated pressures without brackets. The pipe was the pneumatic tube between Fenchurch Street and the Central Station, 2818 yds. in length. The pressures are given in inches of mercury.

Fig. 100.

Variation of Velocity in the Pipe.—Let p₀, u₀ be the pressure and velocity at a given section of the pipe; p, u, the pressure and velocity at any other section. From equation (3a)

up＝cτW / Ω＝constant;

so that, for any given uniform pipe,

up＝u0p0, u＝u0p0 / p;

(12)

which gives the velocity at any section in terms of the pressure, which has already been determined. Fig. 101 gives the velocity curves for the two experiments of Culley and Sabine, for which the pressure curves have already been drawn. It will be seen that the velocity increases considerably towards that end of the pipe where the pressure is least.

Fig. 101.

§ 93. Weight of Air Flowing per Second.—The weight of air discharged per second is (equation 3a)—

W＝Ωu₀p₀ / cτ.

From equation (7b), for a pipe of circular section and diameter d,

W＝1/4π √ (gd 5 (p02 − p12) / ζ lcτ), ＝.611 √ (d 5 (p02 − p12) / ζ lτ).

(13)

Approximately

W＝(.6916p0 − .4438p1) (d 5 / ζ lτ)1/2.

(13a)

§ 94. Application to the Case of Pneumatic Tubes for the Transmission of Messages.—In Paris, Berlin, London, and other towns, it has been found cheaper to transmit messages in pneumatic tubes than to telegraph by electricity. The tubes are laid underground with easy curves; the messages are made into a roll and placed in a light felt carrier, the resistance of which in the tubes in London is only 3/4 oz. A current of air forced into the tube or drawn through it propels the carrier. In most systems the current of air is steady and continuous, and the carriers are introduced or removed without materially altering the flow of air.

Time of Transit through the Tube.—Putting t for the time of transit from 0 to l,

$t=\int _{0}^{l}dl/u,$

From (4a) neglecting dH/H, and putting m = d/4,

dl＝gdΩ²p dp / 2ζW²cr.

From (1) and (3)

u＝Wcτ / pΩ;
dl/u＝gdΩ³p² dp / 2ζW³c²τ²;

t＝$\int _{p_{1}}^{p_{0}};$gdΩ³p² dp / 2ζW³c²τ²,

＝gdΩ3 (p03 − p13) / 6ζW3c2τ2.

(14)

But

W = p0u0Ω / cτ;

∴ t＝gdcτ (p03 − p13) / 6ζp03u03, ＝ζ1/2 l3/2 (p03 − p13) / 6(gcτd)1/2 (p02 − p12)3/2;

(15)

If τ = 521°, corresponding to 60° F.,

t＝.001412 ζ1/2 l3/2 (p03 − p13) / d 1/2 (p02 − p12)3/2;

(15a)

which gives the time of transmission in terms of the initial and final pressures and the dimensions of the tube.

Mean Velocity of Transmission.—The mean velocity is l/t; or, for τ = 521°,

umean＝0.708 √ {d (p02 − p12)3/2 / ζ l (p03 − p13)}.

(16)

The following table gives some results:—

		Absolute  Pressures in  ℔ per sq. in.		Mean Velocities for Tubes of a length in feet.
		p0	p1	1000	2000	3000	4000	5000
Vacuum  Working	$\scriptstyle {\left\{{\begin{matrix}\ \\\ \end{matrix}}\right.}$	15	5	99.4	70.3	57.4	49.7	44.5
		15	10	67.2	47.5	38.8	34.4	30.1
Pressure  Working	$\scriptstyle {\left\{{\begin{matrix}\ \\\\\ \ \end{matrix}}\right.}$	20	15	57.2	40.5	33.0	28.6	25.6
		25	15	74.6	52.7	43.1	37.3	33.3
		30	15	84.7	60.0	49.0	42.4	37.9

Limiting Velocity in the Pipe when the Pressure at one End is diminished indefinitely.—If in the last equation there be put p₁ = 0, then

u′_mean＝0.708 √ (d / ζl);

where the velocity is independent of the pressure p₀ at the other end, a result which apparently must be absurd. Probably for long pipes, as for orifices, there is a limit to the ratio of the initial and terminal pressures for which the formula is applicable.

X. FLOW IN RIVERS AND CANALS

§ 95. Flow of Water in Open Canals and Rivers.—When water flows in a pipe the section at any point is determined by the form of the boundary. When it flows in an open channel with free upper surface, the section depends on the velocity due to the dynamical conditions.

Suppose water admitted to an unfilled canal. The channel will gradually fill, the section and velocity at each point gradually changing. But if the inflow to the canal at its head is constant, the increase of cross section and diminution of velocity at each point attain after a time a limit. Thenceforward the section and velocity at each point are constant, and the motion is steady, or permanent regime is established.

If when the motion is steady the sections of the stream are all equal, the motion is uniform. By hypothesis, the inflow Ωv is constant for all sections, and Ω is constant; therefore v must be constant also from section to section. The case is then one of uniform steady motion. In most artificial channels the form of section is constant, and the bed has a uniform slope. In that case the motion is uniform, the depth is constant, and the stream surface is parallel to the bed. If when steady motion is established the sections are unequal, the motion is steady motion with varying velocity from section to section. Ordinary rivers are in this condition, especially where the flow is modified by weirs or obstructions. Short unobstructed lengths of a river may be treated as of uniform section without great error, the mean section in the length being put for the actual sections.

 

Fig. 102.

In all actual streams the different fluid filaments have different velocities, those near the surface and centre moving faster than those near the bottom and sides. The ordinary formulae for the flow of streams rest on a hypothesis that this variation of velocity may be neglected, and that all the filaments may be treated as having a common velocity equal to the mean velocity of the stream. On this hypothesis, a plane layer abab (fig. 102) between sections normal to the direction of motion is treated as sliding down the channel to a′a′b′b′ without deformation. The component of the weight parallel to the channel bed balances the friction against the channel, and in estimating the friction the velocity of rubbing is taken to be the mean velocity of the stream. In actual streams, however, the velocity of rubbing on which the friction depends is not the mean velocity of the stream, and is not in any simple relation with it, for channels of different forms. The theory is therefore obviously based on an imperfect hypothesis. However, by taking variable values for the coefficient of friction, the errors of the ordinary formulae are to a great extent neutralized, and they may be used without leading to practical errors. Formulae have been obtained based on less restricted hypotheses, but at present they are not practically so reliable, and are more complicated than the formulae obtained in the manner described above.

§ 96. Steady Flow of Water with Uniform Velocity in Channels of Constant Section.—Let aa′, bb′ (fig. 103) be two cross sections normal to the direction of motion at a distance dl. Since the mass aa′bb′ moves uniformly, the external forces acting on it are in equilibrium. Let Ω be the area of the cross sections, χ the wetted perimeter, pq + qr + rs, of a section. Then the quantity m = Ω/χ is termed the hydraulic mean depth of the section. Let v be the mean velocity of the stream, which is taken as the common velocity of all the particles, i, the slope or fall of the stream in feet, per foot, being the ratio bc/ab.

Fig. 103.

The external forces acting on aa′bb′ parallel to the direction of motion are three:—(a) The pressures on aa′ and bb′, which are equal and opposite since the sections are equal and similar, and the mean pressures on each are the same. (b) The component of the weight W of the mass in the direction of motion, acting at its centre of gravity g. The weight of the mass aa′bb′ is GΩ dl, and the component of the weight in the direction of motion is GΩdl × the cosine of the angle between Wg and ab, that is, GΩdl cos abc = GΩ dl bc/ab = GΩidl. (c) There is the friction of the stream on the sides and bottom of the channel. This is proportional to the area χdl of rubbing surface and to a function of the velocity which may be written ƒ(v); ƒ(v) being the friction per sq. ft. at a velocity v. Hence the friction is −χ dl ƒ(v). Equating the sum of the forces to zero,

GΩi dl − χ dl ƒ(v) = 0, ƒ(v) / G = Ωi / χ = mi.

(1)

But it has been already shown (§ 66) that ƒ(v) = ζGv²/2g,

∴ ζv2 / 2g = mi.

(2)

This may be put in the form

v = √ (2g/ζ) √ (mi) = c √ (mi);

(2a)

where c is a coefficient depending on the roughness and form of the channel.

The coefficient of friction ζ varies greatly with the degree of roughness of the channel sides, and somewhat also with the velocity. It must also be made to depend on the absolute dimensions of the section, to eliminate the error of neglecting the variations of velocity in the cross section. A common mean value assumed for ζ is 0.00757. The range of values will be discussed presently.

It is often convenient to estimate the fall of the stream in feet per mile, instead of in feet per foot. If ƒ is the fall in feet per mile,

ƒ = 5280 i.

Putting this and the above value of ζ in (2a), we get the very simple and long-known approximate formula for the mean velocity of a stream—

v = 1/4 1/2 √ (2mf).

(3)

The flow down the stream per second, or discharge of the stream, is

Q = Ωv = Ωc √ (mi).

(4)

§ 97. Coefficient of Friction for Open Channels.—Various expressions have been proposed for the coefficient of friction for channels as for pipes. Weisbach, giving attention chiefly to the variation of the coefficient of friction with the velocity, proposed an expression of the form

ζ = α (1 + β/v),

(5)

and from 255 experiments obtained for the constants the values

α = 0.007409; β = 0.1920.

This gives the following values at different velocities:—

v =  ζ =

0.3  0.01215

0.5  0.01025

0.7  0.00944

1  0.00883

11/2  0.00836

2  0.00812

3  0.90788

5  0.00769

7  0.00761

10  0.00755

15  0.00750

In using this value of ζ when v is not known, it is best to proceed by approximation.

§ 98. Darcy and Bazin’s Expression for the Coefficient of Friction.—Darcy and Bazin’s researches have shown that ζ varies very greatly for different degrees of roughness of the channel bed, and that it also varies with the dimensions of the channel. They give for ζ an empirical expression (similar to that for pipes) of the form

ζ = α (1 + β/m);

(6)

where m is the hydraulic mean depth. For different kinds of channels they give the following values of the coefficient of friction:—

Kind of Channel.		α	β
I.	Very smooth channels, sides of smooth cement or planed timber	.00294	0.10
II.	Smooth channels, sides of ashlar, brickwork, planks	.00373	0.23
III.	Rough channels, sides of rubble masonry or pitched with stone	.00471	0.82
IV.	Very rough canals in earth	.00549	4.10
V.	Torrential streams encumbered with detritus	.00785	5.74

The last values (Class V.) are not Darcy and Bazin’s, but are taken from experiments by Ganguillet and Kutter on Swiss streams.

The following table very much facilitates the calculation of the mean velocity and discharge of channels, when Darcy and Bazin’s value of the coefficient of friction is used. Taking the general formula for the mean velocity already given in equation (2a) above,

v = c √ (mi),

where c = √ (2g/ζ), the following table gives values of c for channels of different degrees of roughness, and for such values of the hydraulic mean depths as are likely to occur in practical calculations:—

Values of c in v = c √ (mi), deduced from Darcy and Bazin’s Values.

Hydraulic Mean.  Depth = m.	Very Smooth  Channels. Cement.	Smooth Channels. Ashlar or  Brickwork.	Rough  Channels.  Rubble Masonry.	Very Rough  Channels. Canals in Earth.	Excessively  Rough Channels  encumbered with Detritus.
.25	125	95	57	26	18.5
.5	135	110	72	36	25.6
.75	139	116	81	42	30.8
1.0	141	119	87	48	34.9
1.5	143	122	94	56	41.2
2.0	144	124	98	62	46.0
2.5	145	126	101	67	..
3.0	145	126	104	70	53
3.5	146	127	105	73	..
4.0	146	128	106	76	58
4.5	146	128	107	78	..
5.0	146	128	108	80	62
5.5	146	129	109	82	..
6.0	147	129	110	84	65
6.5	147	129	110	85	..
7.0	147	129	110	86	67
7.5	147	129	111	87	..
8.0	147	130	111	88	69
8.5	147	130	112	89	..
9.0	147	130	112	90	71
9.5	147	130	112	90	..
10.0	147	130	112	91	72
11	147	130	113	92	..
12	147	130	113	93	74
13	147	130	113	94	..
14	147	130	113	95	..
15	147	130	114	96	77
16	147	130	114	97	..
17	147	130	114	97	..
18	147	130	114	98	..
20	147	131	114	98	80
25	148	131	115	100	..
30	148	131	115	102	83
40	148	131	116	103	85
50	148	131	116	104	86
∞	148	131	117	108	91

§ 99. Ganguillet and Kutter’s Modified Darcy Formula.—Starting from the general expression v = c√mi, Ganguillet and Kutter examined the variations of c for a wider variety of cases than those discussed by Darcy and Bazin. Darcy and Bazin’s experiments were confined to channels of moderate section, and to a limited variation of slope. Ganguillet and Kutter brought into the discussion two very distinct and important additional series of results. The gaugings of the Mississippi by A. A. Humphreys and H. L. Abbot afford data of discharge for the case of a stream of exceptionally large section and or very low slope. On the other hand, their own measurements of the flow in the regulated channels of some Swiss torrents gave data for cases in which the inclination and roughness of the channels were exceptionally great. Darcy and Bazin’s experiments alone were conclusive as to the dependence of the coefficient c on the dimensions of the channel and on its roughness of surface. Plotting values of c for channels of different inclination appeared to indicate that it also depended on the slope of the stream. Taking the Mississippi data only, they found

c = 256 for an inclination of	0.0034 per thousand,
= 154    ”     ”	0.02     ”

so that for very low inclinations no constant value of c independent of the slope would furnish good values of the discharge. In small rivers, on the other hand, the values of c vary little with the slope. As regards the influence of roughness of the sides of the channel a different law holds. For very small channels differences of roughness have a great influence on the discharge, but for very large channels different degrees of roughness have but little influence, and for indefinitely large channels the influence of different degrees of roughness must be assumed to vanish. The coefficients given by Darcy and Bazin are different for each of the classes of channels of different roughness, even when the dimensions of the channel are infinite. But, as it is much more probable that the influence of the nature of the sides diminishes indefinitely as the channel is larger, this must be regarded as a defect in their formula.

Comparing their own measurements in torrential streams in Switzerland with those of Darcy and Bazin, Ganguillet and Kutter found that the four classes of coefficients proposed by Darcy and Bazin were insufficient to cover all cases. Some of the Swiss streams gave results which showed that the roughness of the bed was markedly greater than in any of the channels tried by the French engineers. It was necessary therefore in adopting the plan of arranging the different channels in classes of approximately similar roughness to increase the number of classes. Especially an additional class was required for channels obstructed by detritus.

To obtain a new expression for the coefficient in the formula

v = √ (2g / ζ) √ (mi) = c √ (mi),

Ganguillet and Kutter proceeded in a purely empirical way. They found that an expression of the form

c = α / (1 + β/√ m)

could be made to fit the experiments somewhat better than Darcy’s expression. Inverting this, we get

1/c = 1/α + β/α √ m,

an equation to a straight line having 1/√m for abscissa, 1/c for ordinate, and inclined to the axis of abscissae at an angle the tangent of which is β/α.

Plotting the experimental values of 1/c and 1/√ m, the points so found indicated a curved rather than a straight line, so that β must depend on α. After much comparison the following form was arrived at—

c = (A + l/n) / (1 + An / √ m),

where n is a coefficient depending only on the roughness of the sides of the channel, and A and l are new coefficients, the value of which remains to be determined. From what has been already stated, the coefficient c depends on the inclination of the stream, decreasing as the slope i increases.

Let

A = a + p/i.

Then

c = (a + l/n + p/i) / {1 + (a + p/i) n/√ m},

the form of the expression for c ultimately adopted by Ganguillet and Kutter.

For the constants a, l, p Ganguillet and Kutter obtain the values 23, 1 and 0.00155 for metrical measures, or 41.6, 1.811 and 0.00281 for English feet. The coefficient of roughness n is found to vary from 0.008 to 0.050 for either metrical or English measures.

The most practically useful values of the coefficient of roughness n are given in the following table:—

Nature of Sides of Channel.	Coefficient of Roughness n.
Well-planed timber	0.009
Cement plaster	0.010
Plaster of cement with one-third sand	0.011
Unplaned planks	0.012
Ashlar and brickwork	0.013
Canvas on frames	0.015
Rubble masonry	0.017
Canals in very firm gravel	0.020
Rivers and canals in perfect order, free from stones or weeds	0.025
Rivers and canals in moderately good order, not quite free from stones and weeds${\big \}}$	0.030
Rivers and canals in bad order, with weeds and detritus	0.035
Torrential streams encumbered with detritus	0.050

Ganguillet and Kutter’s formula is so cumbrous that it is difficult to use without the aid of tables.

Lowis D’A. Jackson published complete and extensive tables for facilitating the use of the Ganguillet and Kutter formula (Canal and Culvert Tables, London, 1878). To lessen calculation he puts the formula in this form:—

M = n (41.6 + 0.00281/i);

v = (√ m/n) {(M + 1.811) / (M + √m)} √ (mi).

The following table gives a selection of values of M, taken from Jackson’s tables:—

i =	Values of M for n =
	0.010	0.012	0.015	0.017	0.020	0.025	0.030
.00001	3.2260	3.8712	4.8390	5.4842	6.4520	8.0650	9.6780
.00002	1.8210	2.1852	2.7315	3.0957	3.6420	4.5525	5.4630
.00004	1.1185	1.3422	1.6777	1.9014	2.2370	2.7962	3.3555
.00006	0.8843	1.0612	1.3264	1.5033	1.7686	2.2107	2.6529
.00008	0.7672	0.9206	1.1508	1.3042	1.5344	1.9180	2.3016
.00010	0.6970	0.8364	1.0455	1.1849	1.3940	1.7425	2.0910
.00025	0.5284	0.6341	0.7926	0.8983	1.0568	1.3210	1.5852
.00050	0.4722	0.5666	0.7083	0.8027	0.9444	1.1805	1.4166
.00075	0.4535	0.5442	0.6802	0.7709	0.9070	1.1337	1.3605
.00100	0.4441	0.5329	0.6661	0.7550	0.8882	1.1102	1.3323
.00200	0.4300	0.5160	0.6450	0.7310	0.8600	1.0750	1.2900
.00300	0.4254	0.5105	0.6381	0.7232	0.8508	1.0635	1.2762

A difficulty in the use of this formula is the selection of the coefficient of roughness. The difficulty is one which no theory will overcome, because no absolute measure of the roughness of stream beds is possible. For channels lined with timber or masonry the difficulty is not so great. The constants in that case are few and sufficiently defined. But in the case of ordinary canals and rivers the case is different, the coefficients having a much greater range. For artificial canals in rammed earth or gravel n varies from 0.0163 to 0.0301. For natural channels or rivers n varies from 0.020 to 0.035.

In Jackson’s opinion even Kutter’s numerous classes of channels seem inadequately graduated, and he proposes for artificial canals the following classification:—

I.	Canals in very firm gravel, in perfect order	n = 0.02
II.	Canals in earth, above the average in order	n = 0.0225
III.	Canals in earth, in fair order	n = 0.025
IV.	Canals in earth, below the average in order	n = 0.0275
V.	Canals in earth, in rather bad order, partially over- grown with weeds and obstructed by detritus.${\big \}}$	n = 0.03

Ganguillet and Kutter’s formula has been considerably used partly from its adoption in calculating tables for irrigation work in India. But it is an empirical formula of an unsatisfactory form. Some engineers apparently have assumed that because it is complicated it must be more accurate than simpler formulae. Comparison with the results of gaugings shows that this is not the case. The term involving the slope was introduced to secure agreement with some early experiments on the Mississippi, and there is strong reason for doubting the accuracy of these results.

§ 100. Bazin’s New Formula.—Bazin subsequently re-examined all the trustworthy gaugings of flow in channels and proposed a modification of the original Darcy formula which appears to be more satisfactory than any hitherto suggested (Étude d’une nouvelle formule, Paris, 1898). He points out that Darcy’s original formula, which is of the form mi/v² = α + β/m, does not agree with experiments on channels as well as with experiments on pipes. It is an objection to it that if m increases indefinitely the limit towards which mi/v² tends is different for different values of the roughness. It would seem that if the dimensions of a canal are indefinitely increased the variation of resistance due to differing roughness should vanish. This objection is met if it is assumed that √ (mi/v²) = α + β/√ m, so that if a is a constant mi/v² tends to the limit a when m increases. A very careful discussion of the results of gaugings shows that they can be expressed more satisfactorily by this new formula than by Ganguillet and Kutter’s. Putting the equation in the form ζv²/2g = mi, ζ = 0.002594 (1 + γ/√ m), where γ has the following values:—

I.	Very smooth sides, cement, planed plank, γ =	0.109
II.	Smooth sides, planks, brickwork	0.290
III.	Rubble masonry sides	0.833
IV.	Sides of very smooth earth, or pitching	1.539
V.	Canals in earth in ordinary condition	2.353
VI.	Canals in earth exceptionally rough	3.168

§ 101. The Vertical Velocity Curve.—If at each point along a vertical representing the depth of a stream, the velocity at that point is plotted horizontally, the curve obtained is the vertical velocity curve and it has been shown by many observations that it approximates to a parabola with horizontal axis. The vertex of the parabola is at the level of the greatest velocity. Thus in fig. 104 OA is the vertical at which velocities are observed; v₀ is the surface; v_z the maximum and v_d the bottom velocity. B C D is the vertical velocity curve which corresponds with a parabola having its vertex at C. The mean velocity at the vertical is

v_m = 1/3 [2v_z + v_d + (d_z/d) (v₀ − v_d)].

Fig. 104.

The Horizontal Velocity Curve.—Similarly if at each point along a horizontal representing the width of the stream the velocities are plotted, a curve is obtained called the horizontal velocity curve. In streams of symmetrical section this is a curve symmetrical about the centre line of the stream. The velocity varies little near the centre of the stream, but very rapidly near the banks. In unsymmetrical sections the greatest velocity is at the point where the stream is deepest, and the general form of the horizontal velocity curve is roughly similar to the section of the stream.

§ 102. Curves or Contours of Equal Velocity.—If velocities are observed at a number of points at different widths and depths in a stream, it is possible to draw curves on the cross section through points at which the velocity is the same. These represent contours of a solid, the volume of which is the discharge of the stream per second. Fig. 105 shows the vertical and horizontal velocity curves and the contours of equal velocity in a rectangular channel, from one of Bazin’s gaugings.

§ 103. Experimental Observations on the Vertical Velocity Curve.—A preliminary difficulty arises in observing the velocity at a given point in a stream because the velocity rapidly varies, the motion not being strictly steady. If an average of several velocities at the same point is taken, or the average velocity for a sensible period of time, this average is found to be constant. It may be inferred that though the velocity at a point fluctuates about a mean value, the fluctuations being due to eddying motions superposed on the general motion of the stream, yet these fluctuations produce effects which disappear in the mean of a series of observations and, in calculating the volume of flow, may be disregarded.

Fig. 105.

In the next place it is found that in most of the best observations on the velocity in streams, the greatest velocity at any vertical is found not at the surface but at some distance below it. In various river gaugings the depth d_z at the centre of the stream has been found to vary from 0 to 0.3d.

§ 104. Influence of the Wind.—In the experiments on the Mississippi the vertical velocity curve in calm weather was found to agree fairly with a parabola, the greatest velocity being at 3/10ths of the depth of the stream from the surface. With a wind blowing down stream the surface velocity is increased, and the axis of the parabola approaches the surface. On the contrary, with a wind blowing up stream the surface velocity is diminished, and the axis of the parabola is lowered, sometimes to half the depth of the stream. The American observers drew from their observations the conclusion that there was an energetic retarding action at the surface of a stream like that due to the bottom and sides. If there were such a retarding action the position of the filament of maximum velocity below the surface would be explained.

It is not difficult to understand that a wind acting on surface ripples or waves should accelerate or retard the surface motion of the stream, and the Mississippi results may be accepted so far as showing that the surface velocity of a stream is variable when the mean velocity of the stream is constant. Hence observations of surface velocity by floats or otherwise should only be made in very calm weather. But it is very difficult to suppose that, in still air, there is a resistance at the free surface of the stream at all analogous to that at the sides and bottom. Further, in very careful experiments, P. P. Boileau found the maximum velocity, though raised a little above its position for calm weather, still at a considerable distance below the surface, even when the wind was blowing down stream with a velocity greater than that of the stream, and when the action of the air must have been an accelerating and not a retarding action. A much more probable explanation of the diminution of the velocity at and near the free surface is that portions of water, with a diminished velocity from retardation by the sides or bottom, are thrown off in eddying masses and mingle with the rest of the stream. These eddying masses modify the velocity in all parts of the stream, but have their greatest influence at the free surface. Reaching the free surface they spread out and remain there, mingling with the water at that level and diminishing the velocity which would otherwise be found there.

Influence of the Wind on the Depth at which the Maximum Velocity is found.—In the gaugings of the Mississippi the vertical velocity curve was found to agree well with a parabola having a horizontal axis at some distance below the water surface, the ordinate of the parabola at the axis being the maximum velocity of the section. During the gaugings the force of the wind was registered on a scale ranging from 0 for a calm to 10 for a hurricane. Arranging the velocity curves in three sets—(1) with the wind blowing up stream, (2) with the wind blowing down stream, (3) calm or wind blowing across stream—it was found that an upstream wind lowered, and a down-stream wind raised, the axis of the parabolic velocity curve. In calm weather the axis was at 3/10ths of the total depth from the surface for all conditions of the stream.

Let h′ be the depth of the axis of the parabola, m the hydraulic mean depth, f the number expressing the force of the wind, which may range from +10 to −10, positive if the wind is up stream, negative if it is down stream. Then Humphreys and Abbot find their results agree with the expression

h′ / m = 0.317 ± 0.06f.

Fig. 106 shows the parabolic velocity curves according to the American observers for calm weather, and for an up- or down-stream wind of a force represented by 4.

Fig. 106.

It is impossible at present to give a theoretical rule for the vertical velocity curve, but in very many gaugings it has been found that a parabola with horizontal axis fits the observed results fairly well. The mean velocity on any vertical in a stream varies from 0.85 to 0.92 of the surface velocity at that vertical, and on the average if v₀ is the surface and v_m the mean velocity at a vertical v_m = 6/7v₀, a result useful in float gauging. On any vertical there is a point at which the velocity is equal to the mean velocity, and if this point were known it would be useful in gauging. Humphreys and Abbot in the Mississippi found the mean velocity at 0.66 of the depth; G. H. L. Hagen and H. Heinemann at 0.56 to 0.58 of the depth. The mean of observations by various observers gave the mean velocity at from 0.587 to 0.62 of the depth, the average of all being almost exactly 0.6 of the depth. The mid-depth velocity is therefore nearly equal to, but a little greater than, the mean velocity on a vertical. If v_md is the mid-depth velocity, then on the average v_m = 0.98v_md.

§ 105. Mean Velocity on a Vertical from Two Velocity Observations.—A. J. C. Cunningham, in gaugings on the Ganges canal, found the following useful results. Let v₀ be the surface, v_m the mean, and v_xd the velocity at the depth xd; then

v_m = 1/4 (v₀+3v_2/3d )
  = 1/2 (v_.211^d+v_.789^d ).

§ 106. Ratio of Mean to Greatest Surface Velocity, for the whole Cross Section in Trapezoidal Channels.—It is often very important to be able to deduce the mean velocity, and thence the discharge, from observation of the greatest surface velocity. The simplest method of gauging small streams and channels is to observe the greatest surface velocity by floats, and thence to deduce the mean velocity. In general in streams of fairly regular section the mean velocity for the whole section varies from 0.7 to 0.85 of the greatest surface velocity. For channels not widely differing from those experimented on by Bazin, the expression obtained by him for the ratio of surface to mean velocity may be relied on as at least a good approximation to the truth. Let v₀ be the greatest surface velocity, v_m the mean velocity of the stream. Then, according to Bazin,

v_m = v₀ − 25.4 √ (mi).

But

v_m = c √ (mi),

where c is a coefficient, the values of which have been already given in the table in § 98. Hence

v_m = cv₀ / (c + 25.4).

Values of Coefficient c/(c + 25.4) in the Formula v_m = cv₀/(c + 25.4).

Hydraulic  Mean Depth  = m.	Very Smooth  Channels.  Cement.	Smooth Channels. Ashlar or  Brickwork.	Rough  Channels.  Rubble Masonry.	Very Rough  Channels. Canals in Earth.	Channels  encumbered  with Detritus.
0.25	.83	.79	.69	.51	.42
0.5	.84	.81	.74	.58	.50
0.75	.84	.82	.76	.63	.55
1.0	.85	..	.77	.65	.58
2.0	..	.83	.79	.71	.64
3.0	..	..	.80	.73	.67
4.0	..	..	.81	.75	.70
5.0	..	..	..	.76	.71
6.0	..	.84	..	.77	.72
7.0	..	..	..	.78	.73
8.0	..	..	..	..	..
9.0	..	..	.82	..	.74
10.0	..	..	..	..	..
15.0	..	..	..	.79	.75
20.0	..	..	..	.80	.76
30.0	..	..	.82	..	.77
40.0	..	..	..	..	..
50.0	..	..	..	..	..
∞	..	..	..	..	.79

Fig. 107.

§ 107. River Bends.—In rivers flowing in alluvial plains, the windings which already exist tend to increase in curvature by the scouring away of material from the outer bank and the deposition of detritus along the inner bank. The sinuosities sometimes increase till a loop is formed with only a narrow strip of land between the two encroaching branches of the river. Finally a “cut off” may occur, a waterway being opened through the strip of land and the loop left separated from the stream, forming a horseshoe shaped lagoon or marsh. Professor James Thomson pointed out (Proc. Roy. Soc., 1877, p. 356; Proc. Inst. of Mech. Eng., 1879, p. 456) that the usual supposition is that the water tending to go forwards in a straight line rushes against the outer bank and scours it, at the same time creating deposits at the inner bank. That view is very far from a complete account of the matter, and Professor Thomson gave a much more ingenious account of the action at the bend, which he completely confirmed by experiment.

 

Fig. 108.

When water moves round a circular curve under the action of gravity only, it takes a motion like that in a free vortex. Its velocity is greater parallel to the axis of the stream at the inner than at the outer side of the bend. Hence the scouring at the outer side and the deposit at the inner side of the bend are not due to mere difference of velocity of flow in the general direction of the stream; but, in virtue of the centrifugal force, the water passing round the bend presses outwards, and the free surface in a radial cross section has a slope from the inner side upwards to the outer side (fig. 108). For the greater part of the water flowing in curved paths, this difference of pressure produces no tendency to transverse motion. But the water immediately in contact with the rough bottom and sides of the channel is retarded, and its centrifugal force is insufficient to balance the pressure due to the greater depth at the outside of the bend. It therefore flows inwards towards the inner side of the bend, carrying with it detritus which is deposited at the inner bank. Conjointly with this flow inwards along the bottom and sides, the general mass of water must flow outwards to take its place. Fig. 107 shows the directions of flow as observed in a small artificial stream, by means of light seeds and specks of aniline dye. The lines CC show the directions of flow immediately in contact with the sides and bottom. The dotted line AB shows the direction of motion of floating particles on the surface of the stream.

§ 108. Discharge of a River when flowing at different Depths.—When frequent observations must be made on the flow of a river or canal, the depth of which varies at different times, it is very convenient to have to observe the depth only. A formula can be established giving the flow in terms of the depth. Let Q be the discharge in cubic feet per second; H the depth of the river in some straight and uniform part. Then Q = aH + bH², where the constants a and b must be found by preliminary gaugings in different conditions of the river. M. C. Moquerey found for part of the upper Saône, Q = 64.7H + 8.2H² in metric measures, or Q = 696H + 26.8H² in English measures.

§ 109. Forms of Section of Channels.—The simplest form of section for channels is the semicircular or nearly semicircular channel (fig. 109), a form now often adopted from the facility with which it can be executed in concrete. It has the advantage that the rubbing surface is less in proportion to the area than in any other form.

Fig. 109.

