SF GIFT OF MICHAEL REE^E $ , \ f WORKS OF PROF. HENRY T. BOVEY PUBLISHED BY JOHN WILEY & SONS. A Treatise on Hydraulics. 8vo, cloth, $4.00. Strength of Materials and Theory of Structures. 830 pages, 8vo, cloth, $7.50. A TREATISE ON HYDRAU LICS. BY HENRY T. ^JO M. INST.C.E., LL.D., F.R.S.C., Professor of Civil Engineering and Applied Mechanics, McGill University, Montreal. FIRST EDITION. FIRST THOUSAND. NEW YORK: JOHN WILEY & SONS. LONDON : CHAPMAN & HALL, LIMITED, 1895. Copyright, 1895, BY HENRY T. BOVEY. ROBERT DRUMMOND, HLHCTROTYPBR AND PRINTER, NEW YORK. PREFACE. THE present treatise is the outcome of lectures delivered in McGill University during the last ten or twelve years, and although intended primarily for the use and convenience of the student of hydraulics, it is hoped that it may also prove acceptable to the engineer in general practice. In order to render the treatment of the subject more com- plete, free reference has been made to standard authors on the subject. The examples introduced to illustrate the text have also been selected in part from the works of such well-known writers as Weisbach, Osborne Reynolds, and Cotterill, but the greater number are such as have occurred in the course of the author's own experience. The tables of coefficients of discharge have been prepared from the results of experiments carried out in the Hydraulic Laboratory of the University. These experiments are still being continued and may probably form the subject of a special paper. The author desires to acknowledge many suggestions offered by Professor Bamford, and to express his deep obliga- tion to Professor Chandler for much labor and time given to the revision of proof sheets. HENRY T. BOVEY. MONTREAL, November, 1895. CONTENTS. CHAPTER I. FLOW THROUGH ORIFICES AND OVER WEIRS. PAGE Definitions i Stream-line Motion 2 Motion in Plain Layers 2 Laminar Motion 2 Density of Water 2 Continuity 2 Bernouilli's Theorem 6 Applications of Bernouilli's Theorem 9 Piezometer 9 Orifice in a Thin Plate 13 Torricelli's Theorem 14 Efflux from Orifice in a Vessel in Motion 16 Flow in a Frictionless Pipe of Gradually Changing Section 18 Hydraulic Resistances 20 Coefficient of Velocity 20 Coefficient of Resistance 21 Coefficient of Contraction 22 Coefficient of Discharge 24 Miner's Inch 26 Energy and Momentum of Jet 27 Inversion of the Jet 27 Time Required to Empty and Fill a Lock 29 General Equations 30 Loss of Energy in Shock 32 Mouthpieces 34 Borda's Mouthpiece 34 Ring Nozzle 37 Cylindrical Mouthpiece 39 Divergent Mouthpiece 42 Convergent Mouthpiece 44 Radiating Current 46 Vortex Motion 47 Free Spiral Vortex 48 Forced Vortex 49 v vi CONTENTS. Compound Vortex 50 Large Orifices 50 Rectangular Orifices of Large Size 50 Circular Orifices of Large Size 53 Notches 54 Weirs 54 Triangular Notch 56 Broad-crested Weir 58 Examples 60 CHAPTER II. FLUID FRICTION. Definition 70 Laws of Fluid Friction 72 Surface Friction in Pipes 73 Resistance of Ships 76 CHAPTER III. FLOW IN PIPES. Assumptions .* 78 Steady Motion .. 78 Influence upon the Flow of the Pipe's Position 83 Transmission of Energy by Hydraulic Pressure 84 Flow in a Uniform Pipe Connecting Two Reservoirs 86 Losses of Head due to Abrupt Changes of Section 89 Remarks on the Law of Resistance to Flow 96 Flow in a Pipe of Varying Diameter 98 Equivalent Uniform Main 100 Branch Main of Uniform Diameter 101 Nozzles 104 Motor Driven by Hydraulic Pressure 107 Siphons 108 Inverted Siphons 109 Air in Pipe no Three Reservoirs Connected by a Branched Pipe m Orifice Fed by Two Reservoirs 115 Variation of Velocity in a Transverse Section 119 Examples 122 CHAPTER IV. FLOW OF WATER IN OPEN CHANNELS. Flow of Water in Channels t 131 Steady Flow in Channels of Constant Section 132 CONTENTS. Vll Form of a Channel 135 Flow in Aqueducts 142 River Bends 143 Value of/ 144 Darcy and Bazin's Formulae 145 Ganguillet and Kiitter's Formulas 147 Variation of Velocity in a Transverse Section 148 Bazin's Formula 152 Boileau's Formula 153 Relations between Surface, Mean, and Bottom Velocities 154 Flow of Water in Open Channels of Varying Cross-section 156 Standing Wave 165 Examples 170 CHAPTER V. METHODS OF GAUGING. Gauging of Streams and Watercourses 173 Hook Gauge '. 173 Surface-floats 175 Subsurface-floats 175 Twin-floats 176 Velocity Rod 176 Pitot Tube 176 Darcy Gauge 178 Current Meters 180 Hydrometric Pendulum 183 Gauging of Pipes 183 Venturi Meter 183 Piston Meter 184 Inferential Meter 184 CHAPTER VI. IMPACT. Impact upon a Flat Vane Oblique to Direction of Jet 186 Impact upon a Flat Vane Normal to Direction of Jet 189 Reaction 190 Jet Propeller 190 Impact upon a Surface of Revolution 192 Impact upon a Flat Vane with Rim 195 Pressure in a Pipe upon a Thin Plate Normal to the Direction of Motion. 196 Pressure in a Pipe upon a Cylindrical Body about Three Diameters in Length 198 Impact upon a Curved Vane 199 Frictional Effect 205 Resistance to the Motion of a Solid in a Fluid Mass 205 Examples ; 209 Vlll CONTENTS. CHAPTER VII. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. PAGE Classification 213 Hydraulic Ram 214 Pressure Engine 215 Accumulator 215 Losses of Energy in Pressure Engines 221 Hydraulic Brakes 223 Water-wheels ^ 225 Undershot Wheels 225 Wheels in a Straight Race 227 Poncelet Wheel 232 Form of Bucket 240 Breast-wheels 242 Sluices 244 Overshot Wheels 254 Effect of Centrifugal Force 255 Weight of Water on Wheel 256 Arc of Discharge 256 Pitch-back Wheel 272 Ventilated Bucket 272 Jet Reaction Wheel > 272 Barker's Mill 272 Scotch Turbine 276 Reaction Turbines 276- Impulse Turbines 276 Hurdy-gurdy Wheel 279 Pelton Wheel 280 Radial-, Axial-, and Mixed-flow Turbines 281-284 Limit Turbine 283 Theory of Turbine 284 Remarks on Centrifugal Head in Turbine-flow 298 Practical Values of the Velocities, etc. , in Turbines 299 Theory of the Section-tube 301 Losses of Energy in Turbines 303 Centrifugal Pumps ." 307 Theory of Centrifugal Pumps 309 Examples 315 HYDRAULICS. CHAPTER I. FLOW THROUGH ORIFICES, OVER WEIRS, ETC. I. Fluid Motion. The term " hydraulics," as its derivation (vdoop, water ; avXos, a tube or pipe) indicates, was primarily applied to the conveyance of water in a tube or pipe, but its meaning now embraces the experimental theory of the motion of fluids. The motion of a fluid is said to be steady or permanent when the molecules successively arriving at any given point are animated with the same velocity, are subjected to the same pressure, and are the same in density. As soon as the motion of a stream becomes steady a permanent regime is said to be established, and hydraulic investigations are usually made on the hypothesis of a permanent regime. With such an hypothesis any portion of the fluid mass which leaves a given region is replaced by a like portion under conditions which are identically the same. The terms "steady motion" and " permanent regime" are often considered to be synonymous. The general problem of flow is the determination of the relation which exists at any point between the density, press- 2 HYDRA ULICS. ure, and velocity of the molecules which successively pass that point. The actual motion of a fluid is exceedingly complex, and in order to simplify the investigations various assumptions are made as to the nature of the flow. 2. (a) Stream-line Motion. The molecules may be re- garded as flowing along definite paths, and a succession of such molecules will form a continuous fluid rope which is termed an elementary stream or a fluid filament, or, if the motion is steady and the paths therefore fixed in space, a stream-line. Experiment shows that the velocity of flow in any cross- section varies from point to point, and hence it is often assumed that the section is made up of an infinite number of indefi- nitely small areas, each area being the section of a fluid filament. (b) Motion in Plane Layers. In this motion it is assumed that the molecules which at any given moment are found in a plane layer will remain in a plane layer after they have moved into any new position. (c) Laminar Motion. On this hypothesis the stream is supposed to consist of an infinite number of indefinitely thin layers. The variation in velocity from point to point of a cross-section may then be allowed for by giving the several layers different velocities based upon the law of fluid resistance between consecutive layers. 