Wooden channels or flumes, of which there are examples on a large scale in America, are rectangular in section, and the same form is adopted for wrought and cast-iron aqueducts. Channels built with brickwork or masonry may be also rectangular, but they are often trapezoidal, and are always so if the sides are pitched with masonry laid dry. In a trapezoidal channel, let b (fig. 110) be the bottom breadth, b₀ the top breadth, d the depth, and let the slope of the sides be n horizontal to 1 vertical. Then the area of section is Ω = (b + nd) d = (b₀ − nd) d, and the wetted perimeter χ = b + 2d √ (n² + 1).

Fig. 110.

When a channel is simply excavated in earth it is always originally trapezoidal, though it becomes more or less rounded in course of time. The slope of the sides then depends on the stability of the earth, a slope of 2 to 1 being the one most commonly adopted.

Figs. 111, 112 show the form of canals excavated in earth, the former being the section of a navigation canal and the latter the section of an irrigation canal.

§ 110. Channels of Circular Section.—The following short table facilitates calculations of the discharge with different depths of water in the channel. Let r be the radius of the channel section; then for a depth of water = κr, the hydraulic mean radius is μr and the area of section of the waterway νr ², where κ, μ, and ν have the following values:—

Depth of water in terms of radius	$\scriptstyle {\left.{\begin{matrix}\ \\\ \end{matrix}}\right\}\,}$κ =	.01	.05	.10	.15	.20	.25	.30	.35	.40	.45	.50	.55	.60	.65	.70	.75	.80	.85	.90	.95	1.0
Hydraulic mean depth in terms of radius	$\scriptstyle {\left.{\begin{matrix}\ \\\ \end{matrix}}\right\}\,}$μ =	.00668	.0321	.0523	.0963	.1278	.1574	.1852	.2142	.242	.269	.293	.320	.343	.365	.387	.408	.429	.449	.466	.484	.500
Waterway in terms of square of radius	$\scriptstyle {\left.{\begin{matrix}\ \\\ \end{matrix}}\right\}\,}$ν =	.00189	.0211	.0598	.1067	.1651	.228	.294	.370	.450	.532	.614	.709	.795	.885	.979	1.075	1.175	1.276	1.371	1.470	1.571

Fig. 111.—Scale 20 ft. = 1 in.

Fig. 112.—Scale 80 ft. = 1 in.

Fig. 113.

§ 111. Egg-Shaped Channels or Sewers.—In sewers for discharging storm water and house drainage the volume of flow is extremely variable; and there is a great liability for deposits to be left when the flow is small, which are not removed during the short periods when the flow is large. The sewer in consequence becomes choked. To obtain uniform scouring action, the velocity of flow should be constant or nearly so; a complete uniformity of velocity cannot be obtained with any form of section suitable for sewers, but an approximation to uniform velocity is obtained by making the sewers of oval section. Various forms of oval have been suggested, the simplest being one in which the radius of the crown is double the radius of the invert, and the greatest width is two-thirds the height. The section of such a sewer is shown in fig. 113, the numbers marked on the figure being proportional numbers.

§ 112. Problems on Channels in which the Flow is Steady and at Uniform Velocity.—The general equations given in §§ 96, 98 are

     ζ = α(1 + β/m);

(1)

ζv²/2g = mi;   

(2)

Q = Ωv.

(3)

Problem I.—Given the transverse section of stream and discharge, to find the slope. From the dimensions of the section find Ω and m; from (1) find ζ, from (3) find v, and lastly from (2) find i.

Problem II.—Given the transverse section and slope, to find the discharge. Find v from (2), then Q from (3).

Problem III.—Given the discharge and slope, and either the breadth, depth, or general form of the section of the channel, to determine its remaining dimensions. This must generally be solved by approximations. A breadth or depth or both are chosen, and the discharge calculated. If this is greater than the given discharge, the dimensions are reduced and the discharge recalculated.

Fig. 114.

Since m lies generally between the limits m = d and m = 1/2d, where d is the depth of the stream, and since, moreover, the velocity varies as √ (m) so that an error in the value of m leads only to a much less error in the value of the velocity calculated from it, we may proceed thus. Assume a value for m, and calculate v from it. Let v₁ be this first approximation to v. Then Q/v₁ is a first approximation to Ω, say Ω₁. With this value of Ω design the section of the channel; calculate a second value for m; calculate from it a second value of v, and from that a second value for Ω. Repeat the process till the successive values of m approximately coincide.

§ 113. Problem IV. Most Economical Form of Channel for given Side Slopes.—Suppose the channel is to be trapezoidal in section (fig. 114), and that the sides are to have a given slope. Let the longitudinal slope of the stream be given, and also the mean velocity. An infinite number of channels could be found satisfying the foregoing conditions. To render the problem determinate, let it be remembered that, since for a given discharge Ω∞ ∛χ, other things being the same, the amount of excavation will be least for that channel which has the least wetted perimeter. Let d be the depth and b the bottom width of the channel, and let the sides slope n horizontal to 1 vertical (fig. 114), then

Ω = (b + nd) d;

χ = b + 2d √ (n² + 1).

Both Ω and χ are to be minima. Differentiating, and equating to zero.

(db/dd + n) d + b + nd = 0,
db/dd + 2 √ (n² + 1) = 0;

eliminating db/dd,

{n − 2√ (n² + 1)} d + b + nd = 0;
b = 2 {√ (n² + 1) − n} d.

But

Ω / χ = (b + nd) d / {b + 2d √ (n² + 1)}.

Inserting the value of b,

m = Ω/χ = {2d √ (n² + 1) − nd} / {4d √ (n² + 1) − 2nd} = 1/2 d.

That is, with given side slopes, the section is least for a given discharge when the hydraulic mean depth is half the actual depth.

A simple construction gives the form of the channel which fulfils this condition, for it can be shown that when m = 1/2d the sides of the channel are tangential to a semicircle drawn on the water line.

Since

Ω / χ = 1/2 d,

therefore

Ω = 1/2 χd.

(1)

Let ABCD be the channel (fig. 115); from E the centre of AD drop perpendiculars EF, EG, EH on the sides.

Let

AB = CD = a; BC = b; EF = EH = c; and EG = d.

Ω = area AEB + BEC + CED,
 = ac + 1/2 bd.
χ  = 2a + b.

Putting these values in (1),

ac + 1/2 bd = (a + 1/2 b) d; and hence c = d.

Fig. 115.

Fig. 116.

That is, EF, EG, EH are all equal, hence a semicircle struck from E with radius equal to the depth of the stream will pass through F and H and be tangential to the sides of the channel.

To draw the channel, describe a semicircle on a horizontal line with radius = depth of channel. The bottom will be a horizontal tangent of that semicircle, and the sides tangents drawn at the required side slopes.

The above result may be obtained thus (fig. 116):—

χ = b + 2d / sin β.

(1)

Ω = d (b + d cot β);

Ω/d = b + d cot β;

(2)

Ω/d ² = b/d + cot β.

(3)

From (1) and (2),

χ = Ω / d − d cot β + 2d / sin β.

This will be a minimum for

dχ / dd = Ω / d ² + cot β − 2 / sin β = 0,

or

Ω/d ² = 2 cosec. β − cot β.

(4)

or

d = √ {Ω sin β / (2 − cos β)}.

From (3) and (4),

b/d = 2 (1 − cos β) / sin β = 2 tan 1/2 β.

Proportions of Channels of Maximum Discharge for given Area and Side Slopes.
Depth of channel = d; Hydraulic mean depth = 1/2d;

Area of section = Ω.

	Inclination  of Sides to  Horizon.	Ratio of  Side Slopes.	Area of  Section Ω.	Bottom  Width.	Top width =  twice length of each Side Slope.
Semicircle	..	..	1.571d 2	0	2d
Semi-hexagon	60°   0′	3   : 5	1.732d 2	1.155d	2.310d
Semi-square	90°   0′	0   : 1	2d 2	2d	2d
	75°  58′	1   : 4	1.812d 2	1.562d	2.062d
	63°  26′	1   : 2	1.736d 2	1.236d	2.236d
	53°   8′	3   : 4	1.750d 2	d	2.500d
	45°   0′	1   : 1	1.828d 2	0.828d	2.828d
	38°  40′	11/4 : 1	1.952d 2	0.702d	3.202d
	33°  42′	11/2 : 1	2.106d 2	0.606d	3.606d
	29°  44′	13/4 : 1	2.282d 2	0.532d	4.032d
	26°  34′	2  : 1	2.472d 2	0.472d	4.472d
	23°  58′	21/4 : 1	2.674d 2	0.424d	4.924d
	21°  48′	21/2 : 1	2.885d 2	0.385d	5.385d
	19°  58′	23/4 : 1	3.104d 2	0.354d	5.854d
	18°  26′	3   : 1	3.325d 2	0.325d	6.325d

Half the top width is the length of each side slope. The wetted
perimeter is the sum of the top and bottom widths.

§ 114. Form of Cross Section of Channel in which the Mean Velocity is Constant with Varying Discharge.—In designing waste channels from canals, and in some other cases, it is desirable that the mean velocity should be restricted within narrow limits with very different volumes of discharge. In channels of trapezoidal form the velocity increases and diminishes with the discharge. Hence when the discharge is large there is danger of erosion, and when it is small of silting or obstruction by weeds. A theoretical form of section for which the mean velocity would be constant can be found, and, although this is not very suitable for practical purposes, it can be more or less approximated to in actual channels.

Fig. 117.

Let fig. 117 represent the cross section of the channel. From the symmetry of the section, only half the channel need be considered. Let obac be any section suitable for the minimum flow, and let it be required to find the curve beg for the upper part of the channel so that the mean velocity shall be constant. Take o as origin of coordinates, and let de, fg be two levels of the water above ob.

Let ob = b/2; de = y, fg = y + dy, od = x, of = x + dx; eg = ds.

The condition to be satisfied is that

v = c √ (mi)

should be constant, whether the water-level is at ob, de, or fg. Consequently

m = constant = k

for all three sections, and can be found from the section obac. Hence also

Increment of section/Increment of perimeter = ydx/ds = k.

y²dx² = k²ds² = k²(dx²+dy²) and dx = kdy / √ (y² − k²).

Integrating,

x = k log_ε {y + √ (y² − k²)} + constant;

and, since y = b/2 when x = 0,

x = k log_ε [{y + √ (y² − k²)} / {1/2 b + √ (1/4 b² − k²) }].

Assuming values for y, the values of x can be found and the curve drawn.

The figure has been drawn for a channel the minimum section of which is a half hexagon of 4 ft. depth. Hence k = 2; b = 9.2; the rapid flattening of the side slopes is remarkable.

Steady Motion of Water in Open Channels of Varying Cross Section and Slope

§ 115. In every stream the discharge of which is constant, or may be regarded as constant for the time considered, the velocity at different places depends on the slope of the bed. Except at certain exceptional points the velocity will be greater as the slope of the bed is greater, and, as the velocity and cross section of the stream vary inversely, the section of the stream will be least where the velocity and slope are greatest. If in a stream of tolerably uniform slope an obstruction such as a weir is built, that will cause an alteration of flow similar to that of an alteration of the slope of the bed for a greater or less distance above the weir, and the originally uniform cross section of the stream will become a varied one. In such cases it is often of much practical importance to determine the longitudinal section of the stream.

The cases now considered will be those in which the changes of velocity and cross section are gradual and not abrupt, and in which the only internal work which needs to be taken into account is that due to the friction of the stream bed, as in cases of uniform motion. Further, the motion will be supposed to be steady, the mean velocity at each given cross section remaining constant, though it varies from section to section along the course of the stream.

Fig. 118.

Let fig. 118 represent a longitudinal section of the stream, A₀A₁ being the water surface, B₀B₁ the stream bed. Let A₀B₀, A₁B₁ be cross sections normal to the direction of flow. Suppose the mass of water A₀B₀A₁B₁ comes in a short time θ to C₀D₀C₁D₁, and let the work done on the mass be equated to its change of kinetic energy during that period. Let l be the length A₀A₁ of the portion of the stream considered, and z the fall, of surface level in that distance. Let Q be the discharge of the stream per second.

Fig. 119.

Change of Kinetic Energy.—At the end of the time θ there are as many particles possessing the same velocities in the space C₀D₀A₁B₁ as at the beginning. The change of kinetic energy is therefore the difference of the kinetic energies of A₀B₀C₀D₀ and A₁B₁C₁D₁.

Let fig. 119 represent the cross section A₀B₀, and let ω be a small element of its area at a point where the velocity is v. Let Ω₀ be the whole area of the cross section and u₀ the mean velocity for the whole cross section. From the definition of mean velocity we have

u₀ = Σ ωv / Ω₀.

Let v = u₀ + w, where w is the difference between the velocity at the small element ω and the mean velocity. For the whole cross section, Σωw = 0.

The mass of fluid passing through the element of section ω, in θ seconds, is (G/g) ωvθ, and its kinetic energy is (G/2g) ωv³θ. For the whole section, the kinetic energy of the mass A₀B₀C₀D₀ passing in θ seconds is

(Gθ / 2g) Σωv³ = (Gθ/2g) Σω (u₀³ + 3u₀²w + 3u₀² + w³),
= (Gθ / 2g) {u₀³Ω + Σωw² (3u₀ + w)}.

The factor 3u₀ + w is equal to 2u₀ + v, a quantity necessarily positive. Consequently Σωv³ > Ω₀u₀³, and consequently the kinetic energy of A₀B₀C₀D₀ is greater than

(Gθ / 2g) Ω₀u₀³ or (Gθ) / 2g) Qu₀²,

which would be its value if all the particles passing the section had the same velocity u₀. Let the kinetic energy be taken at

α (Gθ / 2g) Ω₀u₀³ = α (Gθ / 2g) Qu₀²,

where α is a corrective factor, the value of which was estimated by J. B. C. J. Bélanger at 1.1.^[6] Its precise value is not of great importance.

In a similar way we should obtain for the kinetic energy of A₁B₁C₁D₁ the expression

α (Gθ / 2g) Ω₁u₁³ = α (Gθ / 2g) Qu₁²,

where Ω₁, u₁ are the section and mean velocity at A₁B₁, and where a may be taken to have the same value as before without any important error.

Hence the change of kinetic energy in the whole mass A₀B₀A₁B₁ in θ seconds is

α (Gθ / 2g)Q(u12−u02).

(1)

Motive Work of the Weight and Pressures.—Consider a small filament a₀a₁ which comes in θ seconds to c₀c₁. The work done by gravity during that movement is the same as if the portion a₀c₀ were carried to a₁c₁. Let dQθ be the volume of a₀c₀ or a₁c₁, and y₀, y₁ the depths of a₀, a₁ from the surface of the stream. Then the volume dQθ or GdQθ pounds falls through a vertical height z + y₁ − y₀, and the work done by gravity is

GdQθ(z + y₁ − y₀).

Putting p_a for atmospheric pressure, the whole pressure per unit of area at a₀ is Gy₀ + p_a, and that at a₁ is −(Gy₁ + p_a). The work of these pressures is

G(y₀ + p_a/G − y₁ − p_a/G) dQθ = G (y₀ − y₁) dQθ.

Adding this to the work of gravity, the whole work is GzdQθ; or, for the whole cross section,

GzQθ.

(2)

Work expended in Overcoming the Friction of the Stream Bed.—Let A′B′, A″B″ be two cross sections at distances s and s + ds from A₀B₀. Between these sections the velocity may be treated as uniform, because by hypothesis the changes of velocity from section to section are gradual. Hence, to this short length of stream the equation for uniform motion is applicable. But in that case the work in overcoming the friction of the stream bed between A′B′ and A″B″ is

GQθζ(u² / 2g) (χ / Ω) ds,

where u, χ, Ω are the mean velocity, wetted perimeter, and section at A′B′. Hence the whole work lost in friction from A₀B₀ to A₁B₁ will be

GQθ $\int _{0}^{l}$ζ(u2 / 2g) (χ/Ω)ds.

(3)

Equating the work given in (2) and (3) to the change of kinetic energy given in (1),

α(GQθ / 2g) (u₁² − u₀²) = GQzθ − GQθ $\int _{0}^{l}$ ζ(u² / 2g) (χ / Ω) ds;

∴ z = α(u₁² − u₀²) / 2g + $\int _{0}^{l}$ ζ(u² / 2g) (χ / Ω) ds.

Fig. 120.

§ 116. Fundamental Differential Equation of Steady Varied Motion.—Suppose the equation just found to be applied to an indefinitely short length ds of the stream, limited by the end sections ab, a₁b₁, taken for simplicity normal to the stream bed (fig. 120). For that short length of stream the fall of surface level, or difference of level of a and a₁, may be written dz. Also, if we write u for u₀, and u + du for u₁, the term (u₀² − u₁²)/2g becomes udu/g. Hence the equation applicable to an indefinitely short length of the stream is

dz = udu/g + (χ/Ω) ζ(u²/2g) ds.

(1)

From this equation some general conclusions may be arrived at as to the form of the longitudinal section of the stream, but, as the investigation is somewhat complicated, it is convenient to simplify it by restricting the conditions of the problem.

Modification of the Formula for the Restricted Case of a Stream flowing in a Prismatic Stream Bed of Constant Slope.—Let i be the constant slope of the bed. Draw ad parallel to the bed, and ac horizontal. Then dz is sensibly equal to a′c. The depths of the stream, h and h + dh, are sensibly equal to ab and a′b′, and therefore dh = a′d. Also cd is the fall of the bed in the distance ds, and is equal to ids. Hence

dz = a′c = cd − a′d = i ds − dh.

(2)

Since the motion is steady—

Q = Ωu = constant.

Differentiating,

Ω du + u dΩ = 0;

∴ du = −u dΩ/Ω.

Let x be the width of the stream, then dΩ = xdh very nearly. Inserting this value,

du = −(ux / Ω) dh.

(3)

Putting the values of du and dz found in (2) and (3) in equation (1),

ids − dh = −(u²x / gΩ) dh + (χ / Ω) ζ(u² / 2g) ds.

dh/ds = {i − (χ/Ω) ζ(u2/2g)} / {1 − (u2/g) (x/Ω)}.

(4)

Further Restriction to the Case of a Stream of Rectangular Section and of Indefinite Width.—The equation might be discussed in the form just given, but it becomes a little simpler if restricted in the way just stated. For, if the stream is rectangular, χh = Ω, and if χ is large compared with h, Ω/χ = xh/x = h nearly. Then equation (4) becomes

dh/ds = i (1 − ζu² / 2gih) / (1 − u²/gh).

(5)

§ 117. General Indications as to the Form of Water Surface furnished by Equation (5).—Let A₀A₁ (fig. 121) be the water surface, B₀B₁ the bed in a longitudinal section of the stream, and ab any section at a distance s from B₀, the depth ab being h. Suppose B₀B₁, B₀A₀ taken as rectangular coordinate axes, then dh/ds is the trigonometric tangent of the angle which the surface of the stream at a makes with the axis B₀B₁. This tangent dh/ds will be positive, if the stream is increasing in depth in the direction B₀B₁; negative, if the stream is diminishing in depth from B₀ towards B₁. If dh/ds = 0, the surface of the stream is parallel to the bed, as in cases of uniform motion. But from equation (4)

dh/ds = 0, if i − (χ/Ω) ζ(u²/2g) = 0;

∴ ζ(u²/2g) = (Ω/χ) i = mi,

which is the well-known general equation for uniform motion, based on the same assumptions as the equation for varied steady motion now being considered. The case of uniform motion is therefore a limiting case between two different kinds of varied motion.

Fig. 121.

Consider the possible changes of value of the fraction

(1 − ζu² / 2gih) / (1 − u² / gh).

As h tends towards the limit 0, and consequently u is large, the numerator tends to the limit −∞. On the other hand if h = ∞, in which case u is small, the numerator becomes equal to 1. For a value H of h given by the equation

1 − ζu² / 2giH = 0,
H = ζu²/2gi,

we fall upon the case of uniform motion. The results just stated may be tabulated thus:—

For h = 0, H, > H, ∞,

the numerator has the value −∞, 0, > 0, 1.

Next consider the denominator. If h becomes very small, in which case u must be very large, the denominator tends to the limit −∞. As h becomes very large and u consequently very small, the denominator tends to the limit 1. For h = u²/g, or u = √ (gh), the denominator becomes zero. Hence, tabulating these results as before:—

For  h = 0, u²/g, > u²/g, ∞,

the denominator becomes −∞, 0, > 0, 1.

Fig. 122.

§ 118. Case 1.—Suppose h>u²/g, and also h>H, or the depth greater than that corresponding to uniform motion. In this case dh/ds is positive, and the stream increases in depth in the direction of flow. In fig. 122 let B₀B₁ be the bed, C₀C₁ a line parallel to the bed and at a height above it equal to H. By hypothesis, the surface A₀A₁ of the stream is above C₀C₁, and it has just been shown that the depth of the stream increases from B₀ towards B₁. But going up stream h approaches more and more nearly the value H, and therefore dh/ds approaches the limit 0, or the surface of the stream is asymptotic to C₀C₁. Going down stream h increases and u diminishes, the numerator and denominator of the fraction (1 − ζu²/2gih) / (1 − u²/gh) both tend towards the limit 1, and dh/ds to the limit i. That is, the surface of the stream tends to become asymptotic to a horizontal line D₀D₁.

The form of water surface here discussed is produced when the flow of a stream originally uniform is altered by the construction of a weir. The raising of the water surface above the level C₀C₁ is termed the backwater due to the weir.

Fig. 123.

§ 119. Case 2.—Suppose h > u²/g, and also h < H. Then dh/ds is negative, and the stream is diminishing in depth in the direction of flow. In fig. 123 let B₀B₁ be the stream bed as before; C₀C₁ a line drawn parallel to B₀B₁ at a height above it equal to H. By hypothesis the surface A₀A₁ of the stream is below C₀C₁, and the depth has just been shown to diminish from B₀ towards B₁. Going up stream h approaches the limit H, and dh/ds tends to the limit zero. That is, up stream A₀A₁ is asymptotic to C₀C₁. Going down stream h diminishes and u increases; the inequality h > u²/g diminishes; the denominator of the fraction (1 − ζu²/2gih) / (1 − u²/gh) tends to the limit zero, and consequently dh/ds tends to ∞. That is, down stream A₀A₁ tends to a direction perpendicular to the bed. Before, however, this limit was reached the assumptions on which the general equation is based would cease to be even approximately true, and the equation would cease to be applicable. The filaments would have a relative motion, which would make the influence of internal friction in the fluid too important to be neglected. A stream surface of this form may be produced if there is an abrupt fall in the bed of the stream (fig. 124).

Fig. 124.

On the Ganges canal, as originally constructed, there were abrupt falls precisely of this kind, and it appears that the lowering of the water surface and increase of velocity which such falls occasion, for a distance of some miles up stream, was not foreseen. The result was that, the velocity above the falls being greater than was intended, the bed was scoured and considerable damage was done to the works. “When the canal was first opened the water was allowed to pass freely over the crests of the overfalls, which were laid on the level of the bed of the earthen channel; erosion of bed and sides for some miles up rapidly followed, and it soon became apparent that means must be adopted for raising the surface of the stream at those points (that is, the crests of the falls). Planks were accordingly fixed in the grooves above the bridge arches, or temporary weirs were formed over which the water was allowed to fall; in some cases the surface of the water was thus raised above its normal height, causing a backwater in the channel above” (Crofton’s Report on the Ganges Canal, p. 14). Fig. 125 represents in an exaggerated form what probably occurred, the diagram being intended to represent some miles’ length of the canal bed above the fall. AA parallel to the canal bed is the level corresponding to uniform motion with the intended velocity of the canal. In consequence of the presence of the ogee fall, however, the water surface would take some such form as BB, corresponding to Case 2 above, and the velocity would be greater than the intended velocity, nearly in the inverse ratio of the actual to the intended depth. By constructing a weir on the crest of the fall, as shown by dotted lines, a new water surface CC corresponding to Case 1 would be produced, and by suitably choosing the height of the weir this might be made to agree approximately with the intended level AA.

Fig. 125.

§ 120. Case 3.—Suppose a stream flowing uniformly with a depth h < u²/g. For a stream in uniform motion ζu²/2g = mi, or if the stream is of indefinitely great width, so that m = H, then ζu²/2g = iH, and H = ζu²/2gi. Consequently the condition stated above involves that

ζu² / 2gi < u² / g, or that i > ζ/2.

If such a stream is interfered with by the construction of a weir which raises its level, so that its depth at the weir becomes h₁ > u²/g, then for a portion of the stream the depth h will satisfy the conditions h < u²/g and h > H, which are not the same as those assumed in the two previous cases. At some point of the stream above the weir the depth h becomes equal to u²/g, and at that point dh/ds becomes infinite, or the surface of the stream is normal to the bed. It is obvious that at that point the influence of internal friction will be too great to be neglected, and the general equation will cease to represent the true conditions of the motion of the water. It is known that, in cases such as this, there occurs an abrupt rise of the free surface of the stream, or a standing wave is formed, the conditions of motion in which will be examined presently.

It appears that the condition necessary to give rise to a standing wave is that i > ζ/2. Now ζ depends for different channels on the roughness of the channel and its hydraulic mean depth. Bazin calculated the values of ζ for channels of different degrees of roughness and different depths given in the following table, and the corresponding minimum values of i for which the exceptional case of the production of a standing wave may occur.

Nature of Bed of Stream.	Slope below  which a Standing  Wave is impossible in feet peer foot.	Standing Wave Formed.
		Slope in feet  per foot.	Least Depth  in feet.
Very smooth cemented surface	0.00147	0.002	0.262
		0.003	.098
		0.004	.065
Ashlar or brickwork	0.00186	0.003	.394
		0.004	.197
		0.006	.098
Rubble masonry	0.00235	0.004	1.181
		0.006	.525
		0.010	.262
Earth	0.00275	0.006	3.478
		0.010	1.542
		0.015	.919

Standing Waves

§ 121. The formation of a standing wave was first observed by Bidone. Into a small rectangular masonry channel, having a slope of 0.023 ft. per foot, he admitted water till it flowed uniformly with a depth of 0.2 ft. He then placed a plank across the stream which raised the level just above the obstruction to 0.95 ft. He found that the stream above the obstruction was sensibly unaffected up to a point 15 ft. from it. At that point the depth suddenly increased from 0.2 ft. to 0.56 ft. The velocity of the stream in the part unaffected by the obstruction was 5.54 ft. per second. Above the point where the abrupt change of depth occurred u² = 5.54² = 30.7, and gh = 32.2 × 0.2 = 6.44; hence u² was > gh. Just below the abrupt change of depth u = 5.54 × 0.2/0.56 = 1.97; u² = 3.88; and gh = 32.2 × 0.56 = 18.03; hence at this point u² < gh. Between these two points, therefore, u² = gh; and the condition for the production of a standing wave occurred.

Fig. 126.

The change of level at a standing wave may be found thus. Let fig. 126 represent the longitudinal section of a stream and ab, cd cross sections normal to the bed, which for the short distance considered may be assumed horizontal. Suppose the mass of water abcd to come to a′b′c′d′ in a short time t; and let u₀, u₁ be the velocities at ab and cd, Ω₀, Ω₁ the areas of the cross sections. The force causing change of momentum in the mass abcd estimated horizontally is simply the difference of the pressures on ab and cd. Putting h₀, h₁ for the depths of the centres of gravity of ab and cd measured down from the free water surface, the force is G (h ₀Ω₀ − h₁Ω₁) pounds, and the impulse in t seconds is G (h ₀Ω₀ − h₁Ω₁) t second pounds. The horizontal change of momentum is the difference of the momenta of cdc′d′ and aba′b′; that is,

(G/g) (Ω₁u₁² − Ω₀u₀²) t.

Hence, equating impulse and change of momentum,

G (h₀Ω₀ − h₁Ω₁) t = (G/g) (Ω₁u₁² − Ω₀u₀²) t;

∴ h₀Ω₀ − h₁Ω₁ = (Ω₁u₁² − Ω₀u₀²) / g.

(1)

For simplicity let the section be rectangular, of breadth B and depths H₀ and H₁, at the two cross sections considered; then h₀ = 1/2H₀, and h₁ = 1/2H₁. Hence

H₀² − H₁² = (2/g) (H₁u₁² − H₀u₀²).

But, since Ω₀u₀ = Ω₁u₁, we have

u₁² = u₀²H₀² / H₁²,

H₀² − H₁² = (2u₀²/g) (H₀²/H₁ − H₀).

(2)

This equation is satisfied if H₀ = H₁, which corresponds to the case of uniform motion. Dividing by H₀ − H₁, the equation becomes

(H₁/H₀) (H₀ + H₁) = 2u₀² / g;

(3)

∴ H₁ = √ (2u₀²H₀ / g + 1/4 H₀²) − 1/2 H₀.

(4)

In Bidone’s experiment u₀ = 5.54, and H₀ = 0.2. Hence H₁ = 0.52, which agrees very well with the observed height.

Fig. 127.

§ 122. A standing wave is frequently produced at the foot of a weir. Thus in the ogee falls originally constructed on the Ganges canal a standing wave was observed as shown in fig. 127. The water falling over the weir crest A acquired a very high velocity on the steep slope AB, and the section of the stream at B became very small. It easily happened, therefore, that at B the depth h < u²/g. In flowing along the rough apron of the weir the velocity u diminished and the depth h increased. At a point C, where h became equal to u²/g, the conditions for producing the standing wave occurred. Beyond C the free surface abruptly rose to the level corresponding to uniform motion with the assigned slope of the lower reach of the canal.

Fig. 128.

A standing wave is sometimes formed on the down stream side of bridges the piers of which obstruct the flow of the water. Some interesting cases of this kind are described in a paper on the “Floods in the Nerbudda Valley” in the Proc. Inst. Civ. Eng. vol. xxvii. p. 222, by A. C. Howden. Fig. 128 is compiled from the data given in that paper. It represents the section of the stream at pier 8 of the Towah Viaduct, during the flood of 1865. The ground level is not exactly given by Howden, but has been inferred from data given on another drawing. The velocity of the stream was not observed, but the author states it was probably the same as at the Gunjal river during a similar flood, that is 16.58 ft. per second. Now, taking the depth on the down stream face of the pier at 26 ft., the velocity necessary for the production of a standing wave would be u = √ (gh) = √ (32.2 × 26) = 29 ft. per second nearly. But the velocity at this point was probably from Howden’s statements 16.58 × 40/26 = 25.5 ft. per second, an agreement as close as the approximate character of the data would lead us to expect.

XI. ON STREAMS AND RIVERS

§ 123. Catchment Basin.—A stream or river is the channel for the discharge of the available rainfall of a district, termed its catchment basin. The catchment basin is surrounded by a ridge or watershed line, continuous except at the point where the river finds an outlet. The area of the catchment basin may be determined from a suitable contoured map on a scale of at least 1 in 100,000. Of the whole rainfall on the catchment basin, a part only finds its way to the stream. Part is directly re-evaporated, part is absorbed by vegetation, part may escape by percolation into neighbouring districts. The following table gives the relation of the average stream discharge to the average rainfall on the catchment basin (Tiefenbacher).

	Ratio of average Discharge to average Rainfall.	Loss by Evaporation, &c., in per cent of total Rainfall.
Cultivated land and spring-forming declivities.	3 to .33	67 to 70
Wooded hilly slopes.	.35 to .45	55 to 65
Naked unfissured mountains	.55 to .60	40 to 45

§ 124. Flood Discharge.—The flood discharge can generally only be determined by examining the greatest height to which floods have been known to rise. To produce a flood the rainfall must be heavy and widely distributed, and to produce a flood of exceptional height the duration of the rainfall must be so great that the flood waters of the most distant affluents reach the point considered, simultaneously with those from nearer points. The larger the catchment basin the less probable is it that all the conditions tending to produce a maximum discharge should simultaneously occur. Further, lakes and the river bed itself act as storage reservoirs during the rise of water level and diminish the rate of discharge, or serve as flood moderators. The influence of these is often important, because very heavy rain storms are in most countries of comparatively short duration. Tiefenbacher gives the following estimate of the flood discharge of streams in Europe:—

	Flood discharge of Streams per Second per Square Mile of Catchment Basin.
In flat country	8.7 to 12.5	cub. ft.
In hilly districts	17.5 to 22.5	”
In moderately mountainous districts	36.2 to 45.0	”
In very mountainous districts	50.0 to 75.0	”

It has been attempted to express the decrease of the rate of flood discharge with the increase of extent of the catchment basin by empirical formulae. Thus Colonel P. P. L. O’Connell proposed the formula y = M √ x, where M is a constant called the modulus of the river, the value of which depends on the amount of rainfall, the physical characters of the basin, and the extent to which the floods are moderated by storage of the water. If M is small for any given river, it shows that the rainfall is small, or that the permeability or slope of the sides of the valley is such that the water does not drain rapidly to the river, or that lakes and river bed moderate the rise of the floods. If values of M are known for a number of rivers, they may be used in inferring the probable discharge of other similar rivers. For British rivers M varies from 0.43 for a small stream draining meadow land to 37 for the Tyne. Generally it is about 15 or 20. For large European rivers M varies from 16 for the Seine to 67.5 for the Danube. For the Nile M = 11, a low value which results from the immense length of the Nile throughout which it receives no affluent, and probably also from the influence of lakes. For different tributaries of the Mississippi M varies from 13 to 56. For various Indian rivers it varies from 40 to 303, this variation being due to the great variations of rainfall, slope and character of Indian rivers.

In some of the tank projects in India, the flood discharge has been calculated from the formula D = C³√ n², where D is the discharge in cubic yards per hour from n square miles of basin. The constant C was taken = 61,523 in the designs for the Ekrooka tank, = 75,000 on Ganges and Godavery works, and = 10,000 on Madras works.

Fig. 129.

Fig. 130.

§ 125. Action of a Stream on its Bed.—If the velocity of a stream exceeds a certain limit, depending on its size, and on the size, heaviness, form and coherence of the material of which its bed is composed, it scours its bed and carries forward the materials. The quantity of material which a given stream can carry in suspension depends on the size and density of the particles in suspension, and is greater as the velocity of the stream is greater. If in one part of its course the velocity of a stream is great enough to scour the bed and the water becomes loaded with silt, and in a subsequent part of the river’s course the velocity is diminished, then part of the transported material must be deposited. Probably deposit and scour go on simultaneously over the whole river bed, but in some parts the rate of scour is in excess of the rate of deposit, and in other parts the rate of deposit is in excess of the rate of scour. Deep streams appear to have the greatest scouring power at any given velocity. It is possible that the difference is strictly a difference of transporting, not of scouring action. Let fig. 129 represent a section of a stream. The material lifted at a will be diffused through the mass of the stream and deposited at different distances down stream. The average path of a particle lifted at a will be some such curve as abc, and the average distance of transport each time a particle is lifted will be represented by ac. In a deeper stream such as that in fig. 130, the average height to which particles are lifted, and, since the rate of vertical fall through the water may be assumed the same as before, the average distance a′c′ of transport will be greater. Consequently, although the scouring action may be identical in the two streams, the velocity of transport of material down stream is greater as the depth of the stream is greater. The effect is that the deep stream excavates its bed more rapidly than the shallow stream.

§ 126. Bottom Velocity at which Scour commences.—The following bottom velocities were determined by P. L. G. Dubuat to be the maximum velocities consistent with stability of the stream bed for different materials.

Darcy and Bazin give, for the relation of the mean velocity v_m and bottom velocity v_b.

v_m = v_b + 10.87 √ (mi).

But

√ mi = v_m √ (ζ / 2g);

∴ v_m = v_b / (1 − 10.87 √ (ζ / 2g)).

Taking a mean value for ζ, we get

v_m = 1.312 v_b,

and from this the following values of the mean velocity are obtained:—

	Bottom Velocity  = vb.	Mean Velocity  = vm.
1. Soft earth	0.25	.33
2. Loam	0.50	.65
3. Sand	1.00	1.30
4. Gravel	2.00	2.62
5. Pebbles	3.40	4.46
6. Broken stone, flint	4.00	5.25
7. Chalk, soft shale	5.00	6.56
8. Rock in beds	6.00	7.87
9. Hard rock.	10.00	13.12

The following table of velocities which should not be exceeded in channels is given in the Ingenieurs Taschenbuch of the Verein “Hütte”:—

	Surface  Velocity.	Mean  Velocity.	Bottom  Velocity.
Slimy earth or brown clay	.49	.36	.26
Clay	.98	.75	.52
Firm sand	1.97	1.51	1.02
Pebbly bed	4.00	3.15	2.30
Boulder bed	5.00	4.03	3.08
Conglomerate of slaty fragments	7.28	6.10	4.90
Stratified rocks	8.00	7.45	6.00
Hard rocks	14.00	12.15	10.36

§ 127. Regime of a River Channel.—A river channel is said to be in a state of regime, or stability, when it changes little in draught or form in a series of years. In some rivers the deepest part of the channel changes its position perpetually, and is seldom found in the same place in two successive years. The sinuousness of the river also changes by the erosion of the banks, so that in time the position of the river is completely altered. In other rivers the change from year to year is very small, but probably the regime is never perfectly stable except where the rivers flow over a rocky bed.

Fig. 131.

If a river had a constant discharge it would gradually modify its bed till a permanent regime was established. But as the volume discharged is constantly changing, and therefore the velocity, silt is deposited when the velocity decreases, and scour goes on when the velocity increases in the same place. When the scouring and silting are considerable, a perfect balance between the two is rarely established, and hence continual variations occur in the form of the river and the direction of its currents. In other cases, where the action is less violent, a tolerable balance may be established, and the deepening of the bed by scour at one time is compensated by the silting at another. In that case the general regime is permanent, though alteration is constantly going on. This is more likely to happen if by artificial means the erosion of the banks is prevented. If a river flows in soil incapable of resisting its tendency to scour it is necessarily sinuous (§ 107), for the slightest deflection of the current to either side begins an erosion which increases progressively till a considerable bend is formed. If such a river is straightened it becomes sinuous again unless its banks are protected from scour.

§ 128. Longitudinal Section of River Bed.—The declivity of rivers decreases from source to mouth. In their higher parts rapid and torrential, flowing over beds of gravel or boulders, they enlarge in volume by receiving affluent streams, their slope diminishes, their bed consists of smaller materials, and finally they reach the sea. Fig. 131 shows the length in miles, and the surface fall in feet per mile, of the Tyne and its tributaries.

The decrease of the slope is due to two causes. (1) The action of the transporting power of the water, carrying the smallest debris the greatest distance, causes the bed to be less stable near the mouth than in the higher parts of the river; and, as the river adjusts its slope to the stability of the bed by scouring or increasing its sinuousness when the slope is too great, and by silting or straightening its course if the slope is too small, the decreasing stability of the bed would coincide with a decreasing slope. (2) The increase of volume and section of the river leads to a decrease of slope; for the larger the section the less slope is necessary to ensure a given velocity.

 

Fig. 132.

The following investigation, though it relates to a purely arbitrary case, is not without interest. Let it be assumed, to make the conditions definite—(1) that a river flows over a bed of uniform resistance to scour, and let it be further assumed that to maintain stability the velocity of the river in these circumstances is constant from source to mouth; (2) suppose the sections of the river at all points are similar, so that, b being the breadth of the river at any point, its hydraulic mean depth is ab and its section is cb², where a and c are constants applicable to all parts of the river; (3) let us further assume that the discharge increases uniformly in consequence of the supply from affluents, so that, if l is the length of the river from its source to any given point, the discharge there will be kl, where k is another constant applicable to all points in the course of the river.

Let AB (fig. 132) be the longitudinal section of the river, whose source is at A; and take A for the origin of vertical and horizontal coordinates. Let C be a point whose ordinates are x and y, and let the river at C have the breadth b, the slope i, and the velocity v.

Since velocity × area of section = discharge, vcb² = kl, or b = √ (kl/cv).

Hydraulic mean depth = ab = a √ (kl/cv).

But, by the ordinary formula for the flow of rivers, mi = ζv²;

∴ i = ζv² / m = (ζv^5/2 / a) √ (c / kl).

But i is the tangent of the angle which the curve at C makes with the axis of X, and is therefore = dy/dx. Also, as the slope is small, l = AC = AD = x nearly.

∴ dy/dx = (ζv^5/2 / a) √ (c / kx);

and, remembering that v is constant,

y = (2ζv^5/2 / a) √ (cx / k);

or

y² = constant × x;

so that the curve is a common parabola, of which the axis is horizontal and the vertex at the source. This may be considered an ideal longitudinal section, to which actual rivers approximate more or less, with exceptions due to the varying hardness of their beds, and the irregular manner in which their volume increases.

§ 129. Surface Level of River.—The surface level of a river is a plane changing constantly in position from changes in the volume of water discharged, and more slowly from changes in the river bed, and the circumstances affecting the drainage into the river.

For the purposes of the engineer, it is important to determine (1) the extreme low water level, (2) the extreme high water or flood level, and (3) the highest navigable level.

1. Low Water Level cannot be absolutely known, because a river reaches its lowest level only at rare intervals, and because alterations in the cultivation of the land, the drainage, the removal of forests, the removal or erection of obstructions in the river bed, &c., gradually alter the conditions of discharge. The lowest level of which records can be found is taken as the conventional or approximate low water level, and allowance is made for possible changes.

2. High Water or Flood Level.—The engineer assumes as the highest flood level the highest level of which records can be obtained. In forming a judgment of the data available, it must be remembered that the highest level at one point of a river is not always simultaneous with the attainment of the highest level at other points, and that the rise of a river in flood is very different in different parts of its course. In temperate regions, the floods of rivers seldom rise more than 20 ft. above low-water level, but in the tropics the rise of floods is greater.

3. Highest Navigable Level.—When the river rises above a certain level, navigation becomes difficult from the increase of the velocity of the current, or from submersion of the tow paths, or from the headway under bridges becoming insufficient. Ordinarily the highest navigable level may be taken to be that at which the river begins to overflow its banks.

§ 130. Relative Value of Different Materials for Submerged Works.—That the power of water to remove and transport different materials depends on their density has an important bearing on the selection of materials for submerged works. In many cases, as in the aprons or floorings beneath bridges, or in front of locks or falls, and in the formation of training walls and breakwaters by pierres perdus, which have to resist a violent current, the materials of which the structures are composed should be of such a size and weight as to be able individually to resist the scouring action of the water. The heaviest materials will therefore be the best; and the different value of materials in this respect will appear much more striking, if it is remembered that all materials lose part of their weight in water. A block whose volume is V cubic feet, and whose density in air is w ℔ per cubic foot, weighs in air wV ℔, but in water only (w−62.4)V ℔.

	Weight of a Cub. Ft. in ℔.
	In Air.	In Water.
Basalt	187.3	124.9
Brick	130.0	67.6
Brickwork	112.0	49.6
Granite and limestone	170.0	107.6
Sandstone	144.0	81.6
Masonry	116-144	53.6-81.6

§ 131. Inundation Deposits from a River.—When a river carrying silt periodically overflows its banks, it deposits silt over the area flooded, and gradually raises the surface of the country. The silt is deposited in greatest abundance where the water first leaves the river. It hence results that the section of the country assumes a peculiar form, the river flowing in a trough along the crest of a ridge, from which the land slopes downwards on both sides. The silt deposited from the water forms two wedges, having their thick ends towards the river (fig. 133).

Fig. 133.

This is strikingly the case with the Mississippi, and that river is now kept from flooding immense areas by artificial embankments or levees. In India, the term deltaic segment is sometimes applied to that portion of a river running through deposits formed by inundation, and having this characteristic section. The irrigation of the country in this case is very easy; a comparatively slight raising of the river surface by a weir or annicut gives a command of level which permits the water to be conveyed to any part of the district.

§ 132. Deltas.—The name delta was originally given to the Δ-shaped portion of Lower Egypt, included between seven branches of the Nile. It is now given to the whole of the alluvial tracts round river mouths formed by deposition of sediment from the river, where its velocity is checked on its entrance to the sea. The characteristic feature of these alluvial deltas is that the river traverses them, not in a single channel, but in two or many bifurcating branches. Each branch has a tract of the delta under its influence, and gradually raises the surface of that tract, and extends it seaward. As the delta extends itself seaward, the conditions of discharge through the different branches change. The water finds the passage through one of the branches less obstructed than through the others; the velocity and scouring action in that branch are increased; in the others they diminish. The one channel gradually absorbs the whole of the water supply, while the other branches silt up. But as the mouth of the new main channel extends seaward the resistance increases both from the greater length of the channel and the formation of shoals at its mouth, and the river tends to form new bifurcations AC or AD (fig. 134), and one of these may in time become the main channel of the river.

§ 133. Field Operations preliminary to a Study of River Improvement.—There are required (1) a plan of the river, on which the positions of lines of levelling and cross sections are marked; (2) a longitudinal section and numerous cross sections of the river; (3) a series of gaugings of the discharge at different points and in different conditions of the river.

Longitudinal Section.—This requires to be carried out with great accuracy. A line of stakes is planted, following the sinuosities of the river, and chained and levelled. The cross sections are referred to the line of stakes, both as to position and direction. The determination of the surface slope is very difficult, partly from its extreme smallness, partly from oscillation of the water. Cunningham recommends that the slope be taken in a length of 2000 ft. by four simultaneous observations, two on each side of the river.

Fig. 134.

§ 134. Cross Sections—A stake is planted flush with the water, and its level relatively to some point on the line of levels is determined. Then the depth of the water is determined at a series of points (if possible at uniform distances) in a line starting from the stake and perpendicular to the thread of the stream. To obtain these, a wire may be stretched across with equal distances marked on it by hanging tags. The depth at each of these tags may be obtained by a light wooden staff, with a disk-shaped shoe 4 to 6 in. in diameter. If the depth is great, soundings may be taken by a chain and weight. To ensure the wire being perpendicular to the thread of the stream, it is desirable to stretch two other wires similarly graduated, one above and the other below, at a distance of 20 to 40 yds. A number of floats being then thrown in, it is observed whether they pass the same graduation on each wire.

Fig. 135.

For large and rapid rivers the cross section is obtained by sounding in the following way. Let AC (fig. 135) be the line on which soundings are required. A base line AB is measured out at right angles to AC, and ranging staves are set up at AB and at D in line with AC. A boat is allowed to drop down stream, and, at the moment it comes in line with AD, the lead is dropped, and an observer in the boat takes, with a box sextant, the angle AEB subtended by AB. The sounding line may have a weight of 14 ℔ of lead, and, if the boat drops down stream slowly, it may hang near the bottom, so that the observation is made instantly. In extensive surveys of the Mississippi observers with theodolites were stationed at A and B. The theodolite at A was directed towards C, that at B was kept on the boat. When the boat came on the line AC, the observer at A signalled, the sounding line was dropped, and the observer at B read off the angle ABE. By repeating observations a number of soundings are obtained, which can be plotted in their proper position, and the form of the river bed drawn by connecting the extremities of the lines. From the section can be measured the sectional area of the stream Ω and its wetted perimeter χ; and from these the hydraulic mean depth m can be calculated.

§ 135. Measurement of the Discharge of Rivers.—The area of cross section multiplied by the mean velocity gives the discharge of the stream. The height of the river with reference to some fixed mark should be noted whenever the velocity is observed, as the velocity and area of cross section are different in different states of the river. To determine the mean velocity various methods may be adopted; and, since no method is free from liability to error, either from the difficulty of the observations or from uncertainty as to the ratio of the mean velocity to the velocity observed, it is desirable that more than one method should be used.

Instruments for Measuring the Velocity of Water

§ 136. Surface Floats are convenient for determining the surface velocities of a stream, though their use is difficult near the banks. The floats may be small balls of wood, of wax or of hollow metal, so loaded as to float nearly flush with the water surface. To render them visible they may have a vertical painted stem. In experiments on the Seine, cork balls 13/4 in. diameter were used, loaded to float flush with the water, and provided with a stem. In A. J. C. Cunningham’s observations at Roorkee, the floats were thin circular disks of English deal, 3 in. diameter and 1/4 in. thick. For observations near the banks, floats 1 in. diameter and 1/8 in. thick were used. To render them visible a tuft of cotton wool was used loosely fixed in a hole at the centre.

Fig. 136.

The velocity is obtained by allowing the float to be carried down, and noting the time of passage over a measured length of the stream. If v is the velocity of any float, t the time of passing over a length l, then v = l/t. To mark out distinctly the length of stream over which the floats pass, two ropes may be stretched across the stream at a distance apart, which varies usually from 50 to 250 ft., according to the size and rapidity of the river. In the Roorkee experiments a length of run of 50 ft. was found best for the central two-fifths of the width, and 25 ft. for the remainder, except very close to the banks, where the run was made 121/2 ft. only. The longer the run the less is the proportionate error of the time observations, but on the other hand the greater the deviation of the floats from a straight course parallel to the axis of the stream. To mark the precise position at which the floats cross the ropes, Cunningham used short white rope pendants, hanging so as nearly to touch the surface of the water. In this case the streams were 80 to 180 ft. in width. In wider streams the use of ropes to mark the length of run is impossible, and recourse must be had to box sextants or theodolites to mark the path of the floats.

Let AB (fig. 136) be a measured base line strictly parallel to the thread of the stream, and AA₁, BB₁ lines at right angles to AB marked out by ranging rods at A₁ and B₁. Suppose observers stationed at A and B with sextants or theodolites, and let CD be the path of any float down stream. As the float approaches AA₁, the observer at B keeps it on the cross wire of his instrument. The observer at A observes the instant of the float reaching the line AA₁, and signals to B who then reads off the angle ABC. Similarly, as the float approaches BB₁, the observer at A keeps it in sight, and when signalled to by B reads the angle BAD. The data so obtained are sufficient for plotting the path of the float and determining the distances AC, BD.

The time taken by the float in passing over the measured distance may be observed by a chronograph, started as the float passes the upper rope or line, and stopped when it passes the lower. In Cunningham’s observations two chronometers were sometimes used, the time of passing one end of the run being noted on one, and that of passing the other end of the run being noted on the other. The chronometers were compared immediately before the observations. In other cases a single chronometer was used placed midway of the run. The moment of the floats passing the ends of the run was signalled to a time-keeper at the chronometer by shouting. It was found quite possible to count the chronometer beats to the nearest half second, and in some cases to the nearest quarter second.

Fig. 137.

§ 137. Sub-surface Floats.—The velocity at different depths below the surface of a stream may be obtained by sub-surface floats, used precisely in the same way as surface floats. The most usual arrangement is to have a large float, of slightly greater density than water, connected with a small and very light surface float. The motion of the combined arrangement is not sensibly different from that of the large float, and the small surface float enables an observer to note the path and velocity of the sub-surface float. The instrument is, however, not free from objection. If the large submerged float is made of very nearly the same density as water, then it is liable to be thrown upwards by very slight eddies in the water, and it does not maintain its position at the depth at which it is intended to float. On the other hand, if the large float is made sensibly heavier than water, the indicating or surface float must be made rather large, and then it to some extent influences the motion of the submerged float. Fig. 137 shows one form of sub-surface float. It consists of a couple of tin plates bent at a right angle and soldered together at the angle. This is connected with a wooden ball at the surface by a very thin wire or cord. As the tin alone makes a heavy submerged float, it is better to attach to the tin float some pieces of wood to diminish its weight in water. Fig. 138 shows the form of submerged float used by Cunningham. It consists of a hollow metal ball connected to a slice of cork, which serves as the surface float.

Fig. 138.	Fig. 139.

§ 138. Twin Floats.—Suppose two equal and similar floats (fig. 139) connected by a wire. Let one float be a little lighter and the other a little heavier than water. Then the velocity of the combined floats will be the mean of the surface velocity and the velocity at the depth at which the heavier float swims, which is determined by the length of the connecting wire. Thus if v_s is the surface velocity and v_d the velocity at the depth to which the lower float is sunk, the velocity of the combined floats will be

v = 1/2 (v_s + v_d).

Consequently, if v is observed, and v_s determined by an experiment with a single float,

v_d = 2v − v_s

According to Cunningham, the twin float gives better results than the sub-surface float.

Fig. 140.

§ 139. Velocity Rods.—Another form of float is shown in fig. 140. This consists of a cylindrical rod loaded at the lower end so as to float nearly vertical in water. A wooden rod, with a metal cap at the bottom in which shot can be placed, answers better than anything else, and sometimes the wooden rod is made in lengths, which can be screwed together so as to suit streams of different depths. A tuft of cotton wool at the top serves to make the float more easily visible. Such a rod, so adjusted in length that it sinks nearly to the bed of the stream, gives directly the mean velocity of the whole vertical section in which it floats.

§ 140. Revy’s Current Meter.—No instrument has been so much used in directly determining the velocity of a stream at a given point as the screw current meter. Of this there are a dozen varieties at least. As an example of the instrument in its simplest form, Revy’s meter may be selected. This is an ordinary screw meter of a larger size than usual, more carefully made, and with its details carefully studied (figs. 141, 142). It was designed after experience in gauging the great South American rivers. The screw, which is actuated by the water, is 6 in. in diameter, and is of the type of the Griffiths screw used in ships. The hollow spherical boss serves to make the weight of the screw sensibly equal to its displacement, so that friction is much reduced. On the axis aa of the screw is a worm which drives the counter. This consists of two worm wheels g and h fixed on a common axis. The worm wheels are carried on a frame attached to the pin l. By means of a string attached to l they can be pulled into gear with the worm, or dropped out of gear and stopped at any instant. A nut m can be screwed up, if necessary, to keep the counter permanently in gear. The worm is two-threaded, and the worm wheel g has 200 teeth. Consequently it makes one rotation for 100 rotations of the screw, and the number of rotations up to 100 is marked by the passage of the graduations on its edge in front of a fixed index. The second worm wheel has 196 teeth, and its edge is divided into 49 divisions. Hence it falls behind the first wheel one division for a complete rotation of the latter. The number of hundreds of rotations of the screw are therefore shown by the number of divisions on h passed over by an index fixed to g. One difficulty in the use of the ordinary screw meter is that particles of grit, getting into the working parts, very sensibly alter the friction, and therefore the speed of the meter. Revy obviates this by enclosing the counter in a brass box with a glass face. This box is filled with pure water, which ensures a constant coefficient of friction for the rubbing parts, and prevents any mud or grit finding its way in. In order that the meter may place itself with the axis parallel to the current, it is pivoted on a vertical axis and directed by a large vane shown in fig. 142. To give the vane more directing power the vertical axis is nearer the screw than in ordinary meters, and the vane is larger. A second horizontal vane is attached by the screws x, x, the object of which is to allow the meter to rest on the ground without the motion of the screw being interfered with. The string or wire for starting and stopping the meter is carried through the centre of the vertical axis, so that the strain on it may not tend to pull the meter oblique to the current. The pitch of the screw is about 9 in. The screws at x serve for filling the meter with water. The whole apparatus is fixed to a rod (fig. 142), of a length proportionate to the depth, or for very great depths it is fixed to a weighted bar lowered by ropes, a plan invented by Revy. The instrument is generally used thus. The reading of the counter is noted, and it is put out of gear. The meter is then lowered into the water to the required position from a platform between two boats, or better from a temporary bridge. Then the counter is put into gear for one, two or five minutes. Lastly, the instrument is raised and the counter again read. The velocity is deduced from the number of rotations in unit time by the formulae given below. For surface velocities the counter may be kept permanently in gear, the screw being started and stopped by hand.

Fig. 141.

Fig. 142.

§ 141. The Harlacher Current Meter.—In this the ordinary counting apparatus is abandoned. A worm drives a worm wheel, which makes an electrical contact once for each 100 rotations of the worm. This contact gives a signal above water. With this arrangement, a series of velocity observations can be made, without removing the instrument from the water, and a number of practical difficulties attending the accurate starting and stopping of the ordinary counter are entirely got rid of. Fig. 143 shows the meter. The worm wheel z makes one rotation for 100 of the screw. A pin moving the lever x makes the electrical contact. The wires b, c are led through a gas pipe B; this also serves to adjust the meter to any required position on the wooden rod dd. The rudder or vane is shown at WH. The galvanic current acts on the electromagnet m, which is fixed in a small metal box containing also the battery. The magnet exposes and withdraws a coloured disk at an opening in the cover of the box.

§ 142. Amsler Laffon Current Meter.—A very convenient and accurate current meter is constructed by Amsler Laffon of Schaffhausen. This can be used on a rod, and put into and out of gear by a ratchet. The peculiarity in this case is that there is a double ratchet, so that one pull on the string puts the counter into gear and a second puts it out of gear. The string may be slack during the action of the meter, and there is less uncertainty than when the counter has to be held in gear. For deep streams the meter A is suspended by a wire with a heavy lenticular weight below (fig. 144). The wire is payed out from a small winch D, with an index showing the depth of the meter, and passes over a pulley B. The meter is in gimbals and is directed by a conical rudder which keeps it facing the stream with its axis horizontal. There is an electric circuit from a battery C through the meter, and a contact is made closing the circuit every 100 revolutions. The moment the circuit closes a bell rings. By a subsidiary arrangement, when the foot of the instrument, 0.3 metres below the axis of the meter, touches the ground the circuit is also closed and the bell rings. It is easy to distinguish the continuous ring when the ground is reached from the short ring when the counter signals. A convenient winch for the wire is so graduated that if set when the axis of the meter is at the water surface it indicates at any moment the depth of the meter below the surface. Fig. 144 shows the meter as used on a boat. It is a very convenient instrument for obtaining the velocity at different depths and can also be used as a sounding instrument.

Fig. 143.

§ 143. Determination of the Coefficients of the Current Meter.—Suppose a series of observations has been made by towing the meter in still water at different speeds, and that it is required to ascertain from these the constants of the meter. If v is the velocity of the water and n the observed number of rotations per second, let

v = α + βn

(1)

where α and β are constants. Now let the meter be towed over a measured distance L, and let N be the revolutions of the meter and t the time of transit. Then the speed of the meter relatively to the water is L/t = v feet per second, and the number of revolutions per second is N/t = n. Suppose m observations have been made in this way, furnishing corresponding values of v and n, the speed in each trial being as uniform as possible,

Σn =	n1 + n2 + . . .
Σv =	v1 + v2 + . . .
Σnv =	n1v1 + n2v2 + . . .
Σn2 =	n12 + n22 + . . .
[Σn]2 =	[n1 + n2 + . . .]2

Then for the determination of the constants α and β in (1), by the

method of least squares—

α = Σn²Σv − ΣnΣnv/mΣn² − [Σn]²,

β = mΣnv − ΣvΣn/mΣn² − [Σn]². 

Fig. 144.

In a few cases the constants for screw current meters have been determined by towing them in R. E. Froude’s experimental tank in which the resistance of ship models is ascertained. In that case the data are found with exceptional accuracy.

§ 144. Darcy Gauge or modified Pitot Tube.—A very old instrument for measuring velocities, invented by Henri Pitot in 1730 (Histoire de l’Académie des Sciences, 1732, p. 376), consisted simply of a vertical glass tube with a right-angled bend, placed so that its mouth was normal to the direction of flow (fig. 145).

Fig. 145.

The impact of the stream on the mouth of the tube balances a column in the tube, the height of which is approximately h = v²/2g, where v is the velocity at the depth x. Placed with its mouth parallel to the stream the water inside the tube is nearly at the same level as the surface of the stream, and turned with the mouth down stream, the fluid sinks a depth h′ = v²/2g nearly, though the tube in that case interferes with the free flow of the liquid and somewhat modifies the result. Pitot expanded the mouth of the tube so as to form a funnel or bell mouth. In that case he found by experiment

h = 1.5v² / 2g.

But there is more disturbance of the stream. Darcy preferred to make the mouth of the tube very small to avoid interference with the stream and to check oscillations of the water column. Let the difference of level of a pair of tubes A and B (fig. 145) be taken to be h = kv²/2g, then k may be taken to be a corrective coefficient whose value in well-shaped instruments is very nearly unity. By placing his instrument in front of a boat towed through water Darcy found k = 1.034; by placing the instrument in a stream the velocity of which had been ascertained by floats, he found k = 1.006; by readings taken in different parts of the section of a canal in which a known volume of water was flowing, he found k = 0.993. He believed the first value to be too high in consequence of the disturbance caused by the boat. The mean of the other two values is almost exactly unity (Recherches hydrauliques, Darcy and Bazin, 1865, p. 63). W. B. Gregory used somewhat differently formed Pitot tubes for which the k = 1 (Am. Soc. Mech. Eng., 1903, 25). T. E. Stanton used a Pitot tube in determining the velocity of an air current, and for his instrument he found k = 1.030 to k = 1.032 (“On the Resistance of Plane Surfaces in a Current of Air,” Proc. Inst. Civ. Eng., 1904, 156).

One objection to the Pitot tube in its original form was the great difficulty and inconvenience of reading the height h in the immediate neighbourhood of the stream surface. This is obviated in the Darcy gauge, which can be removed from the stream to be read.

Fig. 146 shows a Darcy gauge. It consists of two Pitot tubes having their mouths at right angles. In the instrument shown, the two tubes, formed of copper in the lower part, are united into one for strength, and the mouths of the tubes open vertically and horizontally. The upper part of the tubes is of glass, and they are provided with a brass scale and two verniers b, b. The whole instrument is supported on a vertical rod or small pile AA, the fixing at B permitting the instrument to be adjusted to any height on the rod, and at the same time allowing free rotation, so that it can be held parallel to the current. At c is a two-way cock, which can be opened or closed by cords. If this is shut, the instrument can be lifted out of the stream for reading. The glass tubes are connected at top by a brass fixing, with a stop cock a, and a flexible tube and mouthpiece m. The use of this is as follows. If the velocity is required at a point near the surface of the stream, one at least of the water columns would be below the level at which it could be read. It would be in the copper part of the instrument. Suppose then a little air is sucked out by the tube m, and the cock a closed, the two columns will be forced up an amount corresponding to the difference between atmospheric pressure and that in the tubes. But the difference of level will remain unaltered.

When the velocities to be measured are not very small, this instrument is an admirable one. It requires observation only of a single linear quantity, and does not require any time observation. The law connecting the velocity and the observed height is a rational one, and it is not absolutely necessary to make any experiments on the coefficient of the instrument. If we take v = k√(2gh), then it appears from Darcy’s experiments that for a well-formed instrument k does not sensibly differ from unity. It gives the velocity at a definite point in the stream. The chief difficulty arises from the fact that at any given point in a stream the velocity is not absolutely constant, but varies a little from moment to moment. Darcy in some of his experiments took several readings, and deduced the velocity from the mean of the highest and lowest.

§ 145. Perrodil Hydrodynamometer.—This consists of a frame abcd (fig. 147) placed vertically in the stream, and of a height not less than the stream’s depth. The two vertical members of this frame are connected by cross bars, and united above water by a circular bar, situated in the vertical plane and carrying a horizontal graduated circle ef. This whole system is movable round its axis, being suspended on a pivot at g connected with the fixed support mn. Other horizontal arms serve as guides. The central vertical rod gr forms a torsion rod, being fixed at r to the frame abcd, and, passing freely upwards through the guides, it carries a horizontal needle moving over the graduated circle ef. The support g, which carries the apparatus, also receives in a tubular guide the end of the torsion rod gr and a set screw for fixing the upper end of the torsion rod when necessary. The impulse of the stream of water is received on a circular disk x, in the plane of the torsion rod and the frame abcd. To raise and lower the apparatus easily, it is not fixed directly to the rod mn, but to a tube kl sliding on mn.

Fig. 146.

Fig. 147.

Suppose the apparatus arranged so that the disk x is at that level in the stream where the velocity is to be determined. The plane abcd is placed parallel to the direction of motion of the water. Then the disk x (acting as a rudder) will place itself parallel to the stream on the down stream side of the frame. The torsion rod will be unstrained, and the needle will be at zero on the graduated circle. If, then, the instrument is turned by pressing the needle, till the plane abcd of the disk and the zero of the graduated circle is at right angles to the stream, the torsion rod will be twisted through an angle which measures the normal impulse of the stream on the disk x. That angle will be given by the distance of the needle from zero. Observation shows that the velocity of the water at a given point is not constant. It varies between limits more or less wide. When the apparatus is nearly in its right position, the set screw at g is made to clamp the torsion spring. Then the needle is fixed, and the apparatus carrying the graduated circle oscillates. It is not, then, difficult to note the mean angle marked by the needle.