3. Density; Compressibility; Head; Continuity. The weight of ice per cubic foot at 23 F. is 57.2 Ibs.; "freshwater" " " " 39.2 F. is 62.425 Ibs.; " " "salt " " " " " 53 F. is 64 Ibs.; "fresh " " " " " 53 F. is 62.4 Ibs., or 1000 kilog. per cubic metre. The following table from the article on " Hydromechanics " in the Encyc. Brit, gives the density of water at different temperatures: FLOW THROUGH ORIFICES, OVER WEIRS, ETC. Temperature. Density. Weight in Lbs. per Cu. Ft. Temperature. Density. Weight in Lbs. per Cu. Ft. Cent. Fahr. Cent. Fahr. 32 .999884 62.417 20 68 .998272 62.316 I 33-8 .999941 62.420 22 71-6 .997839 62.289 2 35-6 .999982 62.423 24 75-2 .997380 62.261 3 37-4 I . 000004 62.424 26 78.8 .996879 62.229 4 39- 2 1.000013 62.425 28 82.4 .996344 62.196 5 4i 1.000003 62.424 30 86 995778 62. 161 6 42.8 .999983 62.423 35 95 .994690 62.093 7 44.6 999946 62.421 40 104 .992360 61.947 8 46.4 .999899 62.418 45 H3 .990380 61.823 9 48.2 999 8 37 62.414 50 122 .988210 61.688 10 50 .999760 62.409 55 131 .985830 61.540 ii 51.8 .999668 62.403 60 140 .983390 61.387 12 53-6 .999562 62.397 65 I 49 .980750 6l.222 13 55-4 .999443 62.389 70 158 977950 61.048 14 57-2 .999312 62.381 75 167 .974990 60.863 15 59 999!73 62.373 80 176 .971950 60.674 16 60.8 .999015 62.363 85 185 .968800 60.477 17 62.6 .998854 62.353 90 194 .965570 60.275 18 64.4 .998667 62.341 IOO 212 .958660 59.844 19 66.2 .998473 62.329 Fluids are sensibly compressed under heavy pressures, and the compression is proportional to the pressure up to about 1000 Ibs. (65 atmospheres) per square inch. Grassi's ex- periments indicate that the compressibility of water diminishes as the temperature increases. TABLE OF ELASTICITY OF VOLUME OF LIQUIDS. (Reduced from Grasses results.) Liquid. Elasticity of Volume. Temperature. Mercury . .. 717,000,000 o C. Water. ... j 42,000,000 1 45,900,000 o C. 18 C. Sea- water.. 52,900,000 Ether . j 1 6, P. 80,000 o C. \ 15,000,000 14 C. Alcohol. . .. ( 25.470,000 ( 23,380,000 7 . 3 o C. 13-1 C. Oil 44,090,000 N. B. The value for mercury is probably erroneous. If a volume Fof a fluid is compressed by an amount AV under an increase Ap of the pressure, then " 4 HYDRA ULICS. AV is called the cubical compression, and V -- is termed the elasticity of volume. This is sensibly constant. The vertical distance between the free surface of a mass of water and any datum plane is called the head with respect to that plane. If the water extends down to the level of the plane, a pressure/ is produced at that level, and the value of p r so long as the water is at rest, is given by the equation ^ = A+4, u -fir.J^ w being the. specific weight of the water and / the pressure at the free surface. Thus the pressure may be measured in terms of the head, and hence the expression "head due to pressure or pressure head." The mean value of the atmospheric pressure is 14.7 Ibs. per square inch. A , , ( is equivalent to A head Of a pressure of 2.3 ft. of fresh water I Ib. per sq. in. 2.25 ft. of salt water I Ib. per sq. in. About 34 ft. of fresh water 14.7 Ibs. per sq. in. " 33 ft. of salt " 14.7 Ibs. per sq. in. A head of water is a source of energy. A volume of water descending from an upper to a lower level may be employed to drive a machine which receives energy from the water and utilizes it again in overcoming the resistances of other machines doing useful work. Let Q cu. ft. of water per second fall through a vertical distance of 1i ft. Then the total power of the fall = wQIi ft. -Ibs. = h. p., w being the weight of the water in pounds per cubic foot. Let K be the proportion of the total power which is absorbed in overcoming frictional and other resistances. Then FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 5 the effective power of the fall = ze/(2^ ( r ~ ^0> an d the efficiency is i - AT. Imagine a bounding surface enclosing a space of invariable volume in the midst of a moving mass of fluid. The principle of continuity affirms that in any interval of time the flow into the space must be equal to the outflow during the same inter- val. Giving the inflow a positive and the outflow a negative sign, the principle may be expressed symbolically by = o. The continuity of a mass of water will be preserved so long as the pressure exceeds the tension of the air held in solution. It is on account of the pressure of this air that pumps cannot draw water to the full height of the water barometer, or about 34 ft. Generally speaking, the pressure at every point of a contin- uous fluid must be positive. A negative pressure is equivalent to a tension which will tend to break up the continuity pre- supposed by the formulae ; and should negative pressures result from the calculations, the inference would be that the latter .are based upon insufficient hypotheses. The pressure in water flowing through the air cannot at any point fall below the atmospheric pressure. There are cases, however, as in water flowing through a closed pipe (Art. 3, Chap. Ill), in which the pressure may fall below this limit and become almost nil. But there is then a danger of the air held in solution being set free, thus tending to interrupt the continuity of the flow, which may be wholly stopped if the air is present in sufficient volume. Consider a length of a canal or stream bounded by two normal sections of areas A lf A t , and let v lt v^ be the mean normal velocities of flow across these sections. Then by the principle of continuity and the velocities are inversely as the sectional areas. Again, assume that a moving mass of fluid consists of an O HYDRA ULICS. infinite number of stream-lines, and consider a portion of the mass bounded by stream-lines and by two planes of areas A lt AI at right angles to the direction of flow. If v^ , ^ 2 are the mean velocities of flow across the planes, V^AI = Q = VyA 9 if the fluid is incompressible. Assuming that the fluid is compressible, and that the mean specific weights at the two planes are w l and w 9 , then the weight of fluid flowing across A l is equal to the weight which flows across A^ , since the weight of fluid between the two planes remains constant. Hence 4. Bernouilli's Theorem. This theorem is based on the following assumptions : (1) That the fluid mass under consideration is a steadily moving stream made up of an infinite number of stream-lines whose paths in space are necessarily fixed. (2) That the velocities of consecutive stream-lines are not widely different, so that viscosity, or the frictional resistance between the stream-lines, is sufficiently small to be disregarded. (3) That the fluid is incompressible, so that there can be no internal zvork due to a change of volume. In any given stream-line let a portion AB, Fig. I, of the fluid move into the position A'B' in / seconds. B B' i' ' FIG. i. Let a l , p l , v l , z l be the normal sectional area, the intensity of the pressure, the velocity of flow, and the elevation above FLO W THROUGH ORIFICES, OVER WEIRS, ETC. / a datum plane ZZ of the fluid at A. Let tf a ,/ 3 , z> 2 , z^ denote similar quantities at B. Since the internal work is nil, the work done by external forces must be equivalent to the change of kinetic energy. Now the external work = the work done by gravity -f- the work done by pressure. But when the fluid AB passes into the position A'ft ', the work done by gravity is equivalent to the work done in the transference of the portion BB' , and therefore, t beng the time. the work dw by gr^^ty = wa^AA'-z^ wa^-BB' ' = wQt (*>-*,), since A A' = vj, BB' = vj, and a l v l = Q = a.