Let r be the radius of the torsion rod, l its length from the needle over ef to r, and α the observed torsion angle. Then the moment of the couple due to the molecular forces in the torsion rod is

M = E_tIα / l;

where E_t is the modulus of elasticity for torsion, and I the polar moment of inertia of the section of the rod. If the rod is of circular section, I = 1/2πr⁴. Let R be the radius of the disk, and b its leverage, or the distance of its centre from the axis of the torsion rod. The moment of the pressure of the water on the disk is

Fb = kb(G/2g) πR²v²,

where G is the heaviness of water and k an experimental coefficient. Then

E_tIα / l = kb(G/ 2g)πR²v².

For any given instrument,

v = c √ α,

where c is a constant coefficient for the instrument.

The instrument as constructed had three disks which could be used at will. Their radii and leverages were in feet

	R =	b =
1st disk	0.052	0.16
2nd ’’	0.105	0.32
3rd  ’’	0.210	0.66

For a thin circular plate, the coefficient k = 1.12. In the actual instrument the torsion rod was a brass wire 0.06 in. diameter and 61/2 ft. long. Supposing α measured in degrees, we get by calculation

v = 0.335 √ α; 0.115 √ α; 0.042 √ α.

Very careful experiments were made with the instrument. It was fixed to a wooden turning bridge, revolving over a circular channel of 2 ft. width, and about 76 ft. circumferential length. An allowance was made for the slight current produced in the channel. These experiments gave for the coefficient c, in the formula v = c√α,

1st disk,	c = 0.3126 for velocities of 3 to 16 ft.
2nd disk,	c = 0.1177 for velocities of 11/4 to 31/4 ft.
3rd disk,	c = 0.0349 for velocities of less than 11/4 ft.

The instrument is preferable to the current meter in giving the velocity in terms of a single observed quantity, the angle of torsion, while the current meter involves the observation of two quantities, the number of rotations and the time. The current meter, except in some improved forms, must be withdrawn from the water to read the result of each experiment, and the law connecting the velocity and number of rotations of a current meter is less well-determined than that connecting the pressure on a disk and the torsion of the wire of a hydrodynamometer.

The Pitot tube, like the hydrodynamometer, does not require a time observation. But, where the velocity is a varying one, and consequently the columns of water in the Pitot tube are oscillating, there is room for doubt as to whether, at any given moment of closing the cock, the difference of level exactly measures the impulse of the stream at the moment. The Pitot tube also fails to give measurable indications of very low velocities.

Processes for Gauging Streams

§ 146. Gauging by Observation of the Maximum Surface Velocity.—The method of gauging which involves the least trouble is to determine the surface velocity at the thread of the stream, and to deduce from it the mean velocity of the whole cross section. The maximum surface velocity may be determined by floats or by a current meter. Unfortunately the ratio of the maximum surface to the mean velocity is extremely variable. Thus putting v₀ for the surface velocity at the thread of the stream, and v_m for the mean velocity of the whole cross section, v_m/v₀ has been found to have the following values:—

	vm/v0
De Prony, experiments on small wooden channels	0.8164
Experiments on the Seine	0.62
Destrem and De Prony, experiments on the Neva	0.78
Boileau, experiments on canals	0.82
Baumgartner, experiments on the Garonne	0.80
Brünings (mean)	0.85
Cunningham, Solani aqueduct	0.823

Various formulae, either empirical or based on some theory of the vertical and horizontal velocity curves, have been proposed for determining the ratio v_m/v₀. Bazin found from his experiments the empirical expression

v_m = v₀ − 25.4 √ (mi);

where m is the hydraulic mean depth and i the slope of the stream.

In the case of irrigation canals and rivers, it is often important to determine the discharge either daily or at other intervals of time, while the depth and consequently the mean velocity is varying. Cunningham (Roorkee Prof. Papers, iv. 47), has shown that, for a given part of such a stream, where the bed is regular and of permanent section, a simple formula may be found for the variation of the central surface velocity with the depth. When once the constants of this formula have been determined by measuring the central surface velocity and depth, in different conditions of the stream, the surface velocity can be obtained by simply observing the depth of the stream, and from this the mean velocity and discharge can be calculated. Let z be the depth of the stream, and v₀ the surface velocity, both measured at the thread of the stream. Then v₀² = cz; where c is a constant which for the Solani aqueduct had the values 1.9 to 2, the depths being 6 to 10 ft., and the velocities 31/2 to 41/2 ft. Without any assumption of a formula, however, the surface velocities, or still better the mean velocities, for different conditions of the stream may be plotted on a diagram in which the abscissae are depths and the ordinates velocities. The continuous curve through points so found would then always give the velocity for any observed depth of the stream, without the need of making any new float or current meter observations.

§ 147. Mean Velocity determined by observing a Series of Surface Velocities.—The ratio of the mean velocity to the surface velocity in one longitudinal section is better ascertained than the ratio of the central surface velocity to the mean velocity of the whole cross section. Suppose the river divided into a number of compartments by equidistant longitudinal planes, and the surface velocity observed in each compartment. From this the mean velocity in each compartment and the discharge can be calculated. The sum of the partial discharges will be the total discharge of the stream. When wires or ropes can be stretched across the stream, the compartments can be marked out by tags attached to them. Suppose two such ropes stretched across the stream, and floats dropped in above the upper rope. By observing within which compartment the path of the float lies, and noting the time of transit between the ropes, the surface velocity in each compartment can be ascertained. The mean velocity in each compartment is 0.85 to 0.91 of the surface velocity in that compartment. Putting k for this ratio, and v₁, v₂ . . . for the observed velocities, in compartments of area Ω₁, Ω₂ . . . then the total discharge is

Q = k (Ω₁v₁ + Ω₂v₂ + . . . ).

If several floats are allowed to pass over each compartment, the mean of all those corresponding to one compartment is to be taken as the surface velocity of that compartment.

Fig. 148.

This method is very applicable in the case of large streams or rivers too wide to stretch a rope across. The paths of the floats are then ascertained in this way. Let fig. 148 represent a portion of the river, which should be straight and free from obstructions. Suppose a base line AB measured parallel to the thread of the stream, and let the mean cross section of the stream be ascertained either by sounding the terminal cross sections AE, BF, or by sounding a series of equidistant cross sections. The cross sections are taken at right angles to the base line. Observers are placed at A and B with theodolites or box sextants. The floats are dropped in from a boat above AE, and picked up by another boat below BF. An observer with a chronograph or watch notes the time in which each float passes from AE to BF. The method of proceeding is this. The observer A sets his theodolite in the direction AE, and gives a signal to drop a float. B keeps his instrument on the float as it comes down. At the moment the float arrives at C in the line AE, the observer at A calls out. B clamps his instrument and reads off the angle ABC, and the time observer begins to note the time of transit. B now points his instrument in the direction BF, and A keeps the float on the cross wire of his instrument. At the moment the float arrives at D in the line BF, the observer B calls out, A clamps his instrument and reads off the angle BAD, and the time observer notes the time of transit from C to D. Thus all the data are determined for plotting the path CD of the float and determining its velocity. By dropping in a series of floats, a number of surface velocities can be determined. When all these have been plotted, the river can be divided into convenient compartments. The observations belonging to each compartment are then averaged, and the mean velocity and discharge calculated. It is obvious that, as the surface velocity is greatly altered by wind, experiments of this kind should be made in very calm weather.

The ratio of the surface velocity to the mean velocity in the same vertical can be ascertained from the formulae for the vertical velocity curve already given (§ 101). Exner, in Erbkam’s Zeitschrift for 1875, gave the following convenient formula. Let v be the mean and V the surface velocity in any given vertical longitudinal section, the depth of which is h

v / V = (1 + 0.1478 √ h) / (1 + 0.2216 √ h).

If vertical velocity rods are used instead of common floats, the mean velocity is directly determined for the vertical section in which the rod floats. No formula of reduction is then necessary. The observed velocity has simply to be multiplied by the area of the compartment to which it belongs.

§ 148. Mean Velocity of the Stream from a Series of Mid Depth Velocities.—In the gaugings of the Mississippi it was found that the mid depth velocity differed by only a very small quantity from the mean velocity in the vertical section, and it was uninfluenced by wind. If therefore a series of mid depth velocities are determined by double floats or by a current meter, they may be taken to be the mean velocities of the compartments in which they occur, and no formula of reduction is necessary. If floats are used, the method is precisely the same as that described in the last paragraph for surface floats. The paths of the double floats are observed and plotted, and the mean taken of those corresponding to each of the compartments into which the river is divided. The discharge is the sum of the products of the observed mean mid depth velocities and the areas of the compartments.

§ 149. P. P. Boileau’s Process for Gauging Streams.—Let U be the mean velocity at a given section of a stream, V the maximum velocity, or that of the principal filament, which is generally a little below the surface, W and w the greatest and least velocities at the surface. The distance of the principal filament from the surface is generally less than one-fourth of the depth of the stream; W is a little less than V; and U lies between W and w. As the surface velocities change continuously from the centre towards the sides there are at the surface two filaments having a velocity equal to U. The determination of the position of these filaments, which Boileau terms the gauging filaments, cannot be effected entirely by theory. But, for sections of a stream in which there are no abrupt changes of depth, their position can be very approximately assigned. Let Δ and l be the horizontal distances of the surface filament, having the velocity W, from the gauging filament, which has the velocity U, and from the bank on one side. Then

Δ / l = c⁴ √ {(W + 2w) / 7 (W − w)},

c being a numerical constant. From gaugings by Humphreys and Abbot, Bazin and Baumgarten, the values c = 0.919, 0.922 and 0.925 are obtained. Boileau adopts as a mean value 0.922. Hence, if W and w are determined by float gauging or otherwise, Δ can be found, and then a single velocity observation at Δ ft. from the filament of maximum velocity gives, without need of any reduction, the mean velocity of the stream. More conveniently W, w, and U can be measured from a horizontal surface velocity curve, obtained from a series of float observations.

§ 150. Direct Determination of the Mean Velocity by a Current Meter or Darcy Gauge.—The only method of determining the mean velocity at a cross section of a stream which involves no assumption of the ratio of the mean velocity to other quantities is this—a plank bridge is fixed across the stream near its surface. From this, velocities are observed at a sufficient number of points in the cross section of the stream, evenly distributed over its area. The mean of these is the true mean velocity of the stream. In Darcy and Bazin’s experiments on small streams, the velocity was thus observed at 36 points in the cross section.

When the stream is too large to fix a bridge across it, the observations may be taken from a boat, or from a couple of boats with a gangway between them, anchored successively at a series of points across the width of the stream. The position of the boat for each series of observations is fixed by angular observations to a base line on shore.

Fig. 149.

Fig. 150.

§ 151. A. R. Harlacher’s Graphic Method of determining the Discharge from a Series of Current Meter Observations.—Let ABC (fig. 149) be the cross section of a river at which a complete series of current meter observations have been taken. Let I., II., III., . . . be the verticals at different points of which the velocities were measured. Suppose the depths at I., II., III., . . . (fig. 149), set off as vertical ordinates in fig. 150, and on these vertical ordinates suppose the velocities set off horizontally at their proper depths. Thus, if v is the measured velocity at the depth h from the surface in fig. 149, on vertical marked III., then at III. in fig. 150 take cd = h and ac = v. Then d is a point in the vertical velocity curve for the vertical III., and, all the velocities for that ordinate being similarly set off, the curve can be drawn. Suppose all the vertical velocity curves I.... V. (fig. 150), thus drawn. On each of these figures draw verticals corresponding to velocities of x, 2x, 3x . . . ft. per second. Then for instance cd at III. (fig. 150) is the depth at which a velocity of 2x ft. per second existed on the vertical III. in fig. 149 and if cd is set off at III. in fig. 149 it gives a point in a curve passing through points of the section where the velocity was 2x ft. per second. Set off on each of the verticals in fig. 149 all the depths thus found in the corresponding diagram in fig. 150. Curves drawn through the corresponding points on the verticals are curves of equal velocity.

The discharge of the stream per second may be regarded as a solid having the cross section of the river (fig. 149) as a base, and cross sections normal to the plane of fig. 149 given by the diagrams in fig. 150. The curves of equal velocity may therefore be considered as contour lines of the solid whose volume is the discharge of the stream per second. Let Ω₀ be the area of the cross section of the river, Ω₁, Ω₂ . . . the areas contained by the successive curves of equal velocity, or, if these cut the surface of the stream, by the curves and that surface. Let x be the difference of velocity for which the successive curves are drawn, assumed above for simplicity at 1 ft. per second. Then the volume of the successive layers of the solid body whose volume represents the discharge, limited by successive planes passing through the contour curves, will be

1/2 x (Ω₀ + Ω₁), 1/2 x (Ω₁ + Ω₂), and so on.

Consequently the discharge is

Q = x {1/2 (Ω₀ + Ω_n) + Ω₁ = Ω₂ + . . . + Ω_n−1}.

The areas Ω₀, Ω₁ . . . are easily ascertained by means of the polar planimeter. A slight difficulty arises in the part of the solid lying above the last contour curve. This will have generally a height which is not exactly x, and a form more rounded than the other layers and less like a conical frustum. The volume of this may be estimated separately, and taken to be the area of its base (the area Ω_n) multiplied by 1/3 to 1/2 its height.

Fig. 151.

Fig. 151 shows the results of one of Harlacher’s gaugings worked out in this way. The upper figure shows the section of the river and the positions of the verticals at which the soundings and gaugings were taken. The lower gives the curves of equal velocity, worked out from the current meter observations, by the aid of vertical velocity curves. The vertical scale in this figure is ten times as great as in the other. The discharge calculated from the contour curves is 14.1087 cubic metres per second. In the lower figure some other interesting curves are drawn. Thus, the uppermost dotted curve is the curve through points at which the maximum velocity was found; it shows that the maximum velocity was always a little below the surface, and at a greater depth at the centre than at the sides. The next curve shows the depth at which the mean velocity for each vertical was found. The next is the curve of equal velocity corresponding to the mean velocity of the stream; that is, it passes through points in the cross section where the velocity was identical with the mean velocity of the stream.

Hydraulic Machines

§ 152. Hydraulic machines may be broadly divided into two classes: (1) Motors, in which water descending from a higher to a lower level, or from a higher to a lower pressure, gives up energy which is available for mechanical operations; (2) Pumps, in which the energy of a steam engine or other motor is expended in raising water from a lower to a higher level. A few machines such as the ram and jet pump combine the functions of motor and pump. It may be noted that constructively pumps are essentially reversed motors. The reciprocating pump is a reversed pressure engine, and the centrifugal pump a reversed turbine. Hydraulic machine tools are in principle motors combined with tools, and they now form an important special class.

Water under pressure conveyed in pipes is a convenient and economical means of transmitting energy and distributing it to many scattered working points. Hence large and important hydraulic systems are adopted in which at a central station water is pumped at high pressure into distributing mains, which convey it to various points where it actuates hydraulic motors operating cranes, lifts, dock gates, and in some cases riveting and shearing machines. In this case the head driving the hydraulic machinery is artificially created, and it is the convenience of distributing power in an easily applied form to distant points which makes the system advantageous. As there is some unavoidable loss in creating an artificial head this system is most suitable for driving machines which work intermittently (see Power Transmission). The development of electrical methods of transmitting and distributing energy has led to the utilization of many natural waterfalls so situated as to be useless without such a means of transferring the power to points where it can be conveniently applied. In some cases, as at Niagara, the hydraulic power can only be economically developed in very large units, and it can be most conveniently subdivided and distributed by transformation into electrical energy. Partly from the development of new industries such as paper-making from wood pulp and electro-metallurgical processes, which require large amounts of cheap power, partly from the facility with which energy can now be transmitted to great distances electrically, there has been a great increase in the utilization of water-power in countries having natural waterfalls. According to the twelfth census of the United States the total amount of water-power reported as used in manufacturing establishments in that country was 1,130,431 h.p. in 1870; 1,263,343 h.p. in 1890; and 1,727,258 h.p. in 1900. The increase was 8.4% in the decade 1870–1880, 3.1% in 1880–1890, and no less than 36.7% in 1890–1900. The increase is the more striking because in this census the large amounts of hydraulic power which are transmitted electrically are not included.

XII. IMPACT AND REACTION OF WATER

§ 153. When a stream of fluid in steady motion impinges on a solid surface, it presses on the surface with a force equal and opposite to that by which the velocity and direction of motion of the fluid are changed. Generally, in problems on the impact of fluids, it is necessary to neglect the effect of friction between the fluid and the surface on which it moves.

During Impact the Velocity of the Fluid relatively to the Surface on which it impinges remains unchanged in Magnitude.—Consider a mass of fluid flowing in contact with a solid surface also in motion, the motion of both fluid and solid being estimated relatively to the earth. Then the motion of the fluid may be resolved into two parts, one a motion equal to that of the solid, and in the same direction, the other a motion relatively to the solid. The motion which the fluid has in common with the solid cannot at all be influenced by the contact. The relative component of the motion of the fluid can only be altered in direction, but not in magnitude. The fluid moving in contact with the surface can only have a relative motion parallel to the surface, while the pressure between the fluid and solid, if friction is neglected, is normal to the surface. The pressure therefore can only deviate the fluid, without altering the magnitude of the relative velocity. The unchanged common component and, combined with it, the deviated relative component give the resultant final velocity, which may differ greatly in magnitude and direction from the initial velocity.

From the principle of momentum, the impulse of any mass of fluid reaching the surface in any given time is equal to the change of momentum estimated in the same direction. The pressure between the fluid and surface, in any direction, is equal to the change of momentum in that direction of so much fluid as reaches the surface in one second. If P_a is the pressure in any direction, m the mass of fluid impinging per second, v_a the change of velocity in the direction of P_a due to impact, then

P_a = mv_a.

Fig. 152.

If v₁ (fig. 152) is the velocity and direction of motion before impact, v₂ that after impact, then v is the total change of motion due to impact. The resultant pressure of the fluid on the surface is in the direction of v, and is equal to v multiplied by the mass impinging per second. That is, putting P for the resultant pressure,

P = mv.

Let P be resolved into two components, N and T, normal and tangential to the direction of motion of the solid on which the fluid impinges. Then N is a lateral force producing a pressure on the supports of the solid, T is an effort which does work on the solid. If u is the velocity of the solid, Tu is the work done per second by the fluid in moving the solid surface.

Let Q be the volume, and GQ the weight of the fluid impinging per second, and let v₁ be the initial velocity of the fluid before striking the surface. Then GQv₁²/2g is the original kinetic energy of Q cub. ft. of fluid, and the efficiency of the stream considered as an arrangement for moving the solid surface is

η = Tu/(GQv₁²/2g).

§ 154. Jet deviated entirely in one Direction.—Geometrical Solution (fig. 153).—Suppose a jet of water impinges on a surface ac with a velocity ab, and let it be wholly deviated in planes parallel to the figure. Also let ae be the velocity and direction of motion of the surface. Join eb; then the water moves with respect to the surface in the direction and with the velocity eb. As this relative velocity is unaltered by contact with the surface, take cd = eb, tangent to the surface at c, then cd is the relative motion of the water with respect to the surface at c. Take df equal and parallel to ae. Then fc (obtained by compounding the relative motion of water to surface and common velocity of water and surface) is the absolute velocity and direction of the water leaving the surface. Take ag equal and parallel to fc. Then, since ab is the initial and ag the final velocity and direction of motion, gb is the total change of motion of the water. The resultant pressure on the plane is in the direction gb. Join eg. In the triangle gae, ae is equal and parallel to df, and ag to fc. Hence eg is equal and parallel to cd. But cd = eb = relative motion of water and surface. Hence the change of motion of the water is represented in magnitude and direction by the third side of an isosceles triangle, of which the other sides are equal to the relative velocity of the water and surface, and parallel to the initial and final directions of relative motion.

Fig. 153.

Special Cases

Fig. 154.

§ 155. (1) A Jet impinges on a plane surface at rest, in a direction normal to the plane (fig. 154).—Let a jet whose section is ω impinge with a velocity v on a plane surface at rest, in a direction normal to the plane. The particles approach the plane, are gradually deviated, and finally flow away parallel to the plane, having then no velocity in the original direction of the jet. The quantity of water impinging per second is ωv. The pressure on the plane, which is equal to the change of momentum per second, is P = (G/g) ωv².

(2) If the plane is moving in the direction of the jet with the velocity ±u, the quantity impinging per second is ω(v ± u). The momentum of this quantity before impact is (G/g)ω(v ± u)v. After impact, the water still possesses the velocity ±u in the direction of the jet; and the momentum, in that direction, of so much water as impinges in one second, after impact, is ±(G/g)ω(v ± u)u. The pressure on the plane, which is the change of momentum per second, is the difference of these quantities or P = (G/g)ω(v ± u)². This differs from the expression obtained in the previous case, in that the relative velocity of the water and plane v ± u is substituted for v. The expression may be written P = 2 × G × ω(v ± u)²/2g, where the last two terms are the volume of a prism of water whose section is the area of the jet and whose length is the head due to the relative velocity. The pressure on the plane is twice the weight of that prism of water. The work done when the plane is moving in the same direction as the jet is Pu = (G/g)ω(v − u)²u foot-pounds per second. There issue from the jet ωv cub. ft. per second, and the energy of this quantity before impact is (G/2g)ωv³. The efficiency of the jet is therefore η = 2(v − u)²u/v³. The value of u which makes this a maximum is found by differentiating and equating the differential coefficient to zero:—

dη/du = 2 (v² − 4vu + 3u²) / v³ = 0;
∴ u = v or 1/3 v.

The former gives a minimum, the latter a maximum efficiency.

Putting u = 1/3v in the expression above,

η max. = 8/27.

(3) If, instead of one plane moving before the jet, a series of planes are introduced at short intervals at the same point, the quantity of water impinging on the series will be ωv instead of ω(v − u), and the whole pressure = (G/g)ωv (v − u). The work done is (G/g)ωvu (v − u). The efficiency η = (G/g)ωvu (v − u) ÷ (G/2g)ωv³ = 2u(v–u)/v². This becomes a maximum for dη/du = 2(v − 2u) = 0, or u = 1/2v, and the η = 1/2. This result is often used as an approximate expression for the velocity of greatest efficiency when a jet of water strikes the floats of a water wheel. The work wasted in this case is half the whole energy of the jet when the floats run at the best speed.

§ 156. (4) Case of a Jet impinging on a Concave Cup Vane, velocity of water v, velocity of vane in the same direction u (fig. 155), weight impinging per second = Gw (v − u).

Fig. 155.

If the cup is hemispherical, the water leaves the cup in a direction parallel to the jet. Its relative velocity is v − u when approaching the cup, and −(v − u) when leaving it. Hence its absolute velocity when leaving the cup is u − (v − u) = 2u − v. The change of momentum per second = (G/g)ω (v − u) {v − (2u − v)} = 2(G/g)ω (v − u)². Comparing this with case 2, it is seen that the pressure on a hemispherical cup is double that on a flat plane. The work done on the cup = 2(G/g)ω (v − u) ²u foot-pounds per second. The efficiency of the jet is greatest when v = 3u; in that case the efficiency = 16/27.

If a series of cup vanes are introduced in front of the jet, so that the quantity of water acted upon is ωv instead of ω(v − u), then the whole pressure on the chain of cups is (G/g)ωv {v − (2u − v)} = 2(G/g)ωv (v − u). In this case the efficiency is greatest when v = 2u, and the maximum efficiency is unity, or all the energy of the water is expended on the cups.

Fig. 156.

§ 157. (5) Case of a Flat Vane oblique to the Jet (fig. 156).—This case presents some difficulty. The water spreading on the plane in all directions from the point of impact, different particles leave the plane with different absolute velocities. Let AB = v = velocity of water, AC = u = velocity of plane. Then, completing the parallelogram, AD represents in magnitude and direction the relative velocity of water and plane. Draw AE normal to the plane and DE parallel to the plane. Then the relative velocity AD may be regarded as consisting of two components, one AE normal, the other DE parallel to the plane. On the assumption that friction is insensible, DE is unaffected by impact, but AE is destroyed. Hence AE represents the entire change of velocity due to impact and the direction of that change. The pressure on the plane is in the direction AE, and its amount is = mass of water impinging per second × AE.

Let DAE = θ, and let AD = v_r. Then AE = v_r cos θ; DE = v_r sin θ. If Q is the volume of water impinging on the plane per second, the change of momentum is (G/g) Qv_r cos θ. Let AC = u = velocity of the plane, and let AC make the angle CAE = δ with the normal to the plane. The velocity of the plane in the direction AE = u cos δ. The work of the jet on the plane = (G/g) Qv_r cos θ u cos δ. The same problem may be thus treated algebraically (fig. 157). Let BAF = α, and CAF = δ. The velocity v of the water may be decomposed into AF = v cos α normal to the plane, and FB = v sin α parallel to the plane. Similarly the velocity of the plane = u = AC = BD can be decomposed into BG = FE = u cos δ normal to the plane, and DG = u sin δ parallel to the plane. As friction is neglected, the velocity of the water parallel to the plane is unaffected by the impact, but its component v cos α normal to the plane becomes after impact the same as that of the plane, that is, u cos δ. Hence the change of velocity during impact = AE = v cos α − u cos δ. The change of momentum per second, and consequently the normal pressure on the plane is N = (G/g) Q(v cos α − u cos δ). The pressure in the direction in which the plane is moving is P = N cos δ = (G/g)Q (v cos α − u cos δ) cos δ, and the work done on the plane is Pu = (G/g)Q(v cos α − u cos δ) u cos δ, which is the same expression as before, since AE = v_r cos θ = v cos α − u cos δ.

Fig. 157.

Fig. 158.

In one second the plane moves so that the point A (fig. 158) comes to C, or from the position shown in full lines to the position shown in dotted lines. If the plane remained stationary, a length AB = v of the jet would impinge on the plane, but, since the plane moves in the same direction as the jet, only the length HB = AB − AH impinges on the plane.

But AH = AC cos δ / cos α = u cos δ / cos α, and therefore HB = v − u cos δ / cos α. Let ω = sectional area of jet; volume impinging on plane per second = Q = ω(v − u cos δ / cos α) = ω (v cos α − u cos δ) / cos α. Inserting this in the formulae above, we get

N = G/g ω/cos α(v cos α − u cos δ)2;

(1)

P =	G		ω cos δ	(v cos α − u cos δ)2;
	g		cos α

(2)

Pu =	G	ωu	cos δ	(v cos α − u cos δ)2;
	g		cos α

(3)

Three cases may be distinguished:—

(a) The plane is at rest. Then u = 0, N = (G/g)ωv²cos α; and the work done on the plane and the efficiency of the jet are zero.

(b) The plane moves parallel to the jet. Then δ = α, and Pu = (G/g)ωu cos² α (v − u)², which is a maximum when u = 1/3v.

When u = 1/3v then Pu max. = 4/27(G/g)ωv³ cos²α, and the efficiency = η = 4/9cos²α.

(c) The plane moves perpendicularly to the jet. Then δ = 90° − α; cos δ = sin α; and Pu = G/g ωu (sin α / cos α) (v cos α − u sin α)². This is a maximum when u = 1/3v cos α.

When u = 1/3v cos α, the maximum work and the efficiency are the same as in the last case.

Fig. 159.

§ 158. Best Form of Vane to receive Water.—When water impinges normally or obliquely on a plane, it is scattered in all directions after impact, and the work carried away by the water is then generally lost, from the impossibility of dealing afterwards with streams of water deviated in so many directions. By suitably forming the vane, however, the water may be entirely deviated in one direction, and the loss of energy from agitation of the water is entirely avoided.

Let AB (fig. 159) be a vane, on which a jet of water impinges at the point A and in the direction AC. Take AC = v = velocity of water, and let AD represent in magnitude and direction the velocity of the vane. Completing the parallelogram, DC or AE represents the direction in which the water is moving relatively to the vane. If the lip of the vane at A is tangential to AE, the water will not have its direction suddenly changed when it impinges on the vane, and will therefore have no tendency to spread laterally. On the contrary it will be so gradually deviated that it will glide up the vane in the direction AB. This is sometimes expressed by saying that the vane receives the water without shock.

Fig. 160.

§ 159. Floats of Poncelet Water Wheels.—Let AC (fig. 160) represent the direction of a thin horizontal stream of water having the velocity v. Let AB be a curved float moving horizontally with velocity u. The relative motion of water and float is then initially horizontal, and equal to v − u.

In order that the float may receive the water without shock, it is necessary and sufficient that the lip of the float at A should be tangential to the direction AC of relative motion. At the end of (v − u)/g seconds the float moving with the velocity u comes to the position A₁B₁, and during this time a particle of water received at A and gliding up the float with the relative velocity v − u, attains a height DE = (v − u)²/2g. At E the water comes to relative rest. It then descends along the float, and when after 2(v − u)/g seconds the float has come to A₂B₂ the water will again have reached the lip at A₂ and will quit it tangentially, that is, in the direction CA₂, with a relative velocity −(v − u) = −√ (2gDE) acquired under the influence of gravity. The absolute velocity of the water leaving the float is therefore u − (v − u) = 2u − v. If u = 1/2v, the water will drop off the bucket deprived of all energy of motion. The whole of the work of the jet must therefore have been expended in driving the float. The water will have been received without shock and discharged without velocity. This is the principle of the Poncelet wheel, but in that case the floats move over an arc of a large circle; the stream of water has considerable thickness (about 8 in.); in order to get the water into and out of the wheel, it is then necessary that the lip of the float should make a small angle (about 15°) with the direction of its motion. The water quits the wheel with a little of its energy of motion remaining.

§ 160. Pressure on a Curved Surface when the Water is deviated wholly in one Direction.—When a jet of water impinges on a curved surface in such a direction that it is received without shock, the pressure on the surface is due to its gradual deviation from its first direction. On any portion of the area the pressure is equal and opposite to the force required to cause the deviation of so much water as rests on that surface. In common language, it is equal to the centrifugal force of that quantity of water.

Fig. 161.

Case 1. Surface Cylindrical and Stationary.—Let AB (fig. 161) be the surface, having its axis at O and its radius = r. Let the water impinge at A tangentially, and quit the surface tangentially at B. Since the surface is at rest, v is both the absolute velocity of the water and the velocity relatively to the surface, and this remains unchanged during contact with the surface, because the deviating force is at each point perpendicular to the direction of motion. The water is deviated through an angle BCD = AOB = φ. Each particle of water of weight p exerts radially a centrifugal force pv²/rg. Let the thickness of the stream = t ft. Then the weight of water resting on unit of surface = Gt ℔; and the normal pressure per unit of surface = n = Gtv²/gr. The resultant of the radial pressures uniformly distributed from A to B will be a force acting in the direction OC bisecting AOB, and its magnitude will equal that of a force of intensity = n, acting on the projection of AB on a plane perpendicular to the direction OC. The length of the chord AB = 2r sin 1/2φ; let b = breadth of the surface perpendicular to the plane of the figure. The resultant pressure on surface

= R = 2rb sin φ/2 × Gt/g·v²/r = 2G/gbtv² sin φ/2,

which is independent of the radius of curvature. It may be inferred that the resultant pressure is the same for any curved surface of the same projected area, which deviates the water through the same angle.

Case 2. Cylindrical Surface moving in the Direction AC with Velocity u.—The relative velocity = v − u. The final velocity BF (fig. 162) is found by combining the relative velocity BD = v − u tangential to the surface with the velocity BE = u of the surface. The intensity of normal pressure, as in the last case, is (G/g)t(v − u)²/r. The resultant normal pressure R = 2(G/g)bt(v−u)² sin 1/2φ. This resultant pressure may be resolved into two components P and L, one parallel and the other perpendicular to the direction of the vane’s motion. The former is an effort doing work on the vane. The latter is a lateral force which does no work.

P = R sin 1/2φ = (G/g)bt(v−u)² (1−cos φ);
 L = R cos 1/2φ = (G/g)bt(v−u)² sin φ.

Fig. 162.

Fig. 163.

The work done by the jet on the vane is Pu = (G/g)btu(v − u)²(1 − cos φ), which is a maximum when u = 1/3v. This result can also be obtained by considering that the work done on the plane must be equal to the energy lost by the water, when friction is neglected.

If φ = 180°, cos φ = −1, 1 − cos φ = 2; then P = 2(G/g)bt(v − u)², the same result as for a concave cup.