^- Again, the work done by the pressures on the ends A and B The work done by the pressure on the surface of the stream- line between A and B is nil, since the pressure is at every point normal to the direction of motion. The change of kinetic energy = kinetic energy of A' B' kinetic energy of AB = kinetic energy of BB' kinetic energy of AA' , since the motion is steady, and there is therefore no change in the kinetic energy of the intermediate portion A' B. Thus, w v w V the change of kinetic energy = - a^BB'^- -- a. A A w Hence, equating the external worl{ and the change of kinetic energy, >Qt (*, - *,) + & (A ~ A) = & -- , 8 HYDRA ULICS, which may be written in the form w v? , w v? . . ,+/>, + - -y = ^,+A + --> ... (i) But A and B are arbitrarily chosen points, and therefore, at any point of a stream-line, the motion being steady and the viscosity nil, the gradual interchange of the energies due to head, pressure, and velocity is expressed by the equation w V* fa i W2 j_ p _L = wH, a constant ; / ... . (3) ~r r \ g 2 VJ/ +/ I Is, ** I V^ \M frb*' f - v Xm r - J V -i or n -iVj / / z being the elevation fef the v point above the datum line, / the pressure at the point, w the specific weight, and v the velocity of flow. This is Bernouilli's theorem. Thus the total constant energy of wH ft.-lbs. per cubic foot of fluid, or H ft.-lbs. per pound of fluid, is distributed uniformly along a stream-line, wH being made up of wz ft.-lbs. due to w z? head,/ ft.-lbs. due to pressure, -- ft.-lbs. due to velocity, and H being made up of z ft.-lbs. due to head, ft.-lbs. due v* to pressure, and ft.-lbs. due to velocity. Assuming that (a) the motion is steady, (ft) the frictional resistance may be disregarded, (c) the fluid is incompressible, Bernouilli's theorem may be applied to currents of finite size at any normal section, if the stream-lines across that section are sensibly rectilinear and parallel. There is then no interior work due to a change of volume, and the distribution of the pressure in the section under consideration will be the same as FLO W THROUGH ORIFICES, OVER WEIRS, ETC. 9 if the fluid were at rest, that is, in accordance with the hydro- static law. This is also true whether the flow takes place under atmospheric pressure only, or whether the fluid is wholly or partially confined by solid boundaries, as in pipes and canals, or whether the flow is through another medium already occu- pied by a volume of the fluid at rest or moving steadily in a parallel direction. In the last case there must necessarily be a lateral connection between the two fluids, but the pressure over the section must follow the hydrostatic law throughout the separate fluids, and there can be no sudden change of pressure at the surface of separation, as this would lead to an interruption of the continuity. The hypotheses, however, upon which these results are based are never exactly realized in actual experience, and the results can only be regarded as tentative. Further, they can only ap- ply to an indefinitely short length of the current, as the viscosity, -which is proportional to the surface of contact, would other- wise become too great to be disregarded. 5. Applications. If a glass tube, open at both ends, and called a piezometer (TrieCeiv, to press ; jterpor, a measure) is inserted vertically in the cur- rent, Fig. 2, at a point N, z ft. above the point O in the datum line, the water will rise in the tube to a height MN dependent upon the pressure at N. The effect of the eddy motion produced at N by obstructing the stream- lines may be diminished by mak- ing this end of the tube parallel to the direction of flow. Neglect- ing altogether the effect of the o eddies, and taking/ to be the in- FlG - 2 - tensity of the pressure at TV, and/ the intensity of the atmos- pheric pressure, then, w w 10 and therefore HYDRA ULICS. w w = ON + MN + - 1 w = Q M + -. 1 w (5) The locus of all such points as M is often designated " the line of hydraulic gradient," or the " virtual slope," terms also- used when friction is taken into account. Let the two piezometers AB, CD, Fig. 3, be inserted in the current at any two points B and D, z^ ft., and z % ft. respect- ively above the points E and F in the datum line. FIG. 3. Let /, be the intensity of the pressure at B in pounds per square foot, / 2 that at D, and let the water rise in these tubes to the heights BA, DC. Then w and therefore = z l +, and l w w w + - + =^-^=^' . . (6) the line AG being parallel to the datum line. FLOW THROUGH ORIFICES, OVER WEIRS, ETC. II Thus , (z l -|- J Ls 2 + J is equal to the fall of the free surface level between the points B and D. Let v l , 7' 2 be the velocities of flow at B and D. Then by Bernoulli's theorem W 2g W 2g and therefore the fall of free surface level between B and D (7) W 2g Equation (7) may also be written in the form V ? V ? , I , Pi\ ( . Pl\ V ? . rr SQ\ - = r (Zi H J l*i H } = r k . (8) 2g 2g ^ W' * W' Zg so that the velocity at D is equal to that acquired by a body with an initial velocity v l falling freely through the vertical distance CG. Froude illustrated Bernouilli's theorem experimentally by means of a tube of varying section, Fig. 4, conveying' a current FIG. 4. between two cisterns. The pressure at different points along the tube is measured by piezometers, and it is found that the 12 HYDRAULICS. water stands higher and the pressure is therefore greater, where the cross-section is larger and the velocity consequently less. If the section of the throat at A is such that the velocity is that acquired by a body falling freely through the vertical dis- tance h between A and the surface level of the water in the cistern, and if / be the pressure at A, and z the elevation of A above datum, then, neglecting friction, W 2g W But v* = 2gh, and therefore / = p Q , so that the pressure at A is that due to atmospheric pressure only. Thus, a portion of the pipe in the neighborhood of A may be removed, as in the throat of the injector. Again, let the cross-section in the throat at B be less than that at A. The pressure at B will be less than the atmospheric pressure, and a column of water will be lifted up in the curved piezometer to a height k' . Let tf , , z l ,p l , v l be the sectional area, elevation above datum, pressure, and velocity at B. Let # 3 , z^ ,pi , z> a be similar symbols at E. Then I J ,, + A + ^+A ^ = , l + A_*' + !i 2 . (9) V W 2g W ' 2g W ' 2g Put //, = #,-[- -, the height above datum to which the w water is observed to rise in the piezometer inserted at E, and also let #;=*, + A - h'. Then w since ap^ = a l v l , # a being the sectional area at E. Therefore ft., a, FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 1 3 an equation giving the theoretical velocity of flow at the throat B. Hence the theoretical quantity of flow across the section at B is - a (10) This is the principle of the Venturi water-meter and also of the aspirator. The actual quantity of flow is found by multiplying equa- tion (10) by a coefficient C whose value is to be determined by experiment. If the pressure at E is positive, then //, is merely the height to which the water is observed to rise in an ordinary piezometer inserted at E. Again, Froude also points out that when any number of combinations of enlargements and contractions occur in a pipe, the pressures on the converging and diverging portions of the pipe will balance each other if the sectional areas and directions of the ends are the same. 6. Orifice in a Thin Plate. If an opening is made in the wall or bottom of a tank containing water, the fluid particles FIG. 6. FIG. 7. immediately move towards the opening, and arrive there with a velocity depending upon its depth below the free surface. The opening is termed an " orifice in a thin plate " when the water springs clear from the inner edge, and escapes without again touching the sides