§ 161. Position which a Movable Plane takes in Flowing Water.—When a rectangular plane, movable about an axis parallel to one of its sides, is placed in an indefinite current of fluid, it takes a position such that the resultant of the normal pressures on the two sides of the axis passes through the axis. If, therefore, planes pivoted so that the ratio a/b (fig. 163) is varied are placed in water, and the angle they make with the direction of the stream is observed, the position of the resultant of the pressures on the plane is determined for different angular positions. Experiments of this kind have been made by Hagen. Some of his results are given in the following table:—

	Larger plane.	Smaller Plane.
a/b = 1.0	φ = . . .	φ = 90°
0.9	75°	721/2°
0.8	60°	57°
0.7	48°	43°
0.6	25°	29°
0.5	13°	13°
0.4	8°	61/2°
0.3	61/2°	..
0.2	4°	..

§ 162. Direct Action distinguished from Reaction (Rankine, Steam Engine, § 147).

The pressure which a jet exerts on a vane can be distinguished into two parts, viz.:—

(1) The pressure arising from changing the direct component of the velocity of the water into the velocity of the vane. In fig. 153, § 154, ab cos bae is the direct component of the water’s velocity, or component in the direction of motion of vane. This is changed into the velocity ae of the vane. The pressure due to direct impulse is then

P₁ = GQ (ab cos bae − ae)/g.

For a flat vane moving normally, this direct action is the only action producing pressure on the vane.

(2) The term reaction is applied to the additional action due to the direction and velocity with which the water glances off the vane. It is this which is diminished by the friction between the water and the vane. In Case 2, § 160, the direct pressure is

P₁ = Gbt (v − u)²/g.

That due to reaction is

P₂ = −Gbt (v − u)² cos φ/g.

If φ < 90°, the direct component of the water’s motion is not wholly converted into the velocity of the vane, and the whole pressure due to direct impulse is not obtained. If φ > 90°, cos φ is negative and an additional pressure due to reaction is obtained.

Fig. 164.

§ 163. Jet Propeller.—In the case of vessels propelled by a jet of water (fig. 164), driven sternwards from orifices at the side of the vessel, the water, originally at rest outside the vessel, is drawn into the ship and caused to move with the forward velocity V of the ship. Afterwards it is projected sternwards from the jets with a velocity v relatively to the ship, or v − V relatively to the earth. If Ω is the total sectional area of the jets, Ωv is the quantity of water discharged per second. The momentum generated per second in a sternward direction is (G/g) Ωv (v − V), and this is equal to the forward acting reaction P which propels the ship.

The energy carried away by the water

= 1/2 (G/g) Ωv (v − V)2.

(1)

The useful work done on the ship

PV = (G/g) Ωv (v − V) V.

(2)

Adding (1) and (2), we get the whole work expended on the water, neglecting friction:—

W = 1/2 (G/g) Ωv (v² − V²). 

Hence the efficiency of the jet propeller is

PV/W = 2V / (v + V).

(3)

This increases towards unity as v approaches V. In other words, the less the velocity of the jets exceeds that of the ship, and therefore the greater the area of the orifice of discharge, the greater is the efficiency of the propeller.

In the “Waterwitch” v was about twice V. Hence in this case the theoretical efficiency of the propeller, friction neglected, was about 2/3.

Fig. 165.

§ 164. Pressure of a Steady Stream in a Uniform Pipe on a Plane normal to the Direction of Motion.—Let CD (fig. 165) be a plane placed normally to the stream which, for simplicity, may be supposed to flow horizontally. The fluid filaments are deviated in front of the plane, form a contraction at A₁A₁, and converge again, leaving a mass of eddying water behind the plane. Suppose the section A₀A₀ taken at a point where the parallel motion has not begun to be disturbed, and A₂A₂ where the parallel motion is re-established. Then since the same quantity of water with the same velocity passes A₀A₀, A₂A₂ in any given time, the external forces produce no change of momentum on the mass A₀A₀A₂A₂, and must therefore be in equilibrium. If Ω is the section of the stream at A₀A₀ or A₂A₂, and ω the area of the plate CD, the area of the contracted section of the stream at A₁A₁ will be c_c(Ω − ω), where c_c is the coefficient of contraction. Hence, if v is the velocity at A₀A₀ or A₂A₂, and v₁ the velocity at A₁A₁,

vΩ = c_cv (Ω − ω);

∴ v₁ = vΩ / c_c (Ω − ω).

(1)

Let p₀, p₁, p₂ be the pressures at the three sections. Applying Bernoulli’s theorem to the sections A₀A₀ and A₁A₁,

p0	+	v2	=	p1	+	v12	.
G		2g		G		2g

Also, for the sections A₁A₁ and A₂A₂, allowing that the head due to the relative velocity v₁ − v is lost in shock:—

p1	+	v12	=	p2	+	v2	+	(v1 − v)2	;
G		2g		G		2g		2g

∴ p₀ − p₂ = G (v₁ − v)² / 2g;

(2)

or, introducing the value in (1),

p0 − p2 =	G	${\Big (}$	Ω	− 1 ${\Big )}$	2	v2
	2g		cc (Ω − ω)

(3)

Now the external forces in the direction of motion acting on the mass A₀A₀A₂A₂ are the pressures p₀Ω₁ − p₂Ω at the ends, and the reaction −R of the plane on the water, which is equal and opposite to the pressure of the water on the plane. As these are in equilibrium,

(p₀ − p₂) Ω − R = 0;

∴ R = GΩ ${\Big (}$	Ω	− 1 ${\Big )}$	2		v2	;
	cc (Ω − ω)				2g

(4)

an expression like that for the pressure of an isolated jet on an indefinitely extended plane, with the addition of the term in brackets, which depends only on the areas of the stream and the plane. For a given plane the expression in brackets diminishes as Ω increases. If Ω/ω = ρ, the equation (4) becomes

R = Gω	v2	${\Big \{}$ ρ ${\Big (}$	ρ	− 1 ${\Big )}$	2	${\Big \}}$,
	2g		cc (ρ − 1)

(4a)

which is of the form

R = Gω (v²/2g) K,

where K depends only on the ratio of the sections of the stream and plane.

For example, let c_c = 0.85, a value which is probable, if we allow that the sides of the pipe act as internal borders to an orifice. Then

K = ρ ${\Big (}$ 1.176ρ/ρ − 1− 1 ${\Big )}$²

ρ =	K =
1	∞
2	3.66
3	1.75
4	1.29
5	1.10
10	.94
50	2.00
100	3.50

The assumption that the coefficient of contraction c_c is constant for different values of ρ is probably only true when ρ is not very large. Further, the increase of K for large values of ρ is contrary to experience, and hence it may be inferred that the assumption that all the filaments have a common velocity v₁ at the section A₁A₁ and a common velocity v at the section A₂A₂ is not true when the stream is very much larger than the plane. Hence, in the expression

R = KGωv² / 2g,

K must be determined by experiment in each special case. For a cylindrical body putting ω for the section, c_c for the coefficient of contraction, c_c (Ω − ω) for the area of the stream at A₁A₁,

v₁ = vΩ / c_c (Ω − ω); v₂ = vΩ / (Ω −ω);

or, putting ρ = Ω/ω,

v₁ = vρ / c_c (ρ − 1), v₂ = vρ / (ρ − 1).

Then

R = K₁Gωv² / 2g,

where

K1 = ρ ${\Big \{}{\Big (}$

ρ

${\Big )}$

2

${\Big (}$

1

− 1 ${\Big )}$

2

+ ${\Big (}$

ρ

− 1 ${\Big )}$

2

${\Big \}}$.

ρ − 1

cc

ρ − 1

Taking c_c = 0.85 and ρ = 4, K₁ = 0.467, a value less than before. Hence there is less pressure on the cylinder than on the thin plane.

Fig. 166.

§ 165. Distribution of Pressure on a Surface on which a Jet impinges normally.—The principle of momentum gives readily enough the total or resultant pressure of a jet impinging on a plane surface, but in some cases it is useful to know the distribution of the pressure. The problem in the case in which the plane is struck normally, and the jet spreads in all directions, is one of great complexity, but even in that case the maximum intensity of the pressure is easily assigned. Each layer of water flowing from an orifice is gradually deviated (fig. 166) by contact with the surface, and during deviation exercises a centrifugal pressure towards the axis of the jet. The force exerted by each small mass of water is normal to its path and inversely as the radius of curvature of the path. Hence the greatest pressure on the plane must be at the axis of the jet, and the pressure must decrease from the axis outwards, in some such way as is shown by the curve of pressure in fig. 167, the branches of the curve being probably asymptotic to the plane.

For simplicity suppose the jet is a vertical one. Let h₁ (fig. 167) be the depth of the orifice from the free surface, and v₁ the velocity of discharge. Then, if ω is the area of the orifice, the quantity of water impinging on the plane is obviously

Q = ωv₁ = ω √ (2gh₁);

that is, supposing the orifice rounded, and neglecting the coefficient of discharge.

The velocity with which the fluid reaches the plane is, however, greater than this, and may reach the value

v = √ (2gh);

where h is the depth of the plane below the free surface. The external layers of fluid subjected throughout, after leaving the orifice, to the atmospheric pressure will attain the velocity v, and will flow away with this velocity unchanged except by friction. The layers towards the interior of the jet, being subjected to a pressure greater than atmospheric pressure, will attain a less velocity, and so much less as they are nearer the centre of the jet. But the pressure can in no case exceed the pressure v²/2g or h measured in feet of water, or the direction of motion of the water would be reversed, and there would be reflux. Hence the maximum intensity of the pressure of the jet on the plane is h ft. of water. If the pressure curve is drawn with pressures represented by feet of water, it will touch the free water surface at the centre of the jet.

Fig. 167.

Suppose the pressure curve rotated so as to form a solid of revolution. The weight of water contained in that solid is the total pressure of the jet on the surface, which has already been determined. Let V = volume of this solid, then GV is its weight in pounds. Consequently

GV = (G/g) ωv₁v;
V = 2ω √ (hh₁).

We have already, therefore, two conditions to be satisfied by the pressure curve.

Fig. 168.—Curves of Pressure of Jets impinging normally on a Plane.

Some very interesting experiments on the distribution of pressure on a surface struck by a jet have been made by J. S. Beresford (Prof. Papers on Indian Engineering, No. cccxxii.), with a view to afford information as to the forces acting on the aprons of weirs. Cylindrical jets 1/2 in. to 2 in. diameter, issuing from a vessel in which the water level was constant, were allowed to fall vertically on a brass plate 9 in. in diameter. A small hole in the brass plate communicated by a flexible tube with a vertical pressure column. Arrangements were made by which this aperture could be moved 1/20 in. at a time across the area struck by the jet. The height of the pressure column, for each position of the aperture, gave the pressure at that point of the area struck by the jet. When the aperture was exactly in the axis of the jet, the pressure column was very nearly level with the free surface in the reservoir supplying the jet; that is, the pressure was very nearly v²/2g. As the aperture moved away from the axis of the jet, the pressure diminished, and it became insensibly small at a distance from the axis of the jet about equal to the diameter of the jet. Hence, roughly, the pressure due to the jet extends over an area about four times the area of section of the jet.

Fig. 168 shows the pressure curves obtained in three experiments with three jets of the sizes shown, and with the free surface level in the reservoir at the heights marked.

Height from Free Surface to Brass Plate in inches.	Distance from Axis of Jet in inches.	Pressure in inches of Water.
Experiment 1. Jet .475 in. diameter.
43	0	40.5
”	.05	39.40
”	.1	37.5-39.5
”	.15	35
”	.2	33.5-37
”	.25	31
”	.3	21-27
”	.35	21
”	.4	14
”	.45	8
”	.5	3.5
”	.55	1
”	.6	0.5
”	.65	0
Experiment 2. Jet .988 in. diameter.
42.15	0	42
”	.05	41.9
”	.1	41.5-41.8
”	.15	41
”	.2	40.3
”	.25	39.2
”	.3	37.5
”	.35	34.8
”	.45	27
42.25	.5	23
”	.55	18.5
”	.6	13
”	.65	8.3
”	.7	5
”	.75	3
”	.8	2.2
42.15	.85	1.6
”	.95	1
Experiment 3. Jet 1.95 in. diameter.
27.15	0	26.9
”	.08	26.9
”	.13	26.8
”	.18	26.5-26.6
”	.23	26.4-26.5
”	.28	26.3-26.6
27	.33	26.2
”	.38	25.9
”	.43	25.5
”	.48	25
”	.53	24.5
”	.58	24
”	.63	23.3
”	.68	22.5
”	.73	21.8
”	.78	21
”	.83	20.3
”	.88	19.3
”	.93	18
”	.98	17
26.5	1.13	13.5
”	1.18	12.5
”	1.23	10.8
”	1.28	9.5
”	1.33	8
”	1.38	7
”	1.43	6.3
”	1.48	5
”	1.53	4.3
”	1.58	3.5
”	1.9	2

As the general form of the pressure curve has been already indicated, it may be assumed that its equation is of the form

y = ab^−x2.

But it has already been shown that for x = 0, y = h, hence a = h. To determine the remaining constant, the other condition may be used, that the solid formed by rotating the pressure curve represents the total pressure on the plane. The volume of the solid is

V = $\int _{0}^{\infty }$ 2πxydx

= 2πh $\int _{0}^{\infty }$ b^−x2xdx

= (πh / log_e b) ${\Big [}$ −b^−x2 ${\Big ]}_{0}^{\infty }$

= πh / log_e b.

Using the condition already stated,

2ω √ (hh₁) = πh / log_e b,
log_ε b = (π/2ω) √ (h/h₁).

Putting the value of b in (2) in eq. (1), and also r for the radius of the jet at the orifice, so that ω = πr ², the equation to the pressure curve is

y = hε^−1/2 √(h / h₁) (x² / r ²).

§ 166. Resistance of a Plane moving through a Fluid, or Pressure of a Current on a Plane.—When a thin plate moves through the air, or through an indefinitely large mass of still water, in a direction normal to its surface, there is an excess of pressure on the anterior face and a diminution of pressure on the posterior face. Let v be the relative velocity of the plate and fluid, Ω the area of the plate, G the density of the fluid, h the height due to the velocity, then the total resistance is expressed by the equation

R = f GΩv² / 2g pounds = f GΩh;

where f is a coefficient having about the value 1.3 for a plate moving in still fluid, and 1.8 for a current impinging on a fixed plane, whether the fluid is air or water. The difference in the value of the coefficient in the two cases is perhaps due to errors of experiment. There is a similar resistance to motion in the case of all bodies of “unfair“ form, that is, in which the surfaces over which the water slides are not of gradual and continuous curvature.

The stress between the fluid and plate arises chiefly in this way. The streams of fluid deviated in front of the plate, supposed for definiteness to be moving through the fluid, receive from it forward momentum. Portions of this forward moving water are thrown off laterally at the edges of the plate, and diffused through the surrounding fluid, instead of falling to their original position behind the plate. Other portions of comparatively still water are dragged into motion to fill the space left behind the plate; and there is thus a pressure less than hydrostatic pressure at the back of the plate. The whole resistance to the motion of the plate is the sum of the excess of pressure in front and deficiency of pressure behind. This resistance is independent of any friction or viscosity in the fluid, and is due simply to its inertia resisting a sudden change of direction at the edge of the plate.

Experiments made by a whirling machine, in which the plate is fixed on a long arm and moved circularly, gave the following values of the coefficient f. The method is not free from objection, as the centrifugal force causes a flow outwards across the plate.

Approximate  Area of Plate  in sq. ft.	Values of f.
	Borda.	Hutton.	Thibault.
0.13	1.39	1.24	..
0.25	1.49	1.43	1.525
0.63	1.64	..	..
1.11	..	..	1.784

There is a steady increase of resistance with the size of the plate, in part or wholly due to centrifugal action.

P. L. G. Dubuat (1734–1809) made experiments on a plane 1 ft. square, moved in a straight line in water at 3 to 61/2 ft. per second. Calling m the coefficient of excess of pressure in front, and n the coefficient of deficiency of pressure behind, so that f = m + n, he found the following values:—

m = 1; n = 0.433; f = 1.433.

The pressures were measured by pressure columns. Experiments by A. J. Morin (1795–1880), G. Piobert (1793–1871) and I. Didion (1798–1878) on plates of 0.3 to 2.7 sq. ft. area, drawn vertically through water, gave f = 2.18; but the experiments were made in a reservoir of comparatively small depth. For similar plates moved through air they found f = 1.36, a result more in accordance with those which precede.

For a fixed plane in a moving current of water E. Mariotte found f = 1.25. Dubuat, in experiments in a current of water like those mentioned above, obtained the values m = 1.186; n = 0.670; f = 1.856. Thibault exposed to wind pressure planes of 1.17 and 2.5 sq. ft. area, and found f to vary from 1.568 to 2.125, the mean value being f = 1.834, a result agreeing well with Dubuat.

Fig. 169.

§ 167. Stanton’s Experiments on the Pressure of Air on Surfaces.—At the National Physical Laboratory, London, T. E. Stanton carried out a series of experiments on the distribution of pressure on surfaces in a current of air passing through an air trunk. These were on a small scale but with exceptionally accurate means of measurement. These experiments differ from those already given in that the plane is small relatively to the cross section of the current (Proc. Inst. Civ. Eng. clvi., 1904). Fig. 169 shows the distribution of pressure on a square plate. ab is the plate in vertical section. acb the distribution of pressure on the windward and adb that on the leeward side of the central section. Similarly aeb is the distribution of pressure on the windward and afb on the leeward side of a diagonal section. The intensity of pressure at the centre of the plate on the windward side was in all cases p = Gv ²/2g ℔ per sq. ft., where G is the weight of a cubic foot of air and v the velocity of the current in ft. per sec. On the leeward side the negative pressure is uniform except near the edges, and its value depends on the form of the plate. For a circular plate the pressure on the leeward side was 0.48 Gv ²/2g and for a rectangular plate 0.66 Gv ²/2g. For circular or square plates the resultant pressure on the plate was P = 0.00126 v ² ℔ per sq. ft. where v is the velocity of the current in ft. per sec. On a long narrow rectangular plate the resultant pressure was nearly 60% greater than on a circular plate. In later tests on larger planes in free air, Stanton found resistances 18% greater than those observed with small planes in the air trunk.

§ 168. Case when the Direction of Motion is oblique to the Plane.—The determination of the pressure between a fluid and surface in this case is of importance in many practical questions, for instance, in assigning the load due to wind pressure on sloping and curved roofs, and experiments have been made by Hutton, Vince, and Thibault on planes moved circularly through air and water on a whirling machine.

Fig. 170.

Let AB (fig. 170) be a plane moving in the direction R making an angle φ with the plane. The resultant pressure between the fluid and the plane will be a normal pressure N. The component R of this normal pressure is the resistance to the motion of the plane and the other component L is a lateral force resisted by the guides which support the plane. Obviously

R = N sin φ;

L = N cos φ.

In the case of wind pressure on a sloping roof surface, R is the horizontal and L the vertical component of the normal pressure.

In experiments with the whirling machine it is the resistance to motion, R, which is directly measured. Let P be the pressure on a plane moved normally through a fluid. Then, for the same plane inclined at an angle φ to its direction of motion, the resistance was found by Hutton to be

R = P (sin φ)^{1.842 cos φ}.

A simpler and more convenient expression given by Colonel Duchemin is

R = 2P sin² φ / (1 + sin² φ).

Consequently, the total pressure between the fluid and plane is

N = 2P sin φ / (1 + sin² φ) = 2P / (cosec φ + sin φ),

and the lateral force is

L = 2P sin φ cos φ / (1 + sin² φ).

In 1872 some experiments were made for the Aeronautical Society on the pressure of air on oblique planes. These plates, of 1 to 2 ft. square, were balanced by ingenious mechanism designed by F. H. Wenham and Spencer Browning, in such a manner that both the pressure in the direction of the air current and the lateral force were separately measured. These planes were placed opposite a blast from a fan issuing from a wooden pipe 18 in. square. The pressure of the blast varied from 6/10 to 1 in. of water pressure. The following are the results given in pounds per square foot of the plane, and a comparison of the experimental results with the pressures given by Duchemin’s rule. These last values are obtained by taking P = 3.31, the observed pressure on a normal surface:—

Angle between Plane and Direction of Blast	15°	20°	60°	90°
Horizontal pressure R	0.4	0.61	2.73	3.31
Lateral pressure L	1.6	1.96	1.26	..
Normal pressure √ (L2 + R2)	1.65	2.05	3.01	3.31
Normal pressure by Duchemin’s rule	1.605	2.027	3.276	3.31

Water Motors

In every system of machinery deriving energy from a natural waterfall there exist the following parts:—

1. A supply channel or head race, leading the water from the highest accessible level to the site of the machine. This may be an open channel of earth, masonry or wood, laid at as small a slope as is consistent with the delivery of the necessary supply of water, or it may be a closed cast or wrought-iron pipe, laid at the natural slope of the ground, and about 3 ft. below the surface. In some cases part of the head race is an open channel, part a closed pipe. The channel often starts from a small storage reservoir, constructed near the stream supplying the water motor, in which the water accumulates when the motor is not working. There are sluices or penstocks by which the supply can be cut off when necessary.

2. Leading from the motor there is a tail race, culvert, or discharge pipe delivering the water after it has done its work at the lowest convenient level.

3. A waste channel, weir, or bye-wash is placed at the origin of the head race, by which surplus water, in floods, escapes.

4. The motor itself, of one of the kinds to be described presently, which either overcomes a useful resistance directly, as in the case of a ram acting on a lift or crane chain, or indirectly by actuating transmissive machinery, as when a turbine drives the shafting, belting and gearing of a mill. With the motor is usually combined regulating machinery for adjusting the power and speed to the work done. This may be controlled in some cases by automatic governing machinery.

§ 169. Water Motors with Artificial Sources of Energy.—The great convenience and simplicity of water motors has led to their adoption in certain cases, where no natural source of water power is available. In these cases, an artificial source of water power is created by using a steam-engine to pump water to a reservoir at a great elevation, or to pump water into a closed reservoir in which there is great pressure. The water flowing from the reservoir through hydraulic engines gives back the energy expended, less so much as has been wasted by friction. Such arrangements are most useful where a continuously acting steam engine stores up energy by pumping the water, while the work done by the hydraulic engines is done intermittently.

§ 170. Energy of a Water-fall.—Let H_t be the total fall of level from the point where the water is taken from a natural stream to the point where it is discharged into it again. Of this total fall a portion, which can be estimated independently, is expended in overcoming the resistances of the head and tail races or the supply and discharge pipes. Let this portion of head wasted be ɧ_r. Then the available head to work the motor is H = H_t − ɧ_r. It is this available head which should be used in all calculations of the proportions of the motor. Let Q be the supply of water per second. Then GQH foot-pounds per second is the gross available work of the fall. The power of the fall may be utilized in three ways. (a) The GQ pounds of water may be placed on a machine at the highest level, and descending in contact with it a distance of H ft., the work done will be (neglecting losses from friction or leakage) GQH foot-pounds per second. (b) Or the water may descend in a closed pipe from the higher to the lower level, in which case, with the same reservation as before, the pressure at the foot of the pipe will be p = GH pounds per square foot. If the water with this pressure acts on a movable piston like that of a steam engine, it will drive the piston so that the volume described is Q cubic feet per second. Then the work done will be pQ = GHQ foot-pounds per second as before. (c) Or lastly, the water may be allowed to acquire the velocity v = √2gH by its descent. The kinetic energy of Q cubic feet will then be 1/2GQv²/g = GQH, and if the water is allowed to impinge on surfaces suitably curved which bring it finally to rest, it will impart to these the same energy as in the previous cases. Motors which receive energy mainly in the three ways described in (a), (b), (c) may be termed gravity, pressure and inertia motors respectively. Generally, if Q ft. per second of water act by weight through a distance h₁, at a pressure p due to h₂ ft. of fall, and with a velocity v due to h₃ ft. of fall, so that h₁ + h₂ + h₃ = H, then, apart from energy wasted by friction or leakage or imperfection of the machine, the work done will be

GQh₁ + pQ + (G/g) Q (v²/2g) = GQH foot pounds,

the same as if the water acted simply by its weight while descending H ft.

§ 171. Site for Water Motor.—Wherever a stream flows from a higher to a lower level it is possible to erect a water motor. The amount of power obtainable depends on the available head and the supply of water. In choosing a site the engineer will select a portion of the stream where there is an abrupt natural fall, or at least a considerable slope of the bed. He will have regard to the facility of constructing the channels which are to convey the water, and will take advantage of any bend in the river which enables him to shorten them. He will have accurate measurements made of the quantity of water flowing in the stream, and he will endeavour to ascertain the average quantity available throughout the year, the minimum quantity in dry seasons, and the maximum for which bye-wash channels must be provided. In many cases the natural fall can be increased by a dam or weir thrown across the stream. The engineer will also examine to what extent the head will vary in different seasons, and whether it is necessary to sacrifice part of the fall and give a steep slope to the tail race to prevent the motor being drowned by backwater in floods. Streams fed from lakes which form natural reservoirs or fed from glaciers are less variable than streams depending directly on rainfall, and are therefore advantageous for water-power purposes.

§ 172. Water Power at Holyoke, U.S.A.—About 85 m. from the mouth of the Connecticut river there was a fall of about 60 ft. in a short distance, forming what were called the Grand Rapids, below which the river turned sharply, forming a kind of peninsula on which the city of Holyoke is built. In 1845 the magnitude of the water-power available attracted attention, and it was decided to build a dam across the river. The ordinary flow of the river is 6000 cub. ft. per sec., giving a gross power of 30,000 h.p. In dry seasons the power is 20,000 h.p., or occasionally less. From above the dam a system of canals takes the water to mills on three levels. The first canal starts with a width of 140 ft. and depth of 22 ft., and supplies the highest range of mills. A second canal takes the water which has driven turbines in the highest mills and supplies it to a second series of mills. There is a third canal on a still lower level supplying the lowest mills. The water then finds its way back to the river. With the grant of a mill site is also leased the right to use the water-power. A mill-power is defined as 38 cub. ft. of water per sec. during 16 hours per day on a fall of 20 ft. This gives about 60 h.p. effective. The charge for the power water is at the rate of 20s. per h.p. per annum.

§ 173. Action of Water in a Water Motor.—Water motors may be divided into water-pressure engines, water-wheels and turbines.

Water-pressure engines are machines with a cylinder and piston or ram, in principle identical with the corresponding part of a steam-engine. The water is alternately admitted to and discharged from the cylinder, causing a reciprocating action of the piston or plunger. It is admitted at a high pressure and discharged at a low one, and consequently work is done on the piston. The water in these machines never acquires a high velocity, and for the most part the kinetic energy of the water is wasted. The useful work is due to the difference of the pressure of admission and discharge, whether that pressure is due to the weight of a column of water of more or less considerable height, or is artificially produced in ways to be described presently.

Water-wheels are large vertical wheels driven by water falling from a higher to a lower level. In most water-wheels, the water acts directly by its weight loading one side of the wheel and so causing rotation. But in all water-wheels a portion, and in some a considerable portion, of the work due to gravity is first employed to generate kinetic energy in the water; during its action on the water-wheel the velocity of the water diminishes, and the wheel is therefore in part driven by the impulse due to the change of the water’s momentum. Water-wheels are therefore motors on which the water acts, partly by weight, partly by impulse.

Turbines are wheels, generally of small size compared with water wheels, driven chiefly by the impulse of the water. Before entering the moving part of the turbine, the water is allowed to acquire a considerable velocity; during its action on the turbine this velocity is diminished, and the impulse due to the change of momentum drives the turbine.

In designing or selecting a water motor it is not sufficient to consider only its efficiency in normal conditions of working. It is generally quite as important to know how it will act with a scanty water supply or a diminished head. The greatest difference in water motors is in their adaptability to varying conditions of working.

Water-pressure Engines.

§ 174. In these the water acts by pressure either due to the height of the column in a supply pipe descending from a high-level reservoir, or created by pumping. Pressure engines were first used in mine-pumping on waterfalls of greater height than could at that time be utilized by water wheels. Usually they were single acting, the water-pressure lifting the heavy pump rods which then made the return or pumping stroke by their own weight. To avoid losses by fluid friction and shock the velocity of the water in the pipes and passages was restricted to from 3 to 10 ft. per second, and the mean speed of plunger to 1 ft. per second. The stroke was long and the number of strokes 3 to 6 per minute. The pumping lift being constant, such engines worked practically always at full load, and the efficiency was high, about 84%. But they were cumbrous machines. They are described in Weisbach’s Mechanics of Engineering.

The convenience of distributing energy from a central station to scattered working-points by pressure water conveyed in pipes—a system invented by Lord Armstrong—has already been mentioned. This system has led to the development of a great variety of hydraulic pressure engines of very various types. The cost of pumping the pressure water to some extent restricts its use to intermittent operations, such as working lifts and cranes, punching, shearing and riveting machines, forging and flanging presses. To keep down the cost of the distributing mains very high pressures are adopted, generally 700 ℔ per sq. in. or 1600 ft. of head or more.

In a large number of hydraulic machines worked by water at high pressure, especially lifting machines, the motor consists of a direct, single acting ram and cylinder. In a few cases double-acting pistons and cylinders are used; but they involve a water-tight packing of the piston not easily accessible. In some cases pressure engines are used to obtain rotative movement, and then two double-acting cylinders or three single-acting cylinders are used, driving a crank shaft. Some double-acting cylinders have a piston rod half the area of the piston. The pressure water acts continuously on the annular area in front of the piston. During the forward stroke the pressure on the front of the piston balances half the pressure on the back. During the return stroke the pressure on the front is unopposed. The water in front of the piston is not exhausted, but returns to the supply pipe. As the frictional losses in a fluid are independent of the pressure, and the work done increases directly as the pressure, the percentage loss decreases for given velocities of flow as the pressure increases. Hence for high-pressure machines somewhat greater velocities are permitted in the passages than for low-pressure machines. In supply mains the velocity is from 3 to 6 ft. per second, in valve passages 5 to 10 ft. per second, or in extreme cases 20 ft. per second, where there is less object in economizing energy. As the water is incompressible, slide valves must have neither lap nor lead, and piston valves are preferable to ordinary slide valves. To prevent injurious compression from exhaust valves closing too soon in rotative engines with a fixed stroke, small self-acting relief valves are fitted to the cylinder ends, opening outwards against the pressure into the valve chest. Imprisoned water can then escape without over-straining the machines.

In direct single-acting lift machines, in which the stroke is fixed, and in rotative machines at constant speed it is obvious that the cylinder must be filled at each stroke irrespective of the amount of work to be done. The same amount of water is used whether much or little work is done, or whether great or small weights are lifted. Hence while pressure engines are very efficient at full load, their efficiency decreases as the load decreases. Various arrangements have been adopted to diminish this defect in engines working with a variable load. In lifting machinery there is sometimes a double ram, a hollow ram enclosing a solid ram. By simple arrangements the solid ram only is used for small loads, but for large loads the hollow ram is locked to the solid ram, and the two act as a ram of larger area. In rotative engines the case is more difficult. In Hastie’s and Rigg’s engines the stroke is automatically varied with the load, increasing when the load is large and decreasing when it is small. But such engines are complicated and have not achieved much success. Where pressure engines are used simplicity is generally a first consideration, and economy is of less importance.

§ 175. Efficiency of Pressure Engines.—It is hardly possible to form a theoretical expression for the efficiency of pressure engines, but some general considerations are useful. Consider the case of a long stroke hydraulic ram, which has a fairly constant velocity v during the stroke, and valves which are fairly wide open during most of the stroke. Let r be the ratio of area of ram to area of valve passage, a ratio which may vary in ordinary cases from 4 to 12. Then the loss in shock of the water entering the cylinder will be (r − 1)²v²/2g in ft. of head. The friction in the supply pipe is also proportional to v². The energy carried away in exhaust will be proportional to v². Hence the total hydraulic losses may be taken to be approximately ζv²/2g ft., where ζ is a coefficient depending on the proportions of the machine. Let f be the friction of the ram packing and mechanism reckoned in ℔ per sq. ft. of ram area. Then if the supply-pipe pressure driving the machine is p ℔ per sq. ft., the effective working pressure will be

p − Gζv² / 2g − f ℔ per sq. ft.

Let A be the area of the ram in sq. ft., v its velocity in ft. per sec. The useful work done will be

(p − Gζv² / 2g − f) Av ft. ℔ per sec.,

and the efficiency of the machine will be

η = (p − Gζv² / 2g − f) /p.

Fig. 171.

This shows that the efficiency increases with the pressure p, and diminishes with the speed v, other things being the same. If in regulating the engine for varying load the pressure is throttled, part of the available head is destroyed at the throttle valve, and p in the bracket above is reduced. Direct-acting hydraulic lifts, without intermediate gearing, may have an efficiency of 95% during the working stroke. If a hydraulic jigger is used with ropes and sheaves to change the speed of the ram to the speed of the lift, the efficiency may be only 50%. E. B. Ellington has given the efficiency of lifts with hydraulic balance at 85% during the working stroke. Large pressure engines have an efficiency of 85%, but small rotative engines probably not more than 50% and that only when fully loaded.

§ 176. Direct-Acting Hydraulic Lift (fig. 171).—This is the simplest of all kinds of hydraulic motor. A cage W is lifted directly by water pressure acting in a cylinder C, the length of which is a little greater than the lift. A ram or plunger R of the same length is attached to the cage. The water-pressure admitted by a cock to the cylinder forces up the ram, and when the supply valve is closed and the discharge valve opened, the ram descends. In this case the ram is 9 in. diameter, with a stroke of 49 ft. It consists of lengths of wrought-iron pipe screwed together perfectly water-tight, the lower end being closed by a cast-iron plug. The ram works in a cylinder 11 in. diameter of 9 ft. lengths of flanged cast-iron pipe. The ram passes water-tight through the cylinder cover, which is provided with double hat leathers to prevent leakage outwards or inwards. As the weight of the ram and cage is much more than sufficient to cause a descent of the cage, part of the weight is balanced. A chain attached to the cage passes over a pulley at the top of the lift, and carries at its free end a balance weight B, working in T iron guides. Water is admitted to the cylinder from a 4-in. supply pipe through a two-way slide, worked by a rack, spindle and endless rope. The lift works under 73 ft. of head, and lifts 1350 lb at 2 ft. per second. The efficiency is from 75 to 80%.

The principal prejudicial resistance to the motion of a ram of this kind is the friction of the cup leathers, which make the joint between the cylinder and ram. Some experiments by John Hick give for the friction of these leathers the following formula. Let F = the total friction in pounds; d = diameter of ram in ft.; p = water-pressure in pounds per sq. ft.; k a coefficient.

F = k p d

k = 0.00393 if the leathers are new or badly lubricated;

  = 0.00262 if the leathers are in good condition and well lubricated.

Since the total pressure on the ram is P = 1/4πd ²p, the fraction of the total pressure expended in overcoming the friction of the leathers is F/P = .005/d to .0033/d, d being in feet.

Let H be the height of the pressure column measured from the free surface of the supply reservoir to the bottom of the ram in its lowest position, H_b the height from the discharge reservoir to the same point, h the height of the ram above its lowest point at any moment, S the length of stroke, Ω the area of the ram, W the weight of cage, R the weight of ram, B the weight of balance weight, w the weight of balance chain per foot run, F the friction of the cup leather and slides. Then, neglecting fluid friction, if the ram is rising the accelerating force is

P₁ = G (H − h) Ω − R − W + B − w (S − h) + wh − F,

and if the ram is descending

P₂ = G (H_b − h) Ω + W + R − B + w (S − h) − wh − F.

If w = 1/2 GΩ, P₁ and P₂ are constant throughout the stroke; and the moving force in ascending and descending is the same, if

B = W + R + wS − GΩ (H + H_b) / 2.

Using the values just found for w and B,

P₁ = P₂ = 1/2 GΩ (H − H_b) − F.

Let W + R + wS + B = U, and let P be the constant accelerating force acting on the system, then the acceleration is (P/U)g. The velocity at the end of the stroke is (assuming the friction to be constant)

v = √ (2PgS/U);

and the mean velocity of ascent is 1/2v.

Fig. 172.

§ 177. Armstrong’s Hydraulic Jigger.—This is simply a single-acting hydraulic cylinder and ram, provided with sheaves so as to give motion to a wire rope or chain. It is used in various forms of lift and crane. Fig. 172 shows the arrangement. A hydraulic ram or plunger B works in a stationary cylinder A. Ram and cylinder carry sets of sheaves over which passes a chain or rope, fixed at one end to the cylinder, and at the other connected over guide pulleys to a lift or crane. For each pair of pulleys, one on the cylinder and one on the ram, the movement of the free end of the rope is doubled compared with that of the ram. With three pairs of pulleys the free end of the rope has a movement equal to six times the stroke of the ram, the force exerted being in the inverse proportion.

§ 178. Rotative Hydraulic Engines.—Valve-gear mechanism similar in principle to that of steam engines can be applied to actuate the admission and discharge valves, and the pressure engine is then converted into a continuously-acting motor.

Let H be the available fall to work the engine after deducting the loss of head in the supply and discharge pipes, Q the supply of water in cubic feet per second, and η the efficiency of the engine. Then the horse-power of the engine is

H.P. = ηGQH/550.

The efficiency of large slow-moving pressure engines is η = .66 to .8. In small motors of this kind probably η is not greater than .5. Let v be the mean velocity of the piston, then its diameter d is given by the relation

Q = πd ²v/4 in double-acting engines,
  = πd ²v/8 in single-acting engines.

If there are n cylinders put Q/n for Q in these equations.

Small rotative pressure engines form extremely convenient motors for hoists, capstans or winches, and for driving small machinery. The single-acting engine has the advantage that the pressure of the piston on the crank pin is always in one direction; there is then no knocking as the dead centres are passed. Generally three single-acting cylinders are used, so that the engine will readily start in all positions, and the driving effort on the crank pin is very uniform.

Fig. 173.
 

Fig. 174.
 

Fig. 175.

Brotherhood Hydraulic Engine.—Three cylinders at angles of 120° with each other are formed in one casting with the frame. The plungers are hollow trunks, and the connecting rods abut in cylindrical recesses in them and are connected to a common crank pin. A circular valve disk with concentric segmental ports revolves at the same rate as the crank over ports in the valve face common to the three cylinders. Each cylinder is always in communication with either an admission or exhaust port. The blank parts of the circular valve close the admission and exhaust ports alternately. The fixed valve face is of lignum vitae in a metal recess, and the revolving valve of gun-metal. In the case of a small capstan engine the cylinders are 31/2 in. diameter and 3 in. stroke. At 40 revs. per minute, the piston speed is 31 ft. per minute. The ports are 1 in. diameter or 1/12 of the piston area, and the mean velocity in the ports 6.4 ft. per sec. With 700 ℔ per sq. in. water pressure and an efficiency of 50%, the engine is about 3 h.p. A common arrangement is to have three parallel cylinders acting on a three-throw crank shaft, the cylinders oscillating on trunnions.

Hastie’s Engine.—Fig. 173 shows a similar engine made by Messrs Hastie of Greenock. G, G, G are the three plungers which pass out of the cylinders through cup leathers, and act on the same crank pin. A is the inlet pipe which communicates with the cock B. This cock controls the action of the engine, being so constructed that it acts as a reversing valve when the handle C is in its extreme positions and as a brake when in its middle position. With the handle in its middle position, the ports of the cylinders are in communication with the exhaust. Two passages are formed in the framing leading from the cock B to the ends of the cylinders, one being in communication with the supply pipe A, the other with the discharge pipe Q. These passages end as shown at E. The oscillation of the cylinders puts them alternately in communication with each of these passages, and thus the water is alternately admitted and exhausted.

In any ordinary rotative engine the length of stroke is invariable. Consequently the consumption of water depends simply on the speed of the engine, irrespective of the effort overcome. If the power of the engine must be varied without altering the number of rotations, then the stroke must be made variable. Messrs Hastie have contrived an exceedingly ingenious method of varying the stroke automatically, in proportion to the amount of work to be done (fig. 174). The crank pin I is carried in a slide H moving in a disk M. In this is a double cam K acting on two small steel rollers J, L attached to the slide H. If the cam rotates it moves the slide and increases or decreases the radius of the circle in which the crank pin I rotates. The disk M is keyed on a hollow shaft surrounding the driving shaft P, to which the cams are attached. The hollow shaft N has two snugs to which the chains RR are attached (fig. 175). The shaft P carries the spring case SS to which also are attached the other ends of the chains. When the engine is at rest the springs extend themselves, rotating the hollow shaft N and the frame M, so as to place the crank pin I at its nearest position to the axis of rotation. When a resistance has to be overcome, the shaft N rotates relatively to P, compressing the springs, till their resistance balances the pressure due to the resistance to the rotation of P. The engine then commences to work, the crank pin being in the position in which the turning effort just overcomes the resistance. If the resistance diminishes, the springs force out the chains and shorten the stroke of the plungers, and vice versa. The following experiments, on an engine of this kind working a hoist, show how the automatic arrangement adjusted the water used to the work done. The lift was 22 ft. and the water pressure in the cylinders 80 ℔ per sq. in.

Weight lifted, in ℔	Chain only	427	633	745	857	969	1081	1193
Water used, in gallons	71/2	10	14	16	17	20	21	22

§ 179. Accumulator Machinery.—It has already been pointed out that it is in some cases convenient to use a steam engine to create an artificial head of water, which is afterwards employed in driving water-pressure machinery. Where power is required intermittently, for short periods, at a number of different points, as, for instance, in moving the cranes, lock gates, &c., of a dockyard, a separate steam engine and boiler at each point is very inconvenient; nor can engines worked from a common boiler be used, because of the great loss of heat and the difficulties which arise out of condensation in the pipes. If a tank, into which water is continuously pumped, can be placed at a great elevation, the water can then be used in hydraulic machinery in a very convenient way. Each hydraulic machine is put in communication with the tank by a pipe, and on opening a valve it commences work, using a quantity of water directly proportional to the work done. No attendance is required when the machine is not working.

Fig. 176.

A site for such an elevated tank is, however, seldom available, and in place of it a beautiful arrangement termed an accumulator, invented by Lord Armstrong, is used. This consists of a tall vertical cylinder; into this works a solid ram through cup leathers or hemp packing, and the ram is loaded by fixed weights, so that the pressure in the cylinder is 700 ℔ or 800 ℔ per sq. in. In some cases the ram is fixed and the cylinder moves on it. The pumping engines which supply the energy that is stored in the accumulator should be a pair coupled at right angles, so as to start in any position. The engines pump into the accumulator cylinder till the ram is at the top of its stroke, when by a catch arrangement acting on the engine throttle valve the engines are stopped. If the accumulator ram descends, in consequence of water being taken to work machinery, the engines immediately recommence working. Pipes lead from the accumulator to each of the machines requiring to be driven, and do not require to be of large size, as the pressure is so great.

Fig. 176 shows a diagrammatic way the scheme of a system of accumulator machinery. A is the accumulator, with its ram carrying a cylindrical wrought-iron tank W, in which weights are placed to load the accumulator. At R is one of the pressure engines or jiggers, worked from the accumulator, discharging the water after use into the tank T. In this case the pressure engine is shown working a set of blocks, the fixed block being on the ram cylinder, the running block on the ram. The chain running over these blocks works a lift cage C, the speed of which is as many times greater than that of the ram as there are plies of chain on the block tackle. B is the balance weight of the cage.

Fig. 177.

In the use of accumulators on shipboard for working gun gear or steering gear, the accumulator ram is loaded by springs, or by steam pressure acting on a piston much larger than the ram.

R. H. Tweddell has used accumulators with a pressure of 2000 ℔ per sq. in. to work hydraulic riveting machinery.

The amount of energy stored in the accumulator, having a ram d in. in diameter, a stroke of S ft., and delivering at p ℔ pressure per sq. in., is

π/4pd ²S foot-pounds.

Thus, if the ram is 9 in., the stroke 20 ft., and the pressure 800 ℔ per sq. in., the work stored in the accumulator when the ram is at the top of the stroke is 1,017,600 foot-pounds, that is, enough to drive a machine requiring one horse power for about half an hour. As, however, the pumping engine replaces water as soon as it is drawn off, the working capacity of the accumulator is very much greater than this. Tweddell found that an accumulator charged at 1250 ℔ discharged at 1225 ℔ per sq. in. Hence the friction was equivalent to 121/2 ℔ per sq. in. and the efficiency 98%.

When a very great pressure is required a differential accumulator (fig. 177) is convenient. The ram is fixed and passes through both ends of the cylinder, but is of different diameters at the two ends, A and B. Hence if d₁, d₂ are the diameters of the ram in inches and p the required pressure in ℔ per sq. in., the load required is 1/4pπ(d₁² − d₂²). An accumulator of this kind used with riveting machines has d₁ = 51/2 in., d₂ = 43/4 in. The pressure is 2000 ℔ per sq. in. and the load 5.4 tons.

Fig. 178.

Sometimes an accumulator is loaded by water or steam pressure instead of by a dead weight. Fig. 178 shows the arrangement. A piston A is connected to a plunger B of much smaller area. Water pressure, say from town mains, is admitted below A, and the high pressure water is pumped into and discharged from the cylinder C in which B works. If r is the ratio of the areas of A and B, then, neglecting friction, the pressure in the upper cylinder is r times that under the piston A. With a variable rate of supply and demand from the upper cylinder, the piston A rises and falls, maintaining always a constant pressure in the upper cylinder.

Water Wheels.

§ 180. Overshot and High Breast Wheels.—When a water fall ranges between 10 and 70 ft. and the water supply is from 3 to 25 cub. ft. per second, it is possible to construct a bucket wheel on which the water acts chiefly by its weight. If the variation of the head-water level does not exceed 2 ft., an overshot wheel may be used (fig. 179). The water is then projected over the summit of the wheel, and falls in a parabolic path into the buckets. With greater variation of head-water level, a pitch-back or high breast wheel is better. The water falls over the top of a sliding sluice into the wheel, on the same side as the head race channel. By adjusting the height of the sluice, the requisite supply is given to the wheel in all positions of the head-water level.

The wheel consists of a cast-iron or wrought-iron axle C supporting the weight of the wheel. To this are attached two sets of arms A of wood or iron, which support circular segmental plates, B, termed shrouds. A cylindrical sole plate dd extends between the shrouds on the inner side. The buckets are formed by wood planks or curved wrought-iron plates extending from shroud to shroud, the back of the buckets being formed by the sole plate.

Fig. 179.

The efficiency may be taken at 0.75. Hence, if h.p. is the effective horse power, H the available fall, and Q the available water supply per second,

h.p. = 0.75 (GQH/550) = 0.085 QH.

If the peripheral velocity of the water wheel is too great, water is thrown out of the buckets before reaching the bottom of the fall. In practice, the circumferential velocity of water wheels of the kind now described is from 41/2 to 10 ft. per second, about 6 ft. being the usual velocity of good iron wheels not of very small size. In order that the water may enter the buckets easily, it must have a greater velocity than the wheel. Usually the velocity of the water at the point where it enters the wheel is from 9 to 12 ft. per second, and to produce this it must enter the wheel at a point 16 to 27 in. below the head-water level. Hence the diameter of an overshot wheel may be

D = H−11/3 to H−21/4 ft.

Overshot and high breast wheels work badly in backwater, and hence if the tail-water level varies, it is better to reduce the diameter of the wheel so that its greatest immersion in flood is not more than 1 ft. The depth d of the shrouds is about 10 to 16 in. The number of buckets may be about

N = πD/d.

Let v be the peripheral velocity of the wheel. Then the capacity of that portion of the wheel which passes the sluice in one second is

Q₁ = vb (Dd − d ²)/D
= v b d nearly,

b being the breadth of the wheel between the shrouds. If, however, this quantity of water were allowed to pass on to the wheel the buckets would begin to spill their contents almost at the top of the fall. To diminish the loss from spilling, it is not only necessary to give the buckets a suitable form, but to restrict the water supply to one-fourth or one-third of the gross bucket capacity. Let m be the value of this ratio; then, Q being the supply of water per second,

Q = mQ₁ = mbdv.

This gives the breadth of the wheel if the water supply is known. The form of the buckets should be determined thus. The outer element of the bucket should be in the direction of motion of the water entering relatively to the wheel, so that the water may enter without splashing or shock. The buckets should retain the water as long as possible, and the width of opening of the buckets should be 2 or 3 in. greater than the thickness of the sheet of water entering.

Fig. 180.

For a wooden bucket (fig. 180, A), take ab = distance between two buckets on periphery of wheel. Make ed = 1/2 eb and bc = 6/5 to 5/4 ab. Join cd. For an iron bucket (fig. 180, B), take ed = 1/3eb; bc = 6/5ab. Draw cO making an angle of 10° to 15° with the radius at c. On Oc take a centre giving a circular arc passing near d, and round the curve into the radial part of the bucket de.

There are two ways in which the power of a water wheel is given off to the machinery driven. In wooden wheels and wheels with rigid arms, a spur or bevil wheel keyed on the axle of the turbine will transmit the power to the shafting. It is obvious that the whole turning moment due to the weight of the water is then transmitted through the arms and axle of the water wheel. When the water wheel is an iron one, it usually has light iron suspension arms incapable of resisting the bending action due to the transmission of the turning effort to the axle. In that case spur segments are bolted to one of the shrouds, and the pinion to which the power is transmitted is placed so that the teeth in gear are, as nearly as may be, on the line of action of the resultant of the weight of the water in the loaded arc of the wheel.

The largest high breast wheels ever constructed were probably the four wheels, each 50 ft. in diameter, and of 125 h.p., erected by Sir W. Fairbairn in 1825 at Catrine in Ayrshire. These wheels are still working.

Fig. 181.

§ 181. Poncelet Water Wheel.—When the fall does not exceed 6 ft., the best water motor to adopt in many cases is the Poncelet undershot water wheel. In this the water acts very nearly in the same way as in a turbine, and the Poncelet wheel, although slightly less efficient than the best turbines, in normal conditions of working, is superior to most of them when working with a reduced supply of water. A general notion of the action of the water on a Poncelet wheel has already been given in § 159. Fig. 181 shows its construction. The water penned back between the side walls of the wheel pit is allowed to flow to the wheel under a movable sluice, at a velocity nearly equal to the velocity due to the whole fall. The water is guided down a slope of 1 in 10, or a curved race, and enters the wheel without shock. Gliding up the curved floats it comes to rest, falls back, and acquires at the point of discharge a backward velocity relative to the wheel nearly equal to the forward velocity of the wheel. Consequently it leaves the wheel deprived of nearly the whole of its original kinetic energy.

Taking the efficiency at 0.60, and putting H for the available fall, h.p. for the horse-power, and Q for the water supply per second,

h.p. = 0.068 QH.

The diameter D of the wheel may be taken arbitrarily. It should not be less than twice the fall and is more often four times the fall. For ordinary cases the smallest convenient diameter is 14 ft. with a straight, or 10 ft. with a curved, approach channel. The radial depth of bucket should be at least half the fall, and radius of curvature of buckets about half the radius of the wheel. The shrouds are usually of cast iron with flanges to receive the buckets. The buckets may be of iron 1/8 in. thick bolted to the flanges with 5/16 in. bolts.

Let H′ be the fall measured from the free surface of the head-water to the point F where the mean layer enters the wheel; then the velocity at which the water enters is v = √ (2gH′), and the best circumferential velocity of the wheel is V = 0.55v to 0.6v. The number of rotations of the wheel per second is N = V/πD. The thickness of the sheet of water entering the wheel is very important. The best thickness according to experiment is 8 to 10 in. The maximum thickness should not exceed 12 to 15 in., when there is a surplus water supply. Let e be the thickness of the sheet of water entering the wheel, and b its width; then

bev = Q; or b = Q/ev.

Grashof takes e = 1/6H, and then

b = 6Q/H √ (2gH).

Allowing for the contraction of the stream, the area of opening through the sluice may be 1.25 be to 1.3 be. The inside width of the wheel is made about 4 in. greater than b.

Several constructions have been given for the floats of Poncelet wheels. One of the simplest is that shown in figs. 181, 182.

Let OA (fig. 181) be the vertical radius of the wheel. Set off OB, OD making angles of 15° with OA. Then BD may be the length of the close breasting fitted to the wheel. Draw the bottom of the head face BC at a slope of 1 in 10. Parallel to this, at distances 1/2e and e, draw EF and GH. Then EF is the mean layer and GH the surface layer entering the wheel. Join OF, and make OFK = 23°. Take FK = 0.5 to 0.7 H. Then K is the centre from which the bucket curve is struck and KF is the radius. The depth of the shrouds must be sufficient to prevent the water from rising over the top of the float. It is 1/2H to 2/3H. The number of buckets is not very important. They are usually 1 ft. apart on the circumference of the wheel.

The efficiency of a Poncelet wheel has been found in experiments to reach 0.68. It is better to take it at 0.6 in estimating the power of the wheel, so as to allow some margin.

Fig. 182.

In fig. 182 v_i is the initial and v_o the final velocity of the water, v_r parallel to the vane the relative velocity of the water and wheel, and V the velocity of the wheel.

Turbines.

§ 182. The name turbine was originally given in France to any water motor which revolved in a horizontal plane, the axis being vertical. The rapid development of this class of motors dates from 1827, when a prize was offered by the Société d’Encouragement for a motor of this kind, which should be an improvement on certain wheels then in use. The prize was ultimately awarded to Benoît Fourneyron (1802–1867), whose turbine, but little modified, is still constructed.

Classification of Turbines.—In some turbines the whole available energy of the water is converted into kinetic energy before the water acts on the moving part of the turbine. Such turbines are termed Impulse or Action Turbines, and they are distinguished by this that the wheel passages are never entirely filled by the water. To ensure this condition they must be placed a little above the tail water and discharge into free air. Turbines in which part only of the available energy is converted into kinetic energy before the water enters the wheel are termed Pressure or Reaction Turbines. In these there is a pressure which in some cases amounts to half the head in the clearance space between the guide vanes and wheel vanes. The velocity with which the water enters the wheel is due to the difference between the pressure due to the head and the pressure in the clearance space. In pressure turbines the wheel passages must be continuously filled with water for good efficiency, and the wheel may be and generally is placed below the tail water level.

Some turbines are designed to act normally as impulse turbines discharging above the tail water level. But the passages are so designed that they are just filled by the water. If the tail water rises and drowns the turbine they become pressure turbines with a small clearance pressure, but the efficiency is not much affected. Such turbines are termed Limit turbines.

Next there is a difference of constructive arrangement of turbines, which does not very essentially alter the mode of action of the water. In axial flow or so-called parallel flow turbines, the water enters and leaves the turbine in a direction parallel to the axis of rotation, and the paths of the molecules lie on cylindrical surfaces concentric with that axis. In radial outward and inward flow turbines, the water enters and leaves the turbine in directions normal to the axis of rotation, and the paths of the molecules lie exactly or nearly in planes normal to the axis of rotation. In outward flow turbines the general direction of flow is away from the axis, and in inward flow turbines towards the axis. There are also mixed flow turbines in which the water enters normally and is discharged parallel to the axis of rotation.

Another difference of construction is this, that the water may be admitted equally to every part of the circumference of the turbine wheel or to a portion of the circumference only. In the former case, the condition of the wheel passages is always the same; they receive water equally in all positions during rotation. In the latter case, they receive water during a part of the rotation only. The former may be termed turbines with complete admission, the latter turbines with partial admission. A reaction turbine should always have complete admission. An impulse turbine may have complete or partial admission.

When two turbine wheels similarly constructed are placed on the same axis, in order to balance the pressures and diminish journal friction, the arrangement may be termed a twin turbine.

If the water, having acted on one turbine wheel, is then passed through a second on the same axis, the arrangement may be termed a compound turbine. The object of such an arrangement would be to diminish the speed of rotation.

Many forms of reaction turbine may be placed at any height not exceeding 30 ft. above the tail water. They then discharge into an air-tight suction pipe. The weight of the column of water in this pipe balances part of the atmospheric pressure, and the difference of pressure, producing the flow through the turbine, is the same as if the turbine were placed at the bottom of the fall.

I. Impulse Turbines.	II. Reaction Turbines.
(Wheel passages not filled, and	(Wheel passages filled, discharging
discharging above the tail water.)	above or below the tail water  or into a suction-pipe.
(a) Complete admission. (Rare.)	Always with complete admission.
(b) Partial admission. (Usual.)
$\scriptstyle {\underbrace {\quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad } }$
Axial flow, outward flow, inward flow, or mixed flow.
$\scriptstyle {\underbrace {\quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad \quad } }$
Simple turbines; twin turbines; compound turbines.

Fig. 183.

§ 183. The Simple Reaction Wheel.—It has been shown, in § 162, that, when water issues from a vessel, there is a reaction on the vessel tending to cause motion in a direction opposite to that of the jet. This principle was applied in a rotating water motor at a very early period, and the Scotch turbine, at one time much used, differs in no essential respect from the older form of reaction wheel.

The old reaction wheel consisted of a vertical pipe balanced on a vertical axis, and supplied with water (fig. 183). From the bottom of the vertical pipe two or more hollow horizontal arms extended, at the ends of which were orifices from which the water was discharged. The reaction of the jets caused the rotation of the machine.

Let H be the available fall measured from the level of the water in the vertical pipe to the centres of the orifices, r the radius from the axis of rotation to the centres of the orifices, v the velocity of discharge through the jets, α the angular velocity of the machine. When the machine is at rest the water issues from the orifices with the velocity √ (2gH) (friction being neglected). But when the machine rotates the water in the arms rotates also, and is in the condition of a forced vortex, all the particles having the same angular velocity. Consequently the pressure in the arms at the orifices is H + α²r ²/2g ft. of water, and the velocity of discharge through the orifices is v = √ (2gH + α²r ²). If the total area of the orifices is ω, the quantity discharged from the wheel per second is

Q = ωv = ω √ (2gH + α²r ²).

While the water passes through the orifices with the velocity v, the orifices are moving in the opposite direction with the velocity αr. The absolute velocity of the water is therefore

v − αr = √ (2gH + α²r ²) − αr.

The momentum generated per second is (GQ/g)(v − αr), which is numerically equal to the force driving the motor at the radius r. The work done by the water in rotating the wheel is therefore

(GQ/g) (v − αr) αr foot-pounds per sec.

The work expended by the water fall is GQH foot-pounds per second. Consequently the efficiency of the motor is

η =	(v − αr) αr	=	{√ (2gH + α2r 2) − αr } αr	.
	gH		gH

Let

√ (2gH + α2r 2) = αr +	gH	−	g2H2	...
	αr		2α3r 3

then

η = 1 − gH / 2αr + . . .

which increases towards the limit 1 as αr increases towards infinity. Neglecting friction, therefore, the maximum efficiency is reached when the wheel has an infinitely great velocity of rotation. But this condition is impracticable to realize, and even, at practicable but high velocities of rotation, the friction would considerably reduce the efficiency. Experiment seems to show that the best efficiency is reached when αr = √ (2gH). Then the efficiency apart from friction is

η = {√ (2α²r ²) − αr } αr / gH

= 0.414 α²r ² / gH = 0.828,

about 17% of the energy of the fall being carried away by the water discharged. The actual efficiency realized appears to be about 60%, so that about 21% of the energy of the fall is lost in friction, in addition to the energy carried away by the water.

§ 184. General Statement of Hydrodynamical Principles necessary for the Theory of Turbines.

(a) When water flows through any pipe-shaped passage, such as the passage between the vanes of a turbine wheel, the relation between the changes of pressure and velocity is given by Bernoulli’s theorem (§ 29). Suppose that, at a section A of such a passage, h₁ is the pressure measured in feet of water, v₁ the velocity, and z₁ the elevation above any horizontal datum plane, and that at a section B the same quantities are denoted by h₂, v₂, z₂. Then

h1 − h2 = (v22 − v12) / 2g + z2 − z1.

(1)

If the flow is horizontal, z₂ = z₁; and

h1 − h2 = (v22 − v12) / 2g.

(1a)

(b) When there is an abrupt change of section of the passage, or an abrupt change of section of the stream due to a contraction, then, in applying Bernoulli’s equation allowance must be made for the loss of head in shock (§ 36). Let v₁, v₂ be the velocities before and after the abrupt change, then a stream of velocity v₁ impinges on a stream at a velocity v₂, and the relative velocity is v₁ − v₂. The head lost is (v₁ − v₂)²/2g. Then equation (1a) becomes

h₂ − h₁ = (v₁² − v₂²) / 2g − (v₁ − v₂)² / 2g = v₂ (v₁ − v₂) / g

(2)

Fig. 184.

To diminish as much as possible the loss of energy from irregular eddying motions, the change of section in the turbine passages must be very gradual, and the curvature without discontinuity.

(c) Equality of Angular Impulse and Change of Angular Momentum.—Suppose that a couple, the moment of which is M, acts on a body of weight W for t seconds, during which it moves from A₁ to A₂ (fig. 184). Let v₁ be the velocity of the body at A₁, v₂ its velocity at A₂, and let p₁, p₂ be the perpendiculars from C on v₁ and v₂. Then Mt is termed the angular impulse of the couple, and the quantity

(W/g) (v₂p₂ − v₁p₁)

is the change of angular momentum relatively to C. Then, from the equality of angular impulse and change of angular momentum

Mt = (W/g) (v₂p₂ − v₁p₁),

or, if the change of momentum is estimated for one second,

M = (W/g) (v₂p₂ − v₁p₁).

Let r₁, r₂ be the radii drawn from C to A₁, A₂, and let w₁, w₂ be the components of v₁, v₂, perpendicular to these radii, making angles β and α with v₁, v₂. Then

v₁ = w₁ sec β; v₂ = w₂ sec α

p₁ = r₁ cos β; p₂ = r₂ cos α,

∴ M = (W/g) (w₂r₂ − w₁r₁),

(3)

where the moment of the couple is expressed in terms of the radii drawn to the positions of the body at the beginning and end of a second, and the tangential components of its velocity at those points.

Now the water flowing through a turbine enters at the admission surface and leaves at the discharge surface of the wheel, with its angular momentum relatively to the axis of the wheel changed. It therefore exerts a couple −M tending to rotate the wheel, equal and opposite to the couple M which the wheel exerts on the water. Let Q cub. ft. enter and leave the wheel per second, and let w₁, w₂ be the tangential components of the velocity of the water at the receiving and discharging surfaces of the wheel, r₁, r₂ the radii of those surfaces. By the principle above,

−M = (GQ/g) (w₂r₂ − w₁r₁).

(4)

If α is the angular velocity of the wheel, the work done by the water on the wheel is

T = Ma = (GQ/g) (w₁r₁ − w₂r₂) α foot-pounds per second.

(5)

§ 185. Total and Available Fall.—Let H_t be the total difference of level from the head-water to the tail-water surface. Of this total head a portion is expended in overcoming the resistances of the head race, tail race, supply pipe, or other channel conveying the water. Let ɧ_p be that loss of head, which varies with the local conditions in which the turbine is placed. Then

H = H_t − ɧ_p

is the available head for working the turbine, and on this the calculations for the turbine should be based. In some cases it is necessary to place the turbine above the tail-water level, and there is then a fall ɧ from the centre of the outlet surface of the turbine to the tail-water level which is wasted, but which is properly one of the losses belonging to the turbine itself. In that case the velocities of the water in the turbine should be calculated for a head H − ɧ, but the efficiency of the turbine for the head H.

§ 186. Gross Efficiency and Hydraulic Efficiency of a Turbine.—Let Td be the useful work done by the turbine, in foot-pounds per second, T_t the work expended in friction of the turbine shaft, gearing, &c., a quantity which varies with the local conditions in which the turbine is placed. Then the effective work done by the water in the turbine is

T = Td + T_t.

The gross efficiency of the whole arrangement of turbine, races, and transmissive machinery is

η_t = Td / CQH_t.

(6)

And the hydraulic efficiency of the turbine alone is

η = T / GQH.

(7)

It is this last efficiency only with which the theory of turbines is concerned.

From equations (5) and (7) we get

ηGQH = (GQ/g) (w₁r₁ − w₂r₂) α;

η = (w₁r₁ − w₂r₂) α/gH.

(8)

This is the fundamental equation in the theory of turbines. In general,^[7] w₁ and w₂, the tangential components of the water’s motion on entering and leaving the wheel, are completely independent. That the efficiency may be as great as possible, it is obviously necessary that w₂ = 0. In that case

η = w₁r₁α / gH.