of the orifice. This occurs when the UNIVERSITY 14 HYDRA ULICS* bounding surface is changed to a sharp edge, as in Fig. 5, and also when the ratio of the thickness of the bounding surface to the least transverse dimension of the orifice does not exceed a certain amount which is usually fixed at unity, as in Figs. 6 and 7. Owing to the inertia acquired by the fluid filaments there will be no sudden change in their direction at the edge of the orifice, and they will continue to converge to a point a little in front of the orifice, where the jet is observed to contract to the smallest section. This portion of the jet is called the vena contracta or contracted vein, and the fluid filaments flow across the minimum section in sensibly parallel lines, so that here, if the motion is steady, Bernouilli's theorem is appli- c cable. The dimensions of the contracted section and F its distance from the orifice depend upon the form and dimensions of the orifice and upon the head of water over the orifice. Let Fig. 8 represent the portion of the jet be- tween a circular orifice of diameter AB and the contracted section of diameter CD, EF being the distance between AB and CD. Then, taking the average results of a number of observations, it is found that AB, CD and EF are in the ratios of 100 to 80 to 50. Thus the areas of the contracted section and of the orifice are in the ratio of 16 to 25, and, generally speaking, this is assumed to be the ratio whatever may be the form of the orifice. 7. Torricelli's Theorem. Let Fig. 9 represent a jet issu- ing from a thin-plate orifice -in the side of a vessel containing water kept at a constant level AB. Let XXbz the datum line, J/A^the contracted section, and consider any stream-line mn, m being in a region where the velocity is sensibly zero, and n in the contracted section. Then by Bernouilli's theorem, the motion being steady, FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 1 5 /, /, being the pressures at n and ;, and z t 2 l their elevations above datum. Hence (2) FIG. 9. If the flow is into the -atmosphere, / the atmospheric pressure = / , and p, = w.Om +/., O being the point in which the vertical through m intersects the free surface. Thus, = z . z + Om = 2 g (3) h being the depth of n below the free surface. The result given by equation (3) was first deduced by Tor- ricelli. The depth below the free surface is very nearly the same for all points of the contracted vein, and the value of v as given by (3) is taken to be the theoretical mean velocity of flow across the contracted section. Equation (3) is equivalent to the statement that when the orifice is opened the hydrostatic energy of the water, viz., // ft.- 7' 2 Ibs. per pound, is converted into the kinetic energy of 10 i HYDRA ULICS. ft.-lbs. per pound. Thus, if the jet is directed vertically upwards, it will very nearly rise to the level of the free surface, and would reach that level were it not for air resistance, or for viscosity, or for friction against the sides of the orifice, or for a combination of these retarding causes. If the jet issues in any other direction, it describes a para- bolic arc of which the directrix lies in the free surface. Let OTV, Fig. 10, be such a jet, its direction at the orifice FIG. 10. at O making an angle a with the vertical. With a properly formed orifice a greater or less length of the jet will have the appearance of a glass rod, and if this portion were suddenly solidified and supported at the ends, it would stand as an arch without any shearing stress in normal sections. Again, the horizontal component of the velocity of flow at any point of the jet is constant (= v sin or), so that, for the unbroken portion of the jet, equidistant vertical planes will intercept equal amounts of water, and the height of the C. G. of the jet above the horizontal line O V will be two thirds of the height of the jet. 8. Efflux through an Orifice in the Bottom or in the Side of a Vessel in Motion. If a vessel containing water zit. deep ascend or descend vertically with an acceleration /, the press- ure/ at the bottom is given by the equation w -*/ = p - p - wz y FLOW THROUGH ORIFICES, OVER WEIRS, ETC. being the atmospheric pressure. Therefore /-A w =( 4 If now an orifice is opened at the bottom, the velocity of efflux v is still taken as due to the head of the pressure /, and therefore by Torricelli's Theorem Let W, be the weight of the vessel and water, and let the vessel be connected with a counterpoise of weight W^ by meansof a rope passing over a pulley. Then by Newton's second law of motion, and neglect- ing pulley friction, g~ W, W, W.+ Wt' T being the tension of the rope. Next let a cylindrical vessel, Fig. 12, of radius r and containing water, rotate with an angular veloc- ity oo about its axis. The surface of the water assumes the form of a paraboloid of which the latus 2fT rectum is ^. If an orifice is made GO at Q in the side of the vessel, the water will flow out with a velocity v due to the head of pressure at FIG. 12. the orifice. This head is PQ, and PQ = ON z = CA) z being the vertical distance OM between the orifice and the vertex of the paraboloid. Hence by Torricelli's theorem i8 HYDRA ULICS. *' or 9. Application to the Flow through a Frictionless Pipe of Gradually Changing Section (Fig. 13). Let the pipe be supplied from a mass of water of which the free surface is H ft. above datum. Let a l9 pv v v be the sectional area, pressure, and velocity of flow at any point A, z l ft. above datum and h^ ft. below the free surface. Let# a ,/ 3 , z/ 3 be similar symbols for a second point B, ^ ft. above datum and h^ ft. below the free surface. FIG. 13. Then by the condition of continuity a&i = a t v t , and by Torricelli's theorem 2g /J t IV FLOW THROUGH ORIFICES, OVER WEIRS, ETC. and 2g ~ *" W Hence so that Bernouilli's theorem, viz., + - + 2 = & + = a constant, 2- ' w ' ' w holds true for the assumed conditions. 10. Hydraulic Resistances (a) Coefficient of Velocity. In reality, the velocity v at the vena contracta is a little less than V2gh (Art. 7, eq. 3) and the ratio of v to V ' 2gh is called the coefficient of velocity, and may be denoted by c vt so that v = c v Again, the equations for the velocity of discharge in the case of moving vessels now become 2g and A mean value of c v for well-formed simple orifices is .974. An easy method of determining the value of c v , experi- mentally, may be indicated by reference to the jet represented in Fig. 10, p. 1 6. Measure the vertical and horizontal distances from the orifice of any two points A, B in the jet. Let jj>,, x^ denote the co-ordinates of A. Let y v x^ denote the co-ordinates of B. Then if t l is the time occupied by a fluid particle in moving from the orifice to A, and t^ the time from the orifice to B, 2O HYDRA ULICS. v sin a . t l ; j^ = v cos <* . /, -- '/ 1 2 ; ^ = v sin a . / 3 ; J 2 = ^ cos ar . / 3 -- gt*. x ,=^ cot a- , . a , 2 z; 2 sin 2 a P- = x cot a - 2 2 v* sin 3 a' By means of the two last equations and 2 sin a (x l cot a yj so that *t ' Hence /> , and 4^ sin 2 a (x l cot a y^ ' and since the values of x^ y v x y^ are known, equation I will give the value of a, and equation 2 the value of c v . Note. If the jet issues from the orifice horizontally, a = 90, and the last equation becomes so that the position of one point only relatively to the orifice need be observed. (b) Coefficient of Resistance. Let h v be the head required to produce the velocity v. Let h r be the head required to overcome the frictional resistance. Then h, the total head, h v -\- h r = h v (i -}- c r ), where h r = t r h v . FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 21 c r is termed the coefficient of resistance, and is approxi- mately constant for varying heads with simple sharp-edged orifices. Again, Hence and therefore so that c r can be found when c v is known, and vice versa. (c) Coefficient of Contraction. The ratio of the area a of the vena contracta to the area A of the orifice is called the co- efficient of contraction, and may be denoted by c c . The value of c c must be determined in each case, but in sharp-edged orifices an average value of c c , as already pointed out, is - = .64. Ce o JS ' >^:s >C c H3 'Cc75 > Q s^ S "^ ."-^ "=-5 ^- *c5 rfj J3 ^ S rC ^ 5 ^ 3 a ^ 5 *+-( 6 "o 9 l.L 1'; || ^"fe III i crd i^s Hi s 3 1*1 rt-o ca u 5-0 4) !.- S'O C w C/3 F 4;C75^ I K Twice r ;^H Rectangle with Vertical Sides equal to Four Times uppermost. the Width. Head in Feet. T S T S T S T S T S I 678 620 657 631 643 627 662 6 4 688 6 7 I 2 618 613 646 623 6 3 I 621 643 629 655 657 4 610 605 628 616 620 615 6 3 I 620 642 643 6 607 601 628 613 615 612 62 7 616 634 636 8 606 60 1 621 610 612 609 624 613 631 6 3 2 10 fo 4 600 6l8 608 6I 3 608 621 613 629 629 12 663 598 '6l 7 607 6n 606 621 611 626 627 14 602 598 6l 7 607 610 606 620 610 623 625 16 602 598 616 606 609 606 619 609 622 625 18 60 1 597 6i5 605 607 605 618 608 622 623 20 60 1 597 615 605 607 604 618 608 621 622 T FLO W THROUGH ORIFICES, OVER WEIRS, ETC. 2$ The jet springs clear from the orifice in all cases repre- sented in Tables A and B. The following inferences may be drawn from an inspection of Tables A and B : (1) The coefficient of discharge diminishes as the head in- creases, but at a diminishing rate. (2) The coefficients for the thick-plate orifices are in all cases greater than the corresponding coefficients for sharp- edged orifices, excepting in the case of the longest rectangular orifice in Table B. Under a head of I ft. the coefficient of discharge for this orifice still exceeds that of the same orifice with sharp edge, but for heads exceeding I ft. the coefficient seems to be a little less, but is practically the same. It may be noted that the thickness of the plate is 2.56 times the width of the orifice, and the contraction for the thick-plate orifice is consequently increased. (3) The coefficient for rectangular orifices seems to be practically the same whether the longest side is vertical or horizontal. (4) The coefficient increases with the area of the orifice, excepting when the head is very small. The coefficient for orifices of small area then rapidly increases, as shown in Table B. (5) With rectangular orifices the coefficient increases as the width of the orifice diminishes, i.e., as the orifice becomes more elongated. The two last results are in accordance with similar results deduced by Weisbach, Buff, and others. The coefficient of discharge is modified when the edges of the orifices are not sharp, but have a sensible thickness, and the formula giving the discharge may be written Q = cA J^H> H being the depth of the axis of the orifice below the free surface. II. Miner's Inch. The miner's inch is a term applied to the flow of water through a standard vertical aperture, one square inch in section, under an average head of 6 inches. 26 HYDRA ULICS. Taking c = .62, the flow = Q = .62 A = - 62 x = i cu. ft. per minute, approximately. The term is more or less indefinite, as the different companies in disposing of water to their customers do not always use the FIG. 19. same head, and the flow is thus found to vary from 1.36 to 1.73 cu. ft. for each square inch of aperture. The aperture is usually 2 in. deep and may be of any re- quired width, Fig. 19. The upper and lower edges of the aperture are formed by ij-in. planks, the lower edge being 2 in. above the bottom of the channel, and the plank forming the upper edge being 5 to 5^ in. deep, so that the head over the centre of the aperture is from 6 to 6^ inches. 12. Energy and Momentum of the Jet. The energy of the jet = wav ft.-lbs. per second wav ft.-lbs. per second s. ^ *< FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 2? = wavhc* ft.-lbs. per second wavhc* , N = - - h. p. (horse-power) p (= wh} being the hydrostatic pressure due to the head h. w The momentum of the jet ~ - av . v wa - = 2wakc* < o and this is equal to the pressure in pounds produced by the jet against a fixed plane perpendicular to its direction. Neglect- ing c v *, the thrust is double the hydrostatic pressure due to the head h. 13. Inversion of the Jet. The phenomenon of the inver- sion of the jet was first noticed by Bidone, and has been subse- quently investigated by Poncelet, Lesbros, Magnus, Lord Rayleigh, the author, and others. When a jet issues from an orifice in a vertical surface, the sections of the jet at points along its path assume singular forms dependent upon the nature of the orifice. Figs. 20 to 27 are from photographs (taken from the same point) of jets issuing under the same head, viz., 12 ft., from orifices of different forms and sizes. The dimensions of these jets are comparable with the jets shown by Figs. 20 and 21, which are issuing from circular orifices of I in. and J in. diameter, respectively. With a square orifice, Fig. 22 (side = I in.), Fig. 23 (side = .443 in.), and Fig. 24 (side = .25 in.), the section is a star of four sheets at right angles to the sides. With a triangular orifice, Fig. 25 (side = .676 in.), the sec- tion is a star of three sheets at right angles to the sides. In general, with a polygonal orifice of n sides the section will be a star of n sheets at right angles to the sides. Fig. 26 is a jet from a rectangular orifice (J in. X J in.), its section near the orifice being a star of four sheets. Fig. 27 is a jet from a semi-circular orifice (diar. .388 in.), FIG. 20. FIG. 21. FIG. 22. FIG. 23. FIG. 24. FIG. 25. Fro. 26. FIG. 27. FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 2Q the section near the orifice being a rounded boundary and a single sheet at right angles to the diameter. The changes in the form of the jet are doubtless due to the mutual action between the fluid particles. A filament issuing horizontally and freely at B, Fig. 28, has a velocity 2g . AB, and FIG. 28. describes a certain parabola BD. A filament issuing horizon- tally and freely at a lower level C has a velocity 2g . AC, and describes a parabola CD of less curvature than BD. Now the two filaments cannot pass simultaneously through the point of intersection D, and must necessarily press upon each other- They are thus deviated out of their natural paths, and the jet spreads out into sheets, as described above. If the orifice is small and the head not large, the jet, on leaving the contracted section at the orifice, spreads out into sheets, and then diminishes to a contracted section similar to the first, after which it again spreads out into sheets, bisect- ing the angles between the first set of sheets, and again dimin- ishes to a contracted section. This action is repeated so long as the jet remains unbroken. 14. Emptying and Filling a Canal Lock. When the head varies, as in filling or emptying a reservoir or a lock, in filling a vessel by means of an orifice underwater, or in empty- ing water out of a vessel through a spout, Torricelli's theorem is still employed. If the lock or vessel is to be filled, Fig, 29, let X sq. ft. be the area of the water-surface when it is x ft. below the surface of the outside water. 