(9)

αr₁ is the circumferential velocity of the wheel at the inlet surface. Calling this V₁, the equation becomes

η = w₁V₁ / gH.

(9a)

This remarkably simple equation is the fundamental equation in the theory of turbines. It was first given by Reiche (Turbinenbaues, 1877).

Fig. 185.
 

Fig. 186.
 

Fig. 187.
 

Fig. 188.

Fig. 189.

§ 187. General Description of a Reaction Turbine.—Professor James Thomson’s inward flow or vortex turbine has been selected as the type of reaction turbines. It is one of the best in normal conditions of working, and the mode of regulation introduced is decidedly superior to that in most reaction turbines. Figs. 185 and 186 are external views of the turbine case; figs. 187 and 188 are the corresponding sections; fig. 189 is the turbine wheel. The example chosen for illustration has suction pipes, which permit the turbine to be placed above the tail-water level. The water enters the turbine by cast-iron supply pipes at A, and is discharged through two suction pipes S, S. The water on entering the case distributes itself through a rectangular supply chamber SC, from which it finds its way equally to the four guide-blade passages G, G, G, G. In these passages it acquires a velocity about equal to that due to half the fall, and is directed into the wheel at an angle of about 10° or 12° with the tangent to its circumference. The wheel W receives the water in equal proportions from each guide-blade passage. It consists of a centre plate p (fig. 189) keyed on the shaft aa, which passes through stuffing boxes on the suction pipes. On each side of the centre plate are the curved wheel vanes, on which the pressure of the water acts, and the vanes are bounded on each side by dished or conical cover plates c, c. Joint-rings j, j on the cover plates make a sufficiently water-tight joint with the casing, to prevent leakage from the guide-blade chamber into the suction pipes. The pressure near the joint rings is not very great, probably not one-fourth the total head. The wheel vanes receive the water without shock, and deliver it into central spaces, from which it flows on either side to the suction pipes. The mode of regulating the power of the turbine is very simple. The guide-blades are pivoted to the case at their inner ends, and they are connected by a link-work, so that they all open and close simultaneously and equally. In this way the area of opening through the guide-blades is altered without materially altering the angle or the other conditions of the delivery into the wheel. The guide-blade gear may be variously arranged. In this example four spindles, passing through the case, are linked to the guide-blades inside the case, and connected together by the links l, l, l on the outside of the case. A worm wheel on one of the spindles is rotated by a worm d, the motion being thus slow enough to adjust the guide-blades very exactly. These turbines are made by Messrs Gilkes & Co. of Kendal.

Fig. 190.

Fig. 190 shows another arrangement of a similar turbine, with some adjuncts not shown in the other drawings. In this case the turbine rotates horizontally, and the turbine case is placed entirely below the tail water. The water is supplied to the turbine by a vertical pipe, over which is a wooden pentrough, containing a strainer, which prevents sticks and other solid bodies getting into the turbine. The turbine rests on three foundation stones, and, the pivot for the vertical shaft being under water, there is a screw and lever arrangement for adjusting it as it wears. The vertical shaft gives motion to the machinery driven by a pair of bevel wheels. On the right are the worm and wheel for working the guide-blade gear.

Fig. 191.

§ 188. Hydraulic Power at Niagara.—The largest development of hydraulic power is that at Niagara. The Niagara Falls Power Company have constructed two power houses on the United States side, the first with 10 turbines of 5000 h.p. each, and the second with 10 turbines of 5500 h.p. The effective fall is 136 to 140 ft. In the first power house the turbines are twin outward flow reaction turbines with vertical shafts running at 250 revs. per minute and driving the dynamos direct. In the second power house the turbines are inward flow turbines with draft tubes or suction pipes. Fig. 191 shows a section of one of these turbines. There is a balancing piston keyed on the shaft, to the under side of which the pressure due to the fall is admitted, so that the weight of turbine, vertical shaft and part of the dynamo is water borne. About 70,000 h.p. is daily distributed electrically from these two power houses. The Canadian Niagara Power Company are erecting a power house to contain eleven units of 10,250 h.p. each, the turbines being twin inward flow reaction turbines. The Electrical Development Company of Ontario are erecting a power house to contain 11 units of 12,500 h.p. each. The Ontario Power Company are carrying out another scheme for developing 200,000 h.p. by twin inward flow turbines of 12,000 h.p. each. Lastly the Niagara Falls Power and Manufacturing Company on the United States side have a station giving 35,000 h.p. and are constructing another to furnish 100,000 h.p. The mean flow of the Niagara river is about 222,000 cub. ft. per second with a fall of 160 ft. The works in progress if completed will utilize 650,000 h.p. and require 48,000 cub. ft. per second or 211/2% of the mean flow of the river (Unwin, “The Niagara Falls Power Stations,” Proc. Inst. Mech. Eng., 1906).

Fig. 192.

§ 189. Different Forms of Turbine Wheel.—The wheel of a turbine or part of the machine on which the water acts is an annular space, furnished with curved vanes dividing it into passages exactly or roughly rectangular in cross section. For radial flow turbines the wheel may have the form A or B, fig. 192, A being most usual with inward, and B with outward flow turbines. In A the wheel vanes are fixed on each side of a centre plate keyed on the turbine shaft. The vanes are limited by slightly-coned annular cover plates. In B the vanes are fixed on one side of a disk, keyed on the shaft, and limited by a cover plate parallel to the disk. Parallel flow or axial flow turbines have the wheel as in C. The vanes are limited by two concentric cylinders.

Theory of Reaction Turbines.

Fig. 193.

§ 190. Velocity of Whirl and Velocity of Flow.—Let acb (fig. 193) be the path of the particles of water in a turbine wheel. That path will be in a plane normal to the axis of rotation in radial flow turbines, and on a cylindrical surface in axial flow turbines. At any point c of the path the water will have some velocity v, in the direction of a tangent to the path. That velocity may be resolved into two components, a whirling velocity w in the direction of the wheel’s rotation at the point c, and a component u at right angles to this, radial in radial flow, and parallel to the axis in axial flow turbines. This second component is termed the velocity of flow. Let v_o, w_o, u_o be the velocity of the water, the whirling velocity and velocity of flow at the outlet surface of the wheel, and v_i, w_i, u_i the same quantities at the inlet surface of the wheel. Let α and β be the angles which the water’s direction of motion makes with the direction of motion of the wheel at those surfaces. Then

wo = vo cos β; uo = vo sin β wi = vi cos α; ui = vi sin α.

${\Big \}}$

(10)

The velocities of flow are easily ascertained independently from the dimensions of the wheel. The velocities of flow at the inlet and outlet surfaces of the wheel are normal to those surfaces. Let Ω_o, Ω_i be the areas of the outlet and inlet surfaces of the wheel, and Q the volume of water passing through the wheel per second; then

v₀ = Q/Ω_o; v_i = Q/Ω_i.

(11)

Using the notation in fig. 191, we have, for an inward flow turbine (neglecting the space occupied by the vanes),

Ω_o = 2πr₀d₀; Ω_i = 2πr_id_i.

(12a)

Similarly, for an outward flow turbine,

Ω_o = 2πr_od; Ω_i = 2πr_id;

(12b)

and, for an axial flow turbine,

Ω_o = Ω_i = π (r₂² − r₁²).

(12c)

Fig. 194.

Relative and Common Velocity of the Water and Wheel.—There is another way of resolving the velocity of the water. Let V be the velocity of the wheel at the point c, fig. 194. Then the velocity of the water may be resolved into a component V, which the water has in common with the wheel, and a component v_r, which is the velocity of the water relatively to the wheel.

Velocity of Flow.—It is obvious that the frictional losses of head in the wheel passages will increase as the velocity of flow is greater, that is, the smaller the wheel is made. But if the wheel works under water, the skin friction of the wheel cover increases as the diameter of the wheel is made greater, and in any case the weight of the wheel and consequently the journal friction increase as the wheel is made larger. It is therefore desirable to choose, for the velocity of flow, as large a value as is consistent with the condition that the frictional losses in the wheel passages are a small fraction of the total head.

The values most commonly assumed in practice are these:—

In axial flow turbines,	uo = ui = 0.15 to 0.2 √(2gH);
In outward flow turbines,	ui = 0.25 √2g (H − ɧ),
	uo = 0.21 to 0.17 √2g (H − ɧ);
In inward flow turbines,	uo = ui = 0.125 √(2gH).

§ 191. Speed of the Wheel.—The best speed of the wheel depends partly on the frictional losses, which the ordinary theory of turbines disregards. It is best, therefore, to assume for V_o and V_i values which experiment has shown to be most advantageous.

In axial flow turbines, the circumferential velocities at the mean radius of the wheel may be taken

V_o = V_i = 0.6 √2gH to 0.66 √2gH.

In a radial outward flow turbine,

V_i = 0.56 √2g(H − ɧ)

V_o = V_ir_o / r_i,

where r_o, r_i are the radii of the outlet and inlet surfaces.

In a radial inward flow turbine,

V_i = 0.66 √2gH,

V_o = V_i r_o / r_i.

If the wheel were stationary and the water flowed through it, the water would follow paths parallel to the wheel vane curves, at least when the vanes were so close that irregular motion was prevented. Similarly, when the wheel is in motion, the water follows paths relatively to the wheel, which are curves parallel to the wheel vanes. Hence the relative component, v_r, of the water’s motion at c is tangential to a wheel vane curve drawn through the point c. Let v_o, V_o, v_ro be the velocity of the water and its common and relative components at the outlet surface of the wheel, and v_i, V_i, v_ri be the same quantities at the inlet surface; and let θ and φ be the angles the wheel vanes make with the inlet and outlet surfaces; then

vo2 = √ (vro2 + Vo2 − 2Vovro cos φ) vi = √ (vr i2 + Vo2 − 2Vivr i  cos θ)

${\bigg \}}$,

(13)

equations which may be used to determine φ and θ.

Fig. 195.

§ 192. Condition determining the Angle of the Vanes at the Outlet Surface of the Wheel.—It has been shown that, when the water leaves the wheel, it should have no tangential velocity, if the efficiency is to be as great as possible; that is, w_o = 0. Hence, from (10), cos β = 0, β = 90°, U_o = V_o, and the direction of the water’s motion is normal to the outlet surface of the wheel, radial in radial flow, and axial in axial flow turbines.

Drawing v_o or u_o radial or axial as the case may be, and V_o tangential to the direction of motion, v_ro can be found by the parallelogram of velocities. From fig. 195,

tan φ = v_o / V_o = u_o / V_o;

(14)

but φ is the angle which the wheel vane makes with the outlet surface of the wheel, which is thus determined when the velocity of flow u_o and velocity of the wheel V_o are known. When φ is thus determined,

v_ro = U_o cosec φ = V_o √ (1 + u_o² / V_o²).

(14a)

Correction of the Angle φ to allow for Thickness of Vanes.—In determining φ, it is most convenient to calculate its value approximately at first, from a value of u_o obtained by neglecting the thickness of the vanes. As, however, this angle is the most important angle in the turbine, the value should be afterwards corrected to allow for the vane thickness.

Let

φ′ = tan⁻¹ (u_o / V_o) = tan⁻¹ (Q / Ω_oV_o)

be the first or approximate value of φ, and let t be the thickness, and n the number of wheel vanes which reach the outlet surface of the wheel. As the vanes cut the outlet surface approximately at the angle φ′, their width measured on that surface is t cosec φ′. Hence the space occupied by the vanes on the outlet surface is

For A, fig. 192, ntdo cosec φ B, fig. 192, ntd cosec φ C, fig. 192, nt(r2 − r1) cosec φ.

${\bigg \}}$.

(15)

Call this area occupied by the vanes ω. Then the true value of the clear discharging outlet of the wheel is Ω_o − ω, and the true value of u_o is Q/(Ω_o − ω). The corrected value of the angle of the vanes will be

φ = tan [Q / V_o (Ω_o − ω) ].

(16)

§ 193. Head producing Velocity with which the Water enters the Wheel.—Consider the variation of pressure in a wheel passage, which satisfies the condition that the sections change so gradually that there is no loss of head in shock. When the flow is in a horizontal plane, there is no work done by gravity on the water passing through the wheel. In the case of an axial flow turbine, in which the flow is vertical, the fall d between the inlet and outlet surfaces should be taken into account.

Let V_i, V_o be the velocities of the wheel at the inlet and
outlet surfaces,
    v_i, v_o the velocities of the water,
    u_i, u_o the velocities of flow,
    v_ri, v_ro the relative velocities,
    h_i, h_o the pressures, measured in feet of water,
    r_i, r_o the radii of the wheel,
      α the angular velocity of the wheel.

At any point in the path of a portion of water, at radius r, the velocity v of the water may be resolved into a component V = αr equal to the velocity at that point of the wheel, and a relative component v_r. Hence the motion of the water may be considered to consist of two parts:—(a) a motion identical with that in a forced vortex of constant angular velocity α; (b) a flow along curves parallel to the wheel vane curves. Taking the latter first, and using Bernoulli’s theorem, the change of pressure due to flow through the wheel passages is given by the equation

h′_i + v_ri² / 2g = h′_o + v_ro² / 2g;
h′_i − h′_o = (v_ro² − v_ri²) / 2g.

The variation of pressure due to rotation in a forced vortex is

h″_i − h″_o = (V_i² − V_o²) / 2g.

Consequently the whole difference of pressure at the inlet and outlet surfaces of the wheel is

h_i − h_o = h′_i + h″_i − h′_o − h″_o
= (V_i² − V_o²) / 2g + (v_ro² − v_ri²) / 2g.

(17)

Case 1. Axial Flow Turbines.—V_i = V_o; and the first term on the right, in equation 17, disappears. Adding, however, the work of gravity due to a fall of d ft. in passing through the wheel,

h_i − h_o = (v_ro² − v_ri²) / 2g − d.

(17a)

Case 2. Outward Flow Turbines.—The inlet radius is less than the outlet radius, and (V_i² − V_o²)/2g is negative. The centrifugal head diminishes the pressure at the inlet surface, and increases the velocity with which the water enters the wheel. This somewhat increases the frictional loss of head. Further, if the wheel varies in velocity from variations in the useful work done, the quantity (V_i² − V_o²)/2g increases when the turbine speed increases, and vice versa. Consequently the flow into the turbine increases when the speed increases, and diminishes when the speed diminishes, and this again augments the variation of speed. The action of the centrifugal head in an outward flow turbine is therefore prejudicial to steadiness of motion. For this reason r_o : r_i is made small, generally about 5 : 4. Even then a governor is sometimes required to regulate the speed of the turbine.

Case 3. Inward Flow Turbines.—The inlet radius is greater than the outlet radius, and the centrifugal head diminishes the velocity of flow into the turbine. This tends to diminish the frictional losses, but it has a more important influence in securing steadiness of motion. Any increase of speed diminishes the flow into the turbine, and vice versa. Hence the variation of speed is less than the variation of resistance overcome. In the so-called centre vent wheels in America, the ratio r_i : r_o is about 5 : 4, and then the influence of the centrifugal head is not very important. Professor James Thomson first pointed out the advantage of a much greater difference of radii. By making r_i : r_o = 2 : 1, the centrifugal head balances about half the head in the supply chamber. Then the velocity through the guide-blades does not exceed the velocity due to half the fall, and the action of the centrifugal head in securing steadiness of speed is considerable.

Since the total head producing flow through the turbine is H − ɧ, of this h_i − h_o is expended in overcoming the pressure in the wheel, the velocity of flow into the wheel is

v_i = c_v √ {2g (H − ɧ − (V_i² − V_o² / 2g + (v_ro² − v_ri²) / 2g) ],

(18)

where c_v may be taken 0.96.

From (14a),

v_ro = V_o √ (1 + u_o² / V_o²).

It will be shown immediately that

v_ri = u_i cosec θ;

or, as this is only a small term, and θ is on the average 90°, we may take, for the present purpose, v_ri = u_i nearly.

Inserting these values, and remembering that for an axial flow turbine V_i = V_o, ɧ = 0, and the fall d in the wheel is to be added,

vi = cv ${\sqrt {\Big \{}}$ 2g ${\Big (}$ H −	Vi2	${\Big (}$ 1 +	uo2	${\Big )}$ +	ui2	− d ${\Big )}{\Big \}}$.
	2g		Vo2		2g

For an outward flow turbine,

vi = cv ${\sqrt {\Big [}}$ 2g { H − ɧ −	Vi2	( 1 +	uo2	${\Big )}$ +	ui2	${\Big \}}{\Big ]}$.
	2g		Vi2		2g

For an inward flow turbine,

vi = cv ${\sqrt {\Big [}}$ 2g ${\Big \{}$ H −	Vi2	${\Big (}$ 1 +	uo2	${\Big )}$ +	ui2	${\Big \}}{\Big ]}$.
	2g		Vi2		2g

§ 194. Angle which the Guide-Blades make with the Circumference of the Wheel.—At the moment the water enters the wheel, the radial component of the velocity is u_i, and the velocity is v_i. Hence, if γ is the angle between the guide-blades and a tangent to the wheel

γ = sin⁻¹ (u_i/v_i).

Fig. 196.

This angle can, if necessary, be corrected to allow for the thickness of the guide-blades.

§ 195. Condition determining the Angle of the Vanes at the Inlet Surface of the Wheel.—The single condition necessary to be satisfied at the inlet surface of the wheel is that the water should enter the wheel without shock. This condition is satisfied if the direction of relative motion of the water and wheel is parallel to the first element of the wheel vanes.

Let A (fig. 196) be a point on the inlet surface of the wheel, and let v_i represent in magnitude and direction the velocity of the water entering the wheel, and V_i the velocity of the wheel. Completing the parallelogram, v_ri is the direction of relative motion. Hence the angle between v_ri and V_i is the angle θ which the vanes should make with the inlet surface of the wheel.

§ 196. Example of the Method of designing a Turbine. Professor James Thomson’s Inward Flow Turbine.—

Let H = the available fall after deducting loss of head in pipes and channels from the gross fall;
   Q = the supply of water in cubic feet per second; and
  η = the efficiency of the turbine.

The work done per second is ηGQH, and the horse-power of the turbine is h.p. = ηGQH/550. If η is taken at 0.75, an allowance will be made for the frictional losses in the turbine, the leakage and the friction of the turbine shaft. Then h.p. = 0.085QH.

The velocity of flow through the turbine (uncorrected for the space occupied by the vanes and guide-blades) may be taken

u_i = u_i = 0.125 √2gH,

in which case about 1/64th of the energy of the fall is carried away by the water discharged.

The areas of the outlet and inlet surface of the wheel are then

2πr_od_o = 2πr_id_i = Q / 0.125 √ (2gH).

If we take r_o, so that the axial velocity of discharge from the central orifices of the wheel is equal to u_o, we get

r_o = 0.3984 √ (Q/√H),
d_o = r_o.

If, to obtain considerable steadying action of the centrifugal head, r_i = 2r_o, then d_i = 1/2d_o.

Speed of the Wheel.—Let V_i = 0.66 √2gH, or the speed due to half the fall nearly. Then the number of rotations of the turbine per second is

N = V_i / 2πr_i = 1.0579 √ (H √ H/Q);

also

V_o = V_ir_o / r_i = 0.33 √2gH.

Angle of Vanes with Outlet Surface.

Tan φ = u_o / V_o = 0.125 / 0.33 = .3788;
φ = 21º nearly.

If this value is revised for the vane thickness it will ordinarily become about 25º.

Velocity with which the Water enters the Wheel.—The head producing the velocity is

 H − (V_i² / 2g) (1 + u_o² / V_i²) + u_i² / 2g
= H {1 − .4356 (1 + 0.0358) + .0156}
= 0.5646H.

Then the velocity is

V_i = .96 √2g (.5646H) = 0.721 √2gH.

Angle of Guide-Blades.

Sin γ = u_i / v_i = 0.125 / 0.721 = 0.173;

γ = 10° nearly.;

Tangential Velocity of Water entering Wheel.

w_i = v_i cos γ = 0.7101 √2gH.

Angle of Vanes at Inlet Surface.

Cot θ = (w_i − V_i) / u_i = (.7101 − .66) / .125 = .4008;
θ = 68° nearly.

Hydraulic Efficiency of Wheel.

η = w_iV_i / gH = .7101 × .66 × 2
 = 0.9373.

This, however, neglects the friction of wheel covers and leakage. The efficiency from experiment has been found to be 0.75 to 0.80.

Impulse and Partial Admission Turbines.

§ 197. The principal defect of most turbines with complete admission is the imperfection of the arrangements for working with less than the normal supply. With many forms of reaction turbine the efficiency is considerably reduced when the regulating sluices are partially closed, but it is exactly when the supply of water is deficient that it is most important to get out of it the greatest possible amount of work. The imperfection of the regulating arrangements is therefore, from the practical point of view, a serious defect. All turbine makers have sought by various methods to improve the regulating mechanism. B. Fourneyron, by dividing his wheel by horizontal diaphragms, virtually obtained three or more separate radial flow turbines, which could be successively set in action at their full power, but the arrangement is not altogether successful, because of the spreading of the water in the space between the wheel and guide-blades. Fontaine similarly employed two concentric axial flow turbines formed in the same casing. One was worked at full power, the other regulated. By this arrangement the loss of efficiency due to the action of the regulating sluice affected only half the water power. Many makers have adopted the expedient of erecting two or three separate turbines on the same waterfall. Then one or more could be put out of action and the others worked at full power. All these methods are rather palliatives than remedies. The movable guide-blades of Professor James Thomson meet the difficulty directly, but they are not applicable to every form of turbine.

Fig. 197.

C. Callon, in 1840, patented an arrangement of sluices for axial or outward flow turbines, which were to be closed successively as the water supply diminished. By preference the sluices were closed by pairs, two diametrically opposite sluices forming a pair. The water was thus admitted to opposite but equal arcs of the wheel, and the forces driving the turbine were symmetrically placed. As soon as this arrangement was adopted, a modification of the mode of action of the water in the turbine became necessary. If the turbine wheel passages remain full of water during the whole rotation, the water contained in each passage must be put into motion each time it passes an open portion of the sluice, and stopped each time it passes a closed portion of the sluice. It is thus put into motion and stopped twice in each rotation. This gives rise to violent eddying motions and great loss of energy in shock. To prevent this, the turbine wheel with partial admission must be placed above the tail water, and the wheel passages be allowed to clear themselves of water, while passing from one open portion of the sluices to the next.

But if the wheel passages are free of water when they arrive at the open guide passages, then there can be no pressure other than atmospheric pressure in the clearance space between guides and wheel. The water must issue from the sluices with the whole velocity due to the head; received on the curved vanes of the wheel, the jets must be gradually deviated and discharged with a small final velocity only, precisely in the same way as when a single jet strikes a curved vane in the free air. Turbines of this kind are therefore termed turbines of free deviation. There is no variation of pressure in the jet during the whole time of its action on the wheel, and the whole energy of the jet is imparted to the wheel, simply by the impulse due to its gradual change of momentum. It is clear that the water may be admitted in exactly the same way to any fraction of the circumference at pleasure, without altering the efficiency of the wheel. The diameter of the wheel may be made as large as convenient, and the water admitted to a small fraction of the circumference only. Then the number of revolutions is independent of the water velocity, and may be kept down to a manageable value.

Fig. 198.	Fig. 199.

§ 198. General Description of an Impulse Turbine or Turbine with Free Deviation.—Fig. 197 shows a general sectional elevation of a Girard turbine, in which the flow is axial. The water, admitted above a horizontal floor, passes down through the annular wheel containing the guide-blades G, G, and thence into the revolving wheel WW. The revolving wheel is fixed to a hollow shaft suspended from the pivot p. The solid internal shaft ss is merely a fixed column supporting the pivot. The advantage of this is that the pivot is accessible for lubrication and adjustment. B is the mortise bevel wheel by which the power of the turbine is given off. The sluices are worked by the hand wheel h, which raises them successively, in a way to be described presently. d, d are the sluice rods. Figs. 198, 199 show the sectional form of the guide-blade chamber and wheel and the curves of the wheel vanes and guide-blades, when drawn on a plane development of the cylindrical section of the wheel; a, a, a are the sluices for cutting off the water; b, b, b are apertures by which the entrance or exit of air is facilitated as the buckets empty and fill. Figs. 200, 201 show the guide-blade gear. a, a, a are the sluice rods as before. At the top of each sluice rod is a small block c, having a projecting tongue, which slides in the groove of the circular cam plate d, d. This circular plate is supported on the frame e, and revolves on it by means of the flanged rollers f. Inside, at the top, the cam plate is toothed, and gears into a spur pinion connected with the hand wheel h. At gg is an inclined groove or shunt. When the tongues of the blocks c, c arrive at g, they slide up to a second groove, or the reverse, according as the cam plate is revolved in one direction or in the other. As this operation takes place with each sluice successively, any number of sluices can be opened or closed as desired. The turbine is of 48 horse power on 5.12 ft. fall, and the supply of water varies from 35 to 112 cub. ft. per second. The efficiency in normal working is given as 73%. The mean diameter of the wheel is 6 ft., and the speed 27.4 revolutions per minute.

Fig. 200.
 

Fig. 201.
 

Fig. 202.

As an example of a partial admission radial flow impulse turbine, a 100 h.p. turbine at Immenstadt may be taken. The fall varies from 538 to 570 ft. The external diameter of the wheel is 41/2 ft., and its internal diameter 3 ft. 10 in. Normal speed 400 revs. per minute. Water is discharged into the wheel by a single nozzle, shown in fig. 202 with its regulating apparatus and some of the vanes. The water enters the wheel at an angle of 22° with the direction of motion, and the final angle of the wheel vanes is 20°. The efficiency on trial was from 75 to 78%.

§ 199. Theory of the Impulse Turbine.—The theory of the impulse turbine does not essentially differ from that of the reaction turbine, except that there is no pressure in the wheel opposing the discharge from the guide-blades. Hence the velocity with which the water enters the wheel is simply

v_i = 0.96 √2g (H − ɧ),

where ɧ is the height of the top of the wheel above the tail water. If the hydropneumatic system is used, then ɧ = 0. Let Q_m be the maximum supply of water, r₁, r₂ the internal and external radii of the wheel at the inlet surface; then

u_i = Q_m / {π(r₂² − r₁²)}.

The value of u_i may be about 0.45 √2g (H − ɧ), whence r₁, r₂ can be determined.

The guide-blade angle is then given by the equation

sin γ = u_i / v_i = 0.45/0.94 = .48;

γ = 29°.

The value of u_i should, however, be corrected for the space occupied by the guide-blades.

The tangential velocity of the entering water is

w_i = v_i cos γ = 0.82 √2g (H − ɧ).

The circumferential velocity of the wheel may be (at mean radius)

V_i = 0.5 √2g (H − ɧ).

Hence the vane angle at inlet surface is given by the equation

cot θ = (w_i − V_i) / u_i = (0.82 − 0.5)/0.45 = .71;
θ = 55°.

The relative velocity of the water striking the vane at the inlet edge is v_ri = u_i cosec θ = 1.22u_i. This relative velocity remains unchanged during the passage of the water over the vane; consequently the relative velocity at the point of discharge is v_ro = 1.22u_i. Also in an axial flow turbine V_o = V_i.

If the final velocity of the water is axial, then

cos φ = V_o /v_ro = V_i /v_ri = 0.5/(1.22 × 0.45) = cos 24º 23′.

This should be corrected for the vane thickness. Neglecting this, u_o = v_ro sin φ = v_ri sin φ = u_i cosec θ sin φ = 0.5u_i. The discharging area of the wheel must therefore be greater than the inlet area in the ratio of at least 2 to 1. In some actual turbines the ratio is 7 to 3. This greater outlet area is obtained by splaying the wheel, as shown in the section (fig. 199).

Fig. 203.

§ 200. Pelton Wheel.—In the mining district of California about 1860 simple impulse wheels were used, termed hurdy-gurdy wheels. The wheels rotated in a vertical plane, being supported on a horizontal axis. Round the circumference were fixed flat vanes which were struck normally by a jet from a nozzle of size varying with the head and quantity of water. Such wheels have in fact long been used. They are not efficient, but they are very simply constructed. Then attempts were made to improve the efficiency, first by using hemispherical cup vanes, and then by using a double cup vane with a central dividing ridge, an arrangement invented by Pelton. In this last form the water from the nozzle passes half to each side of the wheel, just escaping clear of the backs of the advancing buckets. Fig. 203 shows a Pelton vane. Some small modifications have been made by other makers, but they are not of any great importance. Fig. 204 shows a complete Pelton wheel with frame and casing, supply pipe and nozzle. Pelton wheels have been very largely used in America and to some extent in Europe. They are extremely simple and easy to construct or repair and on falls of 100 ft. or more are very efficient. The jet strikes tangentially to the mean radius of the buckets, and the face of the buckets is not quite radial but at right angles to the direction of the jet at the point of first impact. For greatest efficiency the peripheral velocity of the wheel at the mean radius of the buckets should be a little less than half the velocity of the jet. As the radius of the wheel can be taken arbitrarily, the number of revolutions per minute can be accommodated to that of the machinery to be driven. Pelton wheels have been made as small as 4 in. diameter, for driving sewing machines, and as large as 24 ft. The efficiency on high falls is about 80%. When large power is required two or three nozzles are used delivering on one wheel. The width of the buckets should be not less than seven times the diameter of the jet.

Fig. 204.

At the Comstock mines, Nevada, there is a 36-in. Pelton wheel made of a solid steel disk with phosphor bronze buckets riveted to the rim. The head is 2100 ft. and the wheel makes 1150 revolutions per minute, the peripheral velocity being 180 ft. per sec. With a 1/2-in. nozzle the wheel uses 32 cub. ft. of water per minute and develops 100 h.p. At the Chollarshaft, Nevada, there are six Pelton wheels on a fall of 1680 ft. driving electrical generators. With 5/8-in. nozzles each develops 125 h.p.

Fig. 205

§ 201. Theory of the Pelton Wheel.—Suppose a jet with a velocity v strikes tangentially a curved vane AB (fig. 205) moving in the same direction with the velocity u. The water will flow over the vane with the relative velocity v−u and at B will have the tangential relative velocity v−u making an angle α with the direction of the vane’s motion. Combining this with the velocity u of the vane, the absolute velocity of the water leaving the vane will be w = Bc. The component of w in the direction of motion of the vane is Ba = Bb − ab = u − (v − u) cos α. Hence if Q is the quantity of water reaching the vane per second the change of momentum per second in the direction of the vane’s motion is (GQ/g) [v − {u − (v − u) cos α}] = (GQ/g) (v − u) (1 + cos α). If α = 0°, cos α = 1, and the change of momentum per second, which is equal to the effort driving the vane, is P = 2(GQ/g) (v − u). The work done on the vane is Pu = 2(GQ/g) (v − u)u. If a series of vanes are interposed in succession, the quantity of water impinging on the vanes per second is the total discharge of the nozzle, and the energy expended at the nozzle is GQv²/2g. Hence the efficiency of the arrangement is, when α = 0°, neglecting friction,

η = 2Pu/GQv² = 4 (v − u) u/v²,

which is a maximum and equal to unity if u = 1/2v. In that case the whole energy of the jet is usefully expended in driving the series of vanes. In practice α cannot be quite zero or the water leaving one vane would strike the back of the next advancing vane. Fig. 203 shows a Pelton vane. The water divides each way, and leaves the vane on each side in a direction nearly parallel to the direction of motion of the vane. The best velocity of the vane is very approximately half the velocity of the jet.

§ 202. Regulation of the Pelton Wheel.—At first Pelton wheels were adjusted to varying loads merely by throttling the supply. This method involves a total loss of part of the head at the sluice or throttle valve. In addition as the working head is reduced, the relation between wheel velocity and jet velocity is no longer that of greatest efficiency. Next a plan was adopted of deflecting the jet so that only part of the water reached the wheel when the load was reduced, the rest going to waste. This involved the use of an equal quantity of water for large and small loads, but it had, what in some cases is an advantage, the effect of preventing any water hammer in the supply pipe due to the action of the regulator. In most cases now regulation is effected by varying the section of the jet. A conical needle in the nozzle can be advanced or withdrawn so as to occupy more or less of the aperture of the nozzle. Such a needle can be controlled by an ordinary governor.