3 HYDRA ULICS. If the lock or vessel is to be emptied, Fig. 30, then X sq. ft. is the area of the water-surface when it is x ft. above the orifice. FIG. 29. FIG 30. In each case JIT ft. is the effective head over the orifice, and is the head under which the flow takes place. In the time dt the water-surface in the lock or vessel will rise or fall by an amount dx. Then A .dx = quantity which has entered the lock = cA , v be the intensity of the atmospheric pressure and the velocity at MN. Let p^ , v l be the mean intensity of pressure and the veloc- ity at AB. Let P be the mean intensity of the pressure over the annu- lar surface EF, GH. Let # , 2 l be the elevations above datum of the C. G.s of the sections MN and AB. FLO W THROUGH ORIFICES, OVER WEIRS, ETC. 39 Then ow.fo - *.) + P& - P(a, - a,) p. a, impulse in direction of motion change of momentum in same direction in a unit of time difference of the momenta of the fluid masses MNN'M' and ABB' A' ~(av* - a,*,")- Assuming that P = p iy the last equation becomes IV wa,(2 l - *.) + a,(p, - p.) = (av* a,v*). . . (i) o By Bernoulli's theorem, A *>? p. v* and therefore W 2g Now s, z is very small and may be disregarded without sensible error, and then by eqs. (i) and (2) v* v? _ p l p _ i av* 2g w ~ g a, Hence 2 _ a ~ av* av ~~ aa a*av* ~~ a since a 9 v t = av. If the sectional area # a of the pipe is very large as compared with a, so that -- may be disregarded without sensible error, then = -, and therefore the coefficient of contraction a, a = = -, as before. HYDRA ULICS. (c] Cylindrical Mouthpiece. Whe-n water issues from a cylindrical mouthpiece (see Fig. 35) at least two to two and one half diameters in length, the jet issues full bore or without contraction at the point of dis- charge. If A be the sectional area of the mouthpiece, h the depth of its axis below the water-surface, and Q the amount of the dis- charge. Then experiment shows that Q = .S2A |/^. . (i) The coefficient .82 is the pro- duct of the coefficients of veloc- ity and contraction, but the co- efficient of contraction is unity, FIG 35. and therefore the coefficient of velocity is .82. Now the mean coefficient of velocity in the case of a simple sharp- edged orifice is .947, and the difference between .947 and .82 cannot be wholly accounted for by frictional resistances, but is in part due to a loss of head. In fact, the water as it clears the inner edge of the mouthpiece converges to a minimum sec- tion MN of area a and then swells out until at M' N' it again fills the mouthpiece. Energy is wasted in eddy motions between MN and M'N', where the action is similar to that which occurs at an abrupt change of section. Let /, v be the intensity of the pressure and the mean velocity of flow at the point of discharge. Let />, , v 1 be similar symbols for the contracted section MN. Let /> be the intensity of the atmospheric pressure. Remembering that is the loss of head "due to shock " between MN and M'N, then by Bernoulli's theorem (2) , A = A v = /_, v " W ~ W ' 2g W*2f 2g FLO W THROUGH ORIFICES, OVER WEIRS, ETC. 4! Hence W 2g and (I-)'! where 4. = coefficient of contraction = = . Therefore an equation giving the velocity of flow at the point of discharge. If the discharge is into the atmosphere,^ =/ and equation 4 becomes * ' where / \. 1--IJ (5) ^.=-+(7,-.)' If c c = .62, then c v = .85, while experiment gives .82 as the value of c v . The small difference between .85 and .82 is probably due to frictional resistance. The value .82 for c v makes c c approximately .617. Again, the discharge from a simple sharp-edged orifice of same sectional area as the mouthpiece is .62A V2gh, or more than 24 per cent less than the discharge from the cylindrical mouthpiece.. 42 HYDRA ULICS. The loss of head between MN and M'N' (by eqs. 5 and 6) = h(i O = h X .3276 = , approximately. Thus the effective head is only \h, instead of h. By eq. 3 the difference between the pressure-heads at MN and at the point of discharge = . h = // - w = ^, very nearly. Now if one end of a tube is inserted in the mouthpiece at the contracted section (Fig. 35) and the other end immersed in a vessel of water, the water will at once rise to a height /z, in the tube, showing that the pressure at the contracted section is less than that due to the atmosphere. By careful measurement it is found that h^ is very nearly equal to \h y which verifies the theory. (d] Divergent Mouthpiece. Suppose that for the cylin- drical mouthpiece in (c) there I is substituted a divergent F IG - 36. mouthpiece of the exact form of the issuing jet (see Fig. 36), Then FLO W THROUGH ORIFICES, OVER WEIRS, ETC. 43 (1) The mouthpiece will run full bore. (2) There will be no loss of head between the minimum section MN and the plane of discharge AB, as there is now no abrupt change of section. Hence by Bernoulli's theorem, and retaining the same symbols as in (c), =+ = + (.) W W 2g W 2g If the discharge is into the atmosphere,/ = / , and therefore v* = 2gh\ .... '. . . . (2) or introducing a coefficient c v (= .98, nearly, for a smooth well-formed mouthpiece), and the discharge is . , ;, ._., . (4) From the last equation it would appear as if the discharge would increase indefinitely with A, but this is manifestly impossible. In fact, by eq. I, the flow being into the air, and taking , (c) W W 2g\V* since av^ = Av. But/, cannot be negative, and therefore so that a '\ wk+ l (7) gives a maximum limit for the ratio of A to a. 44 HYDRA ULICS. Now = 34 feet very nearly, and the last equation may be written By eqs. 4 and 7, (9) which is also the expression for the discharge through the minimum section a into a vacuum. If, however, the sectional areas of the mouthpiece at the point of discharge and at the throat are in the ratio of A to a, as given by eq. 7, it is found that the full-bore flow will be in- terrupted either by the disengagement of air, or by any slight disturbance, as, for example, a slight blow on the mouthpiece, and hence, in practice, it is usual to make the ratio of A to a sensibly less than that given by eq. 7. (e) Convergent Mouthpiece. With a convergent mouth- piece (Fig. 37) two points are to be noted : (i) There is a contraction within the mouthpiece, followed by a swelling out of the jet until it again fills the mouthpiece. FIG. 37. Thus, as in the case of cylindrical mouthpieces, there is a " loss of head " between the contracted section and the point FLO W THROUGH ORIFICES, OVER WEIRS, ETC. 45 of discharge, and also a consequent diminution in the velocity of discharge. (2) There is a second contraction outside the mouthpiece due to the convergence of the fluid filaments. The mean velocity of flow (V) across the section is v' = C,' C v r being the coefficient of velocity and h the effective head above the centre of the section. Also, the area of this section = C C 'X area of mouthpiece at point of discharge = C C '.A, Cc being the coefficient of contraction. Hence the discharge Q is given by Q = C v f C c 'A = C'A C'(= C v 'C c f ) being the coefficient of discharge. The coefficients C v ' and C e ' depend upon the angle of con- vergence, and Castel found that a convergence of 13 24' gave a maximum discharge through a mouthpiece 2-6 diameters in length, the smallest diameter being .05085 foot. TABLE GIVING CASTEL'S RESULTS. Angles of Convergence. C c ' <5 <5 HYDRA ULICS. If the second point is on the axis of revolution, then r a = o, and the last equation becomes W 2P~ 1 hus the free surface of the pressure columns is evidently a paraboloid of revolution with its vertex at O, as in Fig. 40. A compound vortex is produced by the combination of a central forced vor- tex with a free circular vortex, the free surface being formed by the revolution of a Barlow curve and a parabola. For example, the fan of a centri- fugal pump draws the water into a forced FIG. 40. vortex and delivers it as a free spiral vor- tex into a whirlpool-chamber (Chap. VII.). In this chamber there is thus a gain of pressure-head, and the water is therefore enabled to rise to a corresponding addi- tional height. James Thomson adopted the theory of the corn- compound vortex as the principle of the action of his vortex turbine. 20. Large Orifices in Vertical Plane Surfaces. The issuing jet is approximately of the same sectional form as the orifice, and the fluid filaments converge to a minimum section as in the case of simple sharp-edged orifices. (a) Rectangular Orifice (Fig. 41). Let E, F be the upper and lower edges of a large rectangular orifice of breadth B, and let H^ , H^ be the depths of E and F, respectively, below the free surface at A. If u be the velocity with which the water reaches the orifice, then H = -- is the fall of free surface which must have been expended in producing the velocity u. Hence, H l -\- H and H^ + H are the true depths of the edges E and F below the surface of still water. Let A/TV be the minimum or contracted section, and assume that it is a rectangle of breadth b. FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 5 1 Let h l , h^ be the depths of M and N, respectively, below the free surface at A. Then h l -\- H, h^ -\- H are the true depths of M and N below the surface of still water. First, let the flow be into the air, the orifice being clear above the tail-water level. Consider a lamina of the fluid at the section MN of the 1 j ? ;- "' "==:iEi5= 4 T 1 1 ! i- f : ~ f= 1 i J 1 1 f ' r i E f | i V| f 1 . 2 . M-1-- ; H L.-i__ J L..[- i i 1 FIG. 41. width of the section and between the depths x and ^r + dx below the surface of still water. The elementary discharge dq, in this lamina, is dq = bdx Now the ratio jy is constant in a triangular notch and varies **\ in a rectangular notch. Hence Thomson inferred and proved by experiment that the value of c is more constant for trian- gular than for rectangular notches, so that a triangular notch is more suitable for accurate measurements. Example. A sharp-edged triangular notch is opened in the side of a reservoir, and the water flows out until the free- surface level sinks to the bottom of the notch. The discharge in the short interval dt, when the depth of water in the notch is x ft., = cmx mx being the width of the free surface corresponding to the depth x, and m a coefficient depending upon the angle of the notch. Again, S . dx is the quantity of the water which leaves the reservoir in the same time dt, S being the horizontal sectional area of the reservoir corresponding to the depth x. Hence 4 ,/ and therefore \/2gcmx l dt = Sdx, \f2gcmdt = Sx~*dx, so that the time in which the free surface sinks to the required level x 15 c = -- 7^ / 4 V2gcmJ Q X being the initial depth. 58 HYDRAULICS. If 5 is constant, then the time = 23. Broad-crested Weir. Let Fig. 46 represent a stream flowing over a broad-crested weir. On the up-stream side the FIG. 46. free surface falls from A to B. For a distance BD on the crest the fluid filaments are sensibly rectilinear and parallel; the inner edge of the crest is rounded so as to prevent crest con- traction. Consider a filament ab, the point a being taken in a part of the stream where the velocity of flow is so small that it may be disregarded without sensible error. Let A be the thickness MN of the stream at b. Let the horizontal plane through N be the datum plane. Let # z be the depths below the free surface of a and b. Let h l be the elevation of a above datum. Let/ , /,, p be the atmospheric pressure and the pressures at a and b. Let v be the velocity of flow at b. Then, by Bernoulli's theorem, W W 2g FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 59 But = * i + and = , + ; W WWW therefore and hence _ = k, + z l - A = H, - A, //", being the depth of the crest of the weir below the surface of still water. Thus, if B be the width of the weir, the discharge Q is (16) From this equation it appears that Q is nil both when A = o and when A = //,. Hence there must be some value of A between o and //, for which Q is a maximum. This value may be found by putting and the expression for the discharge becomes , = .3855 V^rf, . ' . (17) which is the maximum discharge for the given conditions. Experiment shows that the more correct value for the dis- charge is . . . (18) 60 HYDRAULICS. This formula agrees with the ordinary expression for the discharge over a weir as given by equation u, if c = .525. It might be inferred that for broad-crested weirs and large masonry sluice-openings the discharge should be determined by means of equation 18 rather than by the ordinary weir formula, viz., equation n. It must be remembered, however, that in deducing equa- tion 17 frictional resistances have been disregarded and the gratuitous assumption has been made that the stream adjusts itself to a thickness / which will give a maximum discharge. The theory is therefore incomplete. FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 6 1 EXAMPLES. I. A frictionless pipe gradually contracts from a 6-in. diameter at A to a 3-in. diameter at B, the rise from A to being 2 ft. If the de- livery is i cubic foot per second, find the difference of pressure between the two points A and B. Ans. 500 Ibs. per sq. ft. . 2. In a frictionless horizontal pipe discharging 10 cubic feet.of water per second, the diameter gradually changes from 4 in. at a point A to i/C 6 in. at a point B. The pressure at the point ^5*is 100 Ibs. per square inch ; find the pressure at the point A. Ans. 4118 Ibs. per sq. ft. 3. A ^-in. horizontal pipe is gradually reduced in diameter to -J- in. and then gradually expanded again to its mouth, where it is open to the atmosphere. Determine the maximum quantity of water which can be forced through the pipe (a) when the diameter of the mouth is \ in., (b) when the diameter is f in. Also determine the corresponding velocities at the throat and the total heads (neglect friction, which, however, is very considerable). Ans. (a) .24 cub. ft. per min.; 46.7 ft. per. sec. (b) .239 cub. ft. per. min.; 46.66 ft. per sec. 4. A short horizontal pipe A BC connecting two reservoirs gradually contracts in diameter from i inch at A to inch at B and then enlarges to i inch again at C. If the height of the water in the reservoir over C be 12 inches, determine the maximum flow through the pipe and sketch the curve of pressures. Also obtain an equation for this curve, assum- ing the rates of contraction and expansion of the pipe to be equal and uniform. Ans. 3.75 cub. ft. per min. 5. The pipe DE in the figure is gradually contracted in diameter from D to E, where it is enclosed in another pipe ABC, expanding from B towards A and C\ at C it is open to the atmosphere and at A it is connected with D/ a reservoir R ; the water surface in R being h' below the horizontal axis of DE. If the velocity in DE at E be v and the velocity in AB at B be F, what will be the common velocity after uniting? Explain what becomes of the energy lost in im- ^ pact. If the diameters at E, B, arid C are \ in., f in., and i in., the distance between the outside of E and inside 62 HYDRA ULICS. of B being T \ inch, find the ratio of the quantity pumped from R to the flow through DE. 6. A 3-in. pipe gradually expands to a bell-mouth ; if the total head, //, be 40 ft., find the greatest diameter of the mouth at which it will run full when open to the atmosphere. Compare the discharge from this pipe with the discharge when the pipe is not expanded at the mouth. Ans. 4.8 in.; discharge is 18.63 cu b. ft- P er minute with bell- mouth and 7.337 cub. ft. per minute without bell-mouth. 7. The pressure in a 12- in. pipe at A is 50 Ibs.; the pipe then en- ^ larges to a i5-in. pipe at B, the rise from A to B being 3 ft.; the dis- / charge is Q cubic feet per minute. Find the pressure at^; also find the pressure at a point C, the rise from B to C being 6 ft, (6637.5 + T Ans. 6637.5 + lbs ' er sc - ft ' 8. Two equal pipes lead, one from the steam-space, the other from the water-space of a boiler at pressure/; Ss is the density of the steam and S w that of the water. Assuming Torricelli's theory to hold for rate of efflux of steam and water, show that vel. of steam-jet _ */ _ quantity of water-jet _ energy of steam-jet vel. of water-jet ~~ * S s ~ quantity of steam-jet ~~ energy of water-jet, and that the momentum of each jet is the same. 9. Find the head required to give i cub. ft. of water per second through an orifice of 2 square inches area, the coefficient of discharge [/ being .625. (g = 32.) Ans. 206 ft. 10. The area of an orifice in a/fcnin plate was 36.3 square centimetres, the discharge under a head 0^/^396 metres was found to be .01825 cubic metre per second, and the velocity of flow at the contracted section, as \^/ determined by measurements of the axis of the jet, was 7.98 metres per second. Find the coefficients of velocity, contraction, discharge, and re- sistance. (^ = 9.81.) Ans. .978; .631; .617; .045. 