§ 203. General Considerations on the Choice of a Type of Turbine.—The circumferential speed of any turbine is necessarily a fraction of the initial velocity of the water, and therefore is greater as the head is greater. In reaction turbines with complete admission the number of revolutions per minute becomes inconveniently great, for the diameter cannot be increased beyond certain limits without greatly reducing the efficiency. In impulse turbines with partial admission the diameter can be chosen arbitrarily and the number of revolutions kept down on high falls to any desired amount. Hence broadly reaction turbines are better and less costly on low falls, and impulse turbines on high falls. For variable water flow impulse turbines have some advantage, being more efficiently regulated. On the other hand, impulse turbines lose efficiency seriously if their speed varies from the normal speed due to the head. If the head is very variable, as it often is on low falls, and the turbine must run at the same speed whatever the head, the impulse turbine is not suitable. Reaction turbines can be constructed so as to overcome this difficulty to a great extent. Axial flow turbines with vertical shafts have the disadvantage that in addition to the weight of the turbine there is an unbalanced water pressure to be carried by the footstep or collar bearing. In radial flow turbines the hydraulic pressures are balanced. The application of turbines to drive dynamos directly has involved some new conditions. The electrical engineer generally desires a high speed of rotation, and a very constant speed at all times. The reaction turbine is generally more suitable than the impulse turbine. As the diameter of the turbine depends on the quantity of water and cannot be much varied without great inefficiency, a difficulty arises on low falls. This has been met by constructing four independent reaction turbines on the same shaft, each having of course the diameter suitable for one-quarter of the whole discharge, and having a higher speed of rotation than a larger turbine. The turbines at Rheinfelden and Chevres are so constructed. To ensure constant speed of rotation when the head varies considerably without serious inefficiency, an axial flow turbine is generally used. It is constructed of three or four concentric rings of vanes, with independent regulating sluices, forming practically independent turbines of different radii. Any one of these or any combination can be used according to the state of the water. With a high fall the turbine of largest radius only is used, and the speed of rotation is less than with a turbine of smaller radius. On the other hand, as the fall decreases the inner turbines are used either singly or together, according to the power required. At the Zürich waterworks there are turbines of 90 h.p. on a fall varying from 101/2 ft. to 43/4 ft. The power and speed are kept constant. Each turbine has three concentric rings. The outermost ring gives 90 h.p. with 105 cub. ft. per second and the maximum fall. The outer and middle compartments give the same power with 140 cub. ft. per second and a fall of 7 ft. 10 in. All three compartments working together develop the power with about 250 cub. ft. per second. In some tests the efficiency was 74% with the outer ring working alone, 75.4% with the outer and middle ring working and a fall of 7 ft., and 80.7% with all the rings working.

Fig. 206.

§ 204. Speed Governing.—When turbines are used to drive dynamos direct, the question of speed regulation is of great importance. Steam engines using a light elastic fluid can be easily regulated by governors acting on throttle or expansion valves. It is different with water turbines using a fluid of great inertia. In one of the Niagara penstocks there are 400 tons of water flowing at 10 ft. per second, opposing enormous resistance to rapid change of speed of flow. The sluices of water turbines also are necessarily large and heavy. Hence relay governors must be used, and the tendency of relay governors to hunt must be overcome. In the Niagara Falls Power House No. 1, each turbine has a very sensitive centrifugal governor acting on a ratchet relay. The governor puts into gear one or other of two ratchets driven by the turbine itself. According as one or the other ratchet is in gear the sluices are raised or lowered. By a subsidiary arrangement the ratchets are gradually put out of gear unless the governor puts them in gear again, and this prevents the over correction of the speed from the lag in the action of the governor. In the Niagara Power House No. 2, the relay is an hydraulic relay similar in principle, but rather more complicated in arrangement, to that shown in fig. 206, which is a governor used for the 1250 h.p. turbines at Lyons. The sensitive governor G opens a valve and puts into action a plunger driven by oil pressure from an oil reservoir. As the plunger moves forward it gradually closes the oil admission valve by lowering the fulcrum end f of the valve lever which rests on a wedge w attached to the plunger. If the speed is still too high, the governor reopens the valve. In the case of the Niagara turbines the oil pressure is 1200 ℔ per sq. in. One millimetre of movement of the governor sleeve completely opens the relay valve, and the relay plunger exerts a force of 50 tons. The sluices can be completely opened or shut in twelve seconds. The ordinary variation of speed of the turbine with varying load does not exceed 1%. If all the load is thrown off, the momentary variation of speed is not more than 5%. To prevent hydraulic shock in the supply pipes, a relief valve is provided which opens if the pressure is in excess of that due to the head.

Fig. 207.

§ 205. The Hydraulic Ram.—The hydraulic ram is an arrangement by which a quantity of water falling a distance h forces a portion of the water to rise to a height h₁, greater than h. It consists of a supply reservoir (A, fig. 207), into which the water enters from some natural stream. A pipe s of considerable length conducts the water to a lower level, where it is discharged intermittently through a self-acting pulsating valve at d. The supply pipe s may be fitted with a flap valve for stopping the ram, and this is attached in some cases to a float, so that the ram starts and stops itself automatically, according as the supply cistern fills or empties. The lower float is just sufficient to keep open the flap after it has been raised by the action of the upper float. The length of chain is adjusted so that the upper float opens the flap when the level in the cistern is at the desired height. If the water-level falls below the lower float the flap closes. The pipe s should be as long and as straight as possible, and as it is subjected to considerable pressure from the sudden arrest of the motion of the water, it must be strong and strongly jointed. a is an air vessel, and e the delivery pipe leading to the reservoir at a higher level than A, into which water is to be pumped. Fig. 208 shows in section the construction of the ram itself. d is the pulsating discharge valve already mentioned, which opens inwards and downwards. The stroke of the valve is regulated by the cotter through the spindle, under which are washers by which the amount of fall can be regulated. At o is a delivery valve, opening outwards, which is often a ball-valve but sometimes a flap-valve. The water which is pumped passes through this valve into the air vessel a, from which it flows by the delivery pipe in a regular stream into the cistern to which the water is to be raised. In the vertical chamber behind the outer valve a small air vessel is formed, and into this opens an aperture 1/4 in. in diameter, made in a brass screw plug b. The hole is reduced to 1/16 in. in diameter at the outer end of the plug and is closed by a small valve opening inwards. Through this, during the rebound after each stroke of the ram, a small quantity of air is sucked in which keeps the air vessel supplied with its elastic cushion of air.

Fig. 208.

During the recoil after a sudden closing of the valve d, the pressure below it is diminished and the valve opens, permitting outflow. In consequence of the flow through this valve, the water in the supply pipe acquires a gradually increasing velocity. The upward flow of the water, towards the valve d, increases the pressure tending to lift the valve, and at last, if the valve is not too heavy, lifts and closes it. The forward momentum of the column in the supply pipe being destroyed by the stoppage of the flow, the water exerts a pressure at the end of the pipe sufficient to open the delivery valve o, and to cause a portion of the water to flow into the air vessel. As the water in the supply pipe comes to rest and recoils, the valve d opens again and the operation is repeated. Part of the energy of the descending column is employed in compressing the air at the end of the supply pipe and expanding the pipe itself. This causes a recoil of the water which momentarily diminishes the pressure in the pipe below the pressure due to the statical head. This assists in opening the valve d. The recoil of the water is sufficiently great to enable a pump to be attached to the ram body instead of the direct rising pipe. With this arrangement a ram working with muddy water may be employed to raise clear spring water. Instead of lifting the delivery valve as in the ordinary ram, the momentum of the column drives a sliding or elastic piston, and the recoil brings it back. This piston lifts and forces alternately the clear water through ordinary pump valves.

Pumps

§ 206. The different classes of pumps correspond almost exactly to the different classes of water motors, although the mechanical details of the construction are somewhat different. They are properly reversed water motors. Ordinary reciprocating pumps correspond to water-pressure engines. Chain and bucket pumps are in principle similar to water wheels in which the water acts by weight. Scoop wheels are similar to undershot water wheels, and centrifugal pumps to turbines.

Reciprocating Pumps are single or double acting, and differ from water-pressure engines in that the valves are moved by the water instead of by automatic machinery. They may be classed thus:—

1. Lift Pumps.—The water drawn through a foot valve on the ascent of the pump bucket is forced through the bucket valve when it descends, and lifted by the bucket when it reascends. Such pumps give an intermittent discharge.

2. Plunger or Force Pumps, in which the water drawn through the foot valve is displaced by the descent of a solid plunger, and forced through a delivery valve. They have the advantage that the friction is less than that of lift pumps, and the packing round the plunger is easily accessible, whilst that round a lift pump bucket is not. The flow is intermittent.

3. The Double-acting Force Pump is in principle a double plunger pump. The discharge fluctuates from zero to a maximum and back to zero each stroke, but is not arrested for any appreciable time.

4. Bucket and Plunger Pumps consist of a lift pump bucket combined with a plunger of half its area. The flow varies as in a double-acting pump.

5. Diaphragm Pumps have been used, in which the solid plunger is replaced by an elastic diaphragm, alternately depressed into and raised out of a cylinder.

As single-acting pumps give an intermittent discharge three are generally used on cranks at 120°. But with all pumps the variation of velocity of discharge would cause great waste of work in the delivery pipes when they are long, and even danger from the hydraulic ramming action of the long column of water. An air vessel is interposed between the pump and the delivery pipes, of a volume from 5 to 100 times the space described by the plunger per stroke. The air in this must be replenished from time to time, or continuously, by a special air-pump. At low speeds not exceeding 30 ft. per minute the delivery of a pump is about 90 to 95% of the volume described by the plunger or bucket, from 5 to 10% of the discharge being lost by leakage. At high speeds the quantity pumped occasionally exceeds the volume described by the plunger, the momentum of the water keeping the valves open after the turn of the stroke.

The velocity of large mining pumps is about 140 ft. per minute, the indoor or suction stroke being sometimes made at 250 ft. per minute. Rotative pumping engines of large size have a plunger speed of 90 ft. per minute. Small rotative pumps are run faster, but at some loss of efficiency. Fire-engine pumps have a speed of 180 to 220 ft. per minute.

The efficiency of reciprocating pumps varies very greatly. Small reciprocating pumps, with metal valves on lifts of 15 ft., were found by Morin to have an efficiency of 16 to 40%, or on the average 25%. When used to pump water at considerable pressure, through hose pipes, the efficiency rose to from 28 to 57%, or on the average, with 50 to 100 ft. of lift, about 50%. A large pump with barrels 18 in. diameter, at speeds under 60 ft. per minute, gave the following results:—

Lift in feet	141/2	34	47
Efficiency	.46	.66	.70

The very large steam-pumps employed for waterworks, with 150 ft. or more of lift, appear to reach an efficiency of 90%, not including the friction of the discharge pipes. Reckoned on the indicated work of the steam-engine the efficiency may be 80%.

Many small pumps are now driven electrically and are usually three-throw single-acting pumps driven from the electric motor by gearing. It is not convenient to vary the speed of the motor to accommodate it to the varying rate of pumping usually required. Messrs Hayward Tyler have introduced a mechanism for varying the stroke of the pumps (Sinclair’s patent) from full stroke to nil, without stopping the pumps.

§ 207. Centrifugal Pump.—For large volumes of water on lifts not exceeding about 60 ft. the most convenient pump is the centrifugal pump. Recent improvements have made it available also for very high lifts. It consists of a wheel or fan with curved vanes enclosed in an annular chamber. Water flows in at the centre and is discharged at the periphery. The fan may rotate in a vertical or horizontal plane and the water may enter on one or both sides of the fan. In the latter case there is no axial unbalanced pressure. The fan and its casing must be filled with water before it can start, so that if not drowned there must be a foot valve on the suction pipe. When no special attention needs to be paid to efficiency the water may have a velocity of 6 to 7 ft. in the suction and delivery pipes. The fan often has 6 to 12 vanes. For a double-inlet fan of diameter D, the diameter of the inlets is D/2. If Q is the discharge in cub. ft. per second D = about 0.6 √Q in average cases. The peripheral speed is a little greater than the velocity due to the lift. Ordinary centrifugal pumps will have an efficiency of 40 to 60%.

The first pump of this kind which attracted notice was one exhibited by J. G. Appold in 1851, and the special features of his pump have been retained in the best pumps since constructed. Appold’s pump raised continuously a volume of water equal to 1400 times its own capacity per minute. It had no valves, and it permitted the passage of solid bodies, such as walnuts and oranges, without obstruction to its working. Its efficiency was also found to be good.

Fig. 209.

Fig. 209 shows the ordinary form of a centrifugal pump. The pump disk and vanes B are cast in one, usually of bronze,

and the disk is keyed on the driving shaft C. The casing A has a spirally enlarging discharge passage into the discharge pipe K. A cover L gives access to the pump. S is the suction pipe which opens into the pump disk on both sides at D.

Fig. 210 shows a centrifugal pump differing from ordinary centrifugal pumps in one feature only. The water rises through a suction pipe S, which divides so as to enter the pump wheel W at the centre on each side. The pump disk or wheel is very similar to a turbine wheel. It is keyed on a shaft driven by a belt on a fast and loose pulley arrangement at P. The water rotating in the pump disk presses outwards, and if the speed is sufficient a continuous flow is maintained through the pump and into the discharge pipe D. The special feature in this pump is that the water, discharged by the pump disk with a whirling velocity of not inconsiderable magnitude, is allowed to continue rotation in a chamber somewhat larger than the pump. The use of this whirlpool chamber was first suggested by Professor James Thomson. It utilizes the energy due to the whirling velocity of the water which in most pumps is wasted in eddies in the discharge pipe. In the pump shown guide-blades are also added which have the direction of the stream lines in a free vortex. They do not therefore interfere with the action of the water when pumping the normal quantity, but only prevent irregular motion. At A is a plug by which the pump case is filled before starting. If the pump is above the water to be pumped, a foot valve is required to permit the pump to be filled. Sometimes instead of the foot valve a delivery valve is used, an air-pump or steam jet pump being employed to exhaust the air from the pump case.

Fig. 210.

§ 208. Design and Proportions of a Centrifugal Pump.—The design of the pump disk is very simple. Let r_i , r_o be the radii of the inlet and outlet surfaces of the pump disk, d_i , d_o the clear axial width at those radii. The velocity of flow through the pump may be taken the same as for a turbine. If Q is the quantity pumped, and H the lift,

u_i = 0.25 √2gH.
2πr_id_i = Q / u_i.

(1)

Also in practice

Hence,

di = 1.2 ri . . . .	${\bigg \}}$
ri = .2571 √ (Q / √H).

(2) 

Usually

ro = 2ri,

and

do = di or 1/2 di

according as the disk is parallel-sided or coned. The water enters the wheel radially with the velocity u_i, and

u_o = Q / 2πr_od_o.

(3)

Fig. 211.

Fig. 211 shows the notation adopted for the velocities. Suppose the water enters the wheel with the velocity v_i, while the velocity of the wheel is V_i. Completing the parallelogram, v_ri is the relative velocity of the water and wheel, and is the proper direction of the wheel vanes. Also, by resolving, u_i and w_i are the component velocities of flow and velocities of whir of the velocity v_i of the water. At the outlet surface, v_o is the final velocity of discharge, and the rest of the notation is similar to that for the inlet surface.

Usually the water flows equally in all directions in the eye of the wheel, in that case v_i is radial. Then, in normal conditions of working, at the inlet surface,

 v_i = u_i  w_i = 0 tan θ = u_i / V_i  v_ri = u_i cosec θ = √ (u_i² + V_i²).

(4)

If the pump is raising less or more than its proper quantity, θ will not satisfy the last condition, and there is then some loss of head in shock.

At the outer circumference of the wheel or outlet surface,

v_ro = u_o cosec φ w_o = V_o − u_o cot φ v_o = √ {u_o² + (V_o − u_o cot φ)²}

(5)

Variation of Pressure in the Pump Disk.—Precisely as in the case of turbines, it can be shown that the variation of pressure between the inlet and outlet surfaces of the pump is

h_o − h_i = (V_o² − V_i²) / 2g − (v_ro² − v_ri²) / 2g.

Inserting the values of v_ro, v_ri in (4) and (5), we get for normal conditions of working

h_o − h_i = (V_o² − V_i²) / 2g − u_o² cosec² φ / 2g + (u_i² + V_i²) / 2g
= V_o² / 2g − u_o² cosec² φ / 2g + u_i² / 2g.

(6)

Hydraulic Efficiency of the Pump.—Neglecting disk friction, journal friction, and leakage, the efficiency of the pump can be found in the same way as that of turbines (§ 186). Let M be the moment of the couple rotating the pump, and α its angular velocity; w_o, r_o the tangential velocity of the water and radius at the outlet surface; w_i, r_i the same quantities at the inlet surface. Q being the discharge per second, the change of angular momentum per second is

(GQ/g) (w_or_o − w_ir_i).

Hence

M = (GQ/g) (w_or_o − w_ir_i).

In normal working, w_i = 0. Also, multiplying by the angular velocity, the work done per second is

Mα = (GQ/g) w_or_oα.

But the useful work done in pumping is GQH. Therefore the efficiency is

η = GQH / Mα = gH / w_or_oα = gH / w_oV_o.

(7)

§ 209. Case 1. Centrifugal Pump with no Whirlpool Chamber.—When no special provision is made to utilize the energy of motion of the water leaving the wheel, and the pump discharges directly into a chamber in which the water is flowing to the discharge pipe, nearly the whole of the energy of the water leaving the disk is wasted. The water leaves the disk with the more or less considerable velocity v_o, and impinges on a mass flowing to the discharge pipe at the much slower velocity v_s. The radial component of v_o is almost necessarily wasted. From the tangential component there is a gain of pressure

(w_o² − v_s²) / 2g − (w_o − v_s)² / 2g
= v_s (w_o − v_s) / g,

which will be small, if v_s is small compared with w_o. Its greatest value, if v_s = 1/2w_o, is 1/2w_o²/2g, which will always be a small part of the whole head. Suppose this neglected. The whole variation of pressure in the pump disk then balances the lift and the head u_i²/2g necessary to give the initial velocity of flow in the eye of the wheel.

u_i² / 2g + H = V_o² / 2g − u_o² cosec² φ / 2g + u_i² / 2g,

H = V_o² / 2g − u_o² cosec² φ / 2g

or

V_o = √ (2gH + u_o² cosec² φ).

(8)

and the efficiency of the pump is, from (7),

η = gH / V_ow_o = gH / {V (V_o − n_o cot φ) },
  = (V_o² − u_o² cosec² φ) / {2V_o (V_o − u_o cot φ) },

(9)

For φ = 90°,

η = (V_o² − u_o²) / 2V_o²,

which is necessarily less than 1/2. That is, half the work expended in driving the pump is wasted. By recurving the vanes, a plan introduced by Appold, the efficiency is increased, because the velocity v_o of discharge from the pump is diminished. If φ is very small,

cosec φ = cot φ;

and then

η = (V_o, + u_o cosec φ) / 2V_o,

which may approach the value 1, as φ tends towards 0. Equation (8) shows that u_o cosec φ cannot be greater than V_o. Putting u_o = 0.25 √(2gH) we get the following numerical values of the efficiency and the circumferential velocity of the pump:—

φ	η	Vo
90°	0.47	1.03 √2gH
45°	0.56	1.06 ”
30°	0.65	1.12 ”
20°	0.73	1.24 ”
10°	0.84	1.75 ”

φ cannot practically be made less than 20°; and, allowing for the frictional losses neglected, the efficiency of a pump in which φ = 20° is found to be about .60.

§ 210. Case 2. Pump with a Whirlpool Chamber, as in fig. 210.—Professor James Thomson first suggested that the energy of the water after leaving the pump disk might be utilized, if a space were left in which a free vortex could be formed. In such a free vortex the velocity varies inversely as the radius. The gain of pressure in the vortex chamber is, putting r_o, r_w for the radii to the outlet surface of wheel and to outside of free vortex,

${\frac {v_{o}^{2}}{2g}}{\Big (}1-{\frac {r_{o}^{2}}{r_{w}^{2}}}{\Big )}={\frac {v_{o}^{2}}{2g}}{\Big (}1-k^{2}{\Big )},$

if

k = ro / rw.

The lift is then, adding this to the lift in the last case,

H＝{V_o² − u_o² cosec² φ + v_o² (1 − k²)} / 2g.

But

vo2 = Vo2 − 2Vouo cot φ + uo2 cosec2 φ;

∴ H = {(2 − k2) Vo2 − 2kVouo cot φ − k2uo2 cosec2 φ} / 2g.

(10)

Putting this in the expression for the efficiency, we find a considerable increase of efficiency. Thus with

φ = 90° and	k = 1/2,	η = 7/8 nearly,
φ a small angle and	k = 1/2,	η = 1 nearly.

With this arrangement of pump, therefore, the angle at the outer ends of the vanes is of comparatively little importance. A moderate angle of 30° or 40° may very well be adopted. The following numerical values of the velocity of the circumference of the pump have been obtained by taking k = 1/2, and u_o = 0.25√(2gH).

φ	Vo
90°	.762	√2gH
45°	.842	”
30°	.911	”
20°	1.023	”

The quantity of water to be pumped by a centrifugal pump necessarily varies, and an adjustment for different quantities of water cannot easily be introduced. Hence it is that the average efficiency of pumps of this kind is in practice less than the efficiencies given above. The advantage of a vortex chamber is also generally neglected. The velocity in the supply and discharge pipes is also often made greater than is consistent with a high degree of efficiency. Velocities of 6 or 7 ft. per second in the discharge and suction pipes, when the lift is small, cause a very sensible waste of energy; 3 to 6 ft. would be much better. Centrifugal pumps of very large size have been constructed. Easton and Anderson made pumps for the North Sea canal in Holland to deliver each 670 tons of water per minute on a lift of 5 ft. The pump disks are 8 ft. diameter. J. and H. Gwynne constructed some pumps for draining the Ferrarese Marshes, which together deliver 2000 tons per minute. A pump made under Professor J. Thomson’s direction for drainage works in Barbados had a pump disk 16 ft. in diameter and a whirlpool chamber 32 ft. in diameter. The efficiency of centrifugal pumps when delivering less or more than the normal quantity of water is discussed in a paper in the Proc. Inst. Civ. Eng. vol. 53.

§ 211. High Lift Centrifugal Pumps.—It has long been known that centrifugal pumps could be worked in series, each pump overcoming a part of the lift. This method has been perfected, and centrifugal pumps for very high lifts with great efficiency have been used by Sulzer and others. C. W. Darley (Proc. Inst. Civ. Eng., supplement to vol. 154, p. 156) has described some pumps of this new type driven by Parsons steam turbines for the water supply of Sydney, N.S.W. Each pump was designed to deliver 11/2 million gallons per twenty-four hours against a head of 240 ft. at 3300 revs. per minute. Three pumps in series give therefore a lift of 720 ft. The pump consists of a central double-sided impeller 12 in. diameter. The water entering at the bottom divides and enters the runner at each side through a bell-mouthed passage. The shaft is provided with ring and groove glands which on the suction side keep the air out and on the pressure side prevent leakage. Some water from the pressure side leaks through the glands, but beyond the first grooves it passes into a pocket and is returned to the suction side of the pump. For the glands on the suction side water is supplied from a low-pressure service. No packing is used in the glands. During the trials no water was seen at the glands. The following are the results of tests made at Newcastle:—

		I.	II.	III.	IV.
Duration of test	hours	2	1.54	1.2	1.55
Steam pressure	℔ per sq. in.	57	57	84	55
Weight of steam per water h.p. hour	℔	27.93	30.67	28.83	27.89
Speed in revs, per min.		3300	3330	3710	3340
Height of suction	ft.	11	11	11	11
Total lift	ft.	762	744	917	756
Million galls. per day pumped—
By Venturi meter		1.573	1.499	1.689	1.503
By orifice		1.623	1.513	1.723	1.555
Water h.p.		252	235	326	239

In trial IV. the steam was superheated 95° F. From other trials under the same conditions as trial I. the Parsons turbine uses 15.6 ℔ of steam per brake h.p. hour, so that the combined efficiency of turbine and pumps is about 56%, a remarkably good result.

Fig. 212.

§ 212. Air-Lift Pumps.—An interesting and simple method of pumping by compressed air, invented by Dr J. Pohlé of Arizona, is likely to be very useful in certain cases. Suppose a rising main placed in a deep bore hole in which there is a considerable depth of water. Air compressed to a sufficient pressure is conveyed by an air pipe and introduced at the lower end of the rising main. The air rising In the main diminishes the average density of the contents of the main, and their aggregate weight no longer balances the pressure at the lower end of the main due to its submersion. An upward flow is set up, and if the air supply is sufficient the water in the rising main is lifted to any required height. The higher the lift above the level in the bore hole the deeper must be the point at which air is injected. Fig. 212 shows an airlift pump constructed for W. H. Maxwell at the Tunbridge Wells waterworks. There is a two-stage steam air compressor, compressing air to from 90 to 100 ℔ per sq. in. The bore hole is 350 ft. deep, lined with steel pipes 15 in. diameter for 200 ft. and with perforated pipes 131/2 in. diameter for the lower 150 ft. The rest level of the water is 96 ft. from the ground-level, and the level when pumping 32,000 gallons per hour is 120 ft. from the ground-level. The rising main is 7 in. diameter, and is carried nearly to the bottom of the bore hole and to 20 ft. above the ground-level. The air pipe is 21/2 in. diameter. In a trial run 31,402 gallons per hour were raised 133 ft. above the level in the well. Trials of the efficiency of the system made at San Francisco with varying conditions will be found in a paper by E. A. Rix (Journ. Amer. Assoc. Eng. Soc. vol. 25, 1900). Maxwell found the best results when the ratio of immersion to lift was 3 to 1 at the start and 2.2 to 1 at the end of the trial. In these conditions the efficiency was 37% calculated on the indicated h.p. of the steam-engine, and 46% calculated on the indicated work of the compressor. 2.7 volumes of free air were used to 1 of water lifted. The system is suitable for temporary purposes, especially as the quantity of water raised is much greater than could be pumped by any other system in a bore hole of a given size. It is useful for clearing a boring of sand and may be advantageously used permanently when a boring is in sand or gravel which cannot be kept out of the bore hole. The initial cost is small.

§ 213. Centrifugal Fans.—Centrifugal fans are constructed similarly to centrifugal pumps, and are used for compressing air to pressures not exceeding 10 to 15 in. of water-column. With this small variation of pressure the variation of volume and density of the air may be neglected without sensible error. The conditions of pressure and discharge for fans are generally less accurately known than in the case of pumps, and the design of fans is generally somewhat crude. They seldom have whirlpool chambers, though a large expanding outlet is provided in the case of the important Guibal fans used in mine ventilation.

It is usual to reckon the difference of pressure at the inlet and outlet of a fan in inches of water-column. One inch of water-column = 64.4 ft. of air at average atmospheric pressure = 5.2℔ per sq. ft.

Roughly the pressure-head produced in a fan without means of utilizing the kinetic energy of discharge would be v²/2g ft. of air, or 0.00024 v² in. of water, where v is the velocity of the tips of the fan blades in feet per second. If d is the diameter of the fan and t the width at the external circumference, then πdt is the discharge area of the fan disk. If Q is the discharge in cub. ft. per sec., u = Q/π dt is the radial velocity of discharge which is numerically equal to the discharge per square foot of outlet in cubic feet per second. As both the losses in the fan and the work done are roughly proportional to u² in fans of the same type, and are also proportional to the gauge pressure p, then if the losses are to be a constant percentage of the work done u may be taken proportional to √p. In ordinary cases u = about 22 √p. The width t of the fan is generally from 0.35 to 0.45d. Hence if Q is given, the diameter of the fan should be:—

For t = 0.35d,    d = 0.20 √ (Q / √p)
For t = 0.45d,    d = 0.18 √ (Q / √p)

If p is the pressure difference in the fan in inches of water, and N the revolutions of fan,

v = πdN/60	ft. per sec.
N = 1230√p/d	revs. per min.

As the pressure difference is small, the work done in compressing the air is almost exactly 5.2pQ foot-pounds per second. Usually, however, the kinetic energy of the air in the discharge pipe is not inconsiderable compared with the work done in compression. If w is the velocity of the air where the discharge pressure is measured, the air carries away w²/2g foot-pounds per ℔ of air as kinetic energy. In Q cubic feet or 0.0807Q ℔ the kinetic energy is 0.00125 Qw² foot-pounds per second.

The efficiency of fans is reckoned in two ways. If B.H.P. is the effective horse-power applied at the fan shaft, then the efficiency reckoned on the work of compression is

η = 5.2pQ / 550 B.H.P.

On the other hand, if the kinetic energy in the delivery pipe is taken as part of the useful work the efficiency is

η₂ = (5.2 pQ + 0.00125 Qw²) / 550 B.H.P.

Although the theory above is a rough one it agrees sufficiently with experiment, with some merely numerical modifications.

An extremely interesting experimental investigation of the action of centrifugal fans has been made by H. Heenan and W. Gilbert (Proc. Inst. Civ. Eng. vol. 123, p. 272). The fans delivered through an air trunk in which different resistances could be obtained by introducing diaphragms with circular apertures of different sizes. Suppose a fan run at constant speed with different resistances and the compression pressure, discharge and brake horse-power measured. The results plot in such a diagram as is shown in fig. 213. The less the resistance to discharge, that is the larger the opening in the air trunk, the greater the quantity of air discharged at the given speed of the fan. On the other hand the compression pressure diminishes. The curve marked total gauge is the compression pressure + the velocity head in the discharge pipe, both in inches of water. This curve falls, but not nearly so much as the compression curve, when the resistance in the air trunk is diminished. The brake horse-power increases as the resistance is diminished because the volume of discharge increases very much. The curve marked efficiency is the efficiency calculated on the work of compression only. It is zero for no discharge, and zero also when there is no resistance and all the energy given to the air is carried away as kinetic energy. There is a discharge for which this efficiency is a maximum; it is about half the discharge which there is when there is no resistance and the delivery pipe is full open. The conditions of speed and discharge corresponding to the greatest efficiency of compression are those ordinarily taken as the best normal conditions of working. The curve marked total efficiency gives the efficiency calculated on the work of compression and kinetic energy of discharge. Messrs Gilbert and Heenan found the efficiencies of ordinary fans calculated on the compression to be 40 to 60% when working at about normal conditions.

Fig. 213.

Taking some of Messrs Heenan and Gilbert’s results for ordinary fans in normal conditions, they have been found to agree fairly with the following approximate rules. Let p_c be the compression pressure and q the volume discharged per second per square foot of outlet area of fan. Then the total gauge pressure due to pressure of compression and velocity of discharge is approximately: p = p_c + 0.0004q² in. of water, so that if p_c is given, p can be found approximately. The pressure p depends on the circumferential speed v of the fan disk—

p = 0.00025 v² in. of water
v = 63 √p ft. per sec.

The discharge per square foot of outlet of fan is—

q = 15 to 18 √p cub. ft. per sec.

The total discharge is

Q = π dtq = 47 to 56 dt √p

t = .45d,   d = 0.20 to 0.22 √(Q / √p) ft.
N = 1203 √ p/d.

These approximate equations, which are derived purely from experiment, do not differ greatly from those obtained by the rough theory given above. The theory helps to explain the reason for the form of the empirical results.  (W. C. U.) 

1. Except where other units are given, the units throughout this article are feet, pounds, pounds per sq. ft., feet per second.
2. Journal de M. Liouville, t. xiii. (1868); Mémoires de l’Académie, des Sciences de l’Institut de France, t. xxiii., xxiv. (1877).
3. The following theorem is taken from a paper by J. H. Cotterill, “On the Distribution of Energy in a Mass of Fluid in Steady Motion,” Phil. Mag., February 1876.
4. The discharge per second varied from .461 to .665 cub. ft. in two experiments. The coefficient .435 is derived from the mean value.
5. “Formulae for the Flow of Water in Pipes,” Industries (Manchester, 1886).
6. Boussinesq has shown that this mode of determining the corrective factor α is not satisfactory.
7. In general, because when the water leaves the turbine wheel it ceases to act on the machine. If deflecting vanes or a whirlpool are added to a turbine at the discharging side, then v₁ may in part depend on v₂, and the statement above is no longer true.