11. The piston of a 12-in. cylinder containing saltwater is pressed down under a force of 3000 lbs. Find the velocity of efflux and the \J volume of discharge at the end of the cylinder through a well-rounded i -in. orifice. Also find the power exerted. Ans. 60.373 ft. per sec.; .1691 cub. ft. per sec.; 1.166 H. P. 12. In the condenser of a marine engine there is a vacuum of 26^ in. of mercury ; the injection orifices are 6 ft. below the sea-level. With what velocity will the injection-water enter the condenser? (Neglect re- sistance.) Ans. 25.3 ft. per sec. 13. Water in the feed-pipe of a steam-engine stands 12 ft. above the FLO W THROUGH ORIFICES, OVER WEIRS, ETC. 63 surface of the water in the boiler ; the pressure per sq. in. of the steam is 20 Ibs., of the atmosphere 15 Ibs. Find the velocity with which the water enters the boiler. Ans. 5.376 ft. per sec. 14. The injection orifice of a jet condenser is 5 ft. below sea-level and vacuum = 27 in. of mercury. Find velocity of water entering con- denser, supposing three fourths of the head lost by frictional resistance. Ans. 23.86 ft. per sec. 15. A vessel containing water is placed on scales and weighed. How will the weight be affected by opening a small orifice in the bottom of the vessel ? 1 6. Water is supplied by a scoop to a locomotive tender at 7 ft. above trough. Find lowest speed of train at which the operation is possible. Ans. 14.44 miles per hour. Also find the velocity of delivery when train travels at 40 miles per hour, assuming half the head lost by frictional resistance. Ans. 35.68 ft. per. sec. 17. The head in a prismatic vessel at the instant of opening an orifice was 6 ft. and at closing it had decreased to 5 ft. Determine the mean constant head h at which, in the same time, the orifice would discharge the same volume of water. Ans. 5.434 ft. 18. A prismatic vessel 5.747 in. in diameter has an orifice of .2 in. diam. at the bottom; the surface sinks from 16 in. to 12 in. in 53 seconds. Find the coefficient of discharge. Ans. .6. 19. A prismatic basin with a horizontal sectional area of 9 sq. ft. has an orifice of .09 sq. ft. at the bottom ; it is filled to a depth of 6 ft. above the centre of the orifice. Find the time required for the surface to sink 2 ft., 3^ ft., 5 ft. Ans. 260 sec.; 502 sec.; 838 sec. 20. The water in a cylindrical cistern of 144 sq. in. sectional area is 16 ft. deep. Upon opening an orifice of I sq. in. in the bottom the water fell 7 ft. in i minute. Find the coefficient of discharge. The co- efficient of contraction being .625, find the coefficients of velocity and resistance. Ans. .6 ; .96 ; 0.85. 21. How long will it take to fill a paraboloidal vessel up to the level of the outside surface through a hole in the bottom 2 ft. under water? (g = 32 and c = .625.) 1 76 |/2 B Ans. j, B being the parameter of the parabola and A the sectional area of the orifice. 22. How long will it take to fill a spherical Vessel of radius r up to the level of the outside surface through a hole of area A at bottom 2 ft. under water ? (g = 32 and c = .625.) Ans ' 64 HYDRAULICS. 23. A vessel full of water weighs 350 Ibs. and is raised vertically by means of a weight of 450 Ibs. Find the velocity of efflux through an orifice in the bottom, the head being 4 ft. Ans. 17.02 ft. per sec. 24. A vessel full of water makes loo-revols. per min. Find the velocity of efflux through an orifice 2 ft. below the surface of the water at the centre. Ans. 33.4 ft. per sec. What will be the velocity if the vessel is at rest ? Ans. 1 1. 35 ft. per sec. 25. The jet from a circular sharp-edged orifice, in. in diameter, un- der a head of 18 ft., strikes a point distant 5 ft. horizontally and 4.665 in. vertically from the orifice. The discharge is 98.987 gallons in 569.218 seconds. Find the coefficients of discharge, velocity, contraction, and resistance. Ans. .6014; .945; .636; .118. 26. A square box 2 ft. in length and i ft. across a diagonal is placed with a diagonal vertical and filled with water. How long will it take for the whole of the water to flow out through a hole at the bottom of .02 sq. ft. area ? (c .625.) Ans. 97.48 sees. 27. A pyramid 2 ft. high, on a square base, is inverted and filled with water. Find the time in which the water will all run out through a hole of .02 sq. ft. at the apex. A side of the base is i ft. in length. (c. .625.) Ans. 15.08 sec, 28. Find the discharge under a head of 25 ft. through a thin-lipped square orifice of i sq. in. sectional area, (a) when it has a border on one side, (b) when it has a border on two sides. Ans. (a) .3575 cu. ft. per sec.; (b) .3706 cu. ft. per sec. 29. A vessel in the form of a paraboloid of revolution has a depth of 16 in. and a diam. of 12 in. at the top. At the bottom is an orifice of i sq. in. sectional area. If water flows into the vessel at the rate of 2 T V cubic feet per minute, to what level will the water ultimately rise ? How long will it take to rise (a) 11 in., (b) 11.9 in., (c) 11.99 m -> (X) I2 in- above the orifice? If the supply is now stopped, how long (e) will it take to empty the vessel ? Ans. 12 inches; (a) 83.095 sec.; (b) 124.2 sec.; (c) 263.9 sec.; (d) an infinite length of time ; (e) 11.3 sec. 30. If the vessel in Question 29 is a semi-sphere i ft. in diameter, to what height will the water rise ? How long will it take for the water to rise (a) 11 in., (b) 12 in. above the orifice ? How long (c) will it take to empty the vessel ? Ans. 12 inches ; (a) 67.16 sec. ; () 81.46 sec. ; (c) 24.13 sec. 31. In a vortical motion two circular filaments of radii ri , r 2 , of ve- locities Vi,Vt, and of equal weight Ware made to change place. Show 7/ 2 that a stable vortex is produced if =const.; and if r 2 > r\ , show that the surfaces of equal pressure are cones. FLOW THROUGH 1 ORIFICES, OVER WEIRS, ETC. 65 32. Prove that for a Borda's mouthpiece running full the coefficient of discharge is . 4/2 33. The surface of the water in a tank is kept at the same level; obtain the discharge at 60 in. below the surface (a) through a circular orifice i sq. in. in area, (U) through a cylindrical ajutage of the same sectional aYea fitted to the outside, (c) through the same ajutage fitted to the inside, and determine the mechanical effect of the efflux in each case. Ans. (a) 4 36 Ibs. per sec. (ff) 6.356 " " " (rf 3.488 " " 20.514 ft.-lbs. per sec. 21.369 " " " 1744 " " " 34. Water is discharged under a head of 64 ft. through a short cylin- drical mouthpiece 12 in. in diameter. Find (a) the loss of head due to shock, () the volume of disdharge in cubic feet per secJnd, (c) the energy of the issuing jet. (g = 32.) Ans. (a) 20.96 ft. ; (8) 51.54 cub. ft. ; (c) 393.8 H. P. 35. If a bell-mouth is substituted for the mouthpiece in the preced- ing question, find the discharge and the mechanical effect of the jet. Ans. 61.6 cub. ft. per sec. ; 470.6 H. P. 36. Compare the energies of a jet issuing under an effective head of 100 ft. through (i) a 12-in. cylindrical ajutage, (2) a 12-in. divergent aju- tage, (3) a 12-in. convergent ajutage, the angle of convergence being 21. Draw the plane of charge in each case. Ans. (i) 393.8 H. P.; (2) 672.28 H. P.; (3) 618.23 H. P. 37. Find the discharge through a rectangular opening 36 in. wide and 10 in. deep in the vertical face of a dam, the upper edge of the opening being 10 ft. below the water surface. Ans. 40.2 cub. ft. per sec. 38. Find the discharge in pounds per minute through a Borda's mouthpiece i in. in diameter, the lip being 12 in. below the water- surface. Ans. 87.714 Ibs. 39. Sometimes the crest of a dam is raised by floating a stick L into the position Zi , where it is supported against the verticals. The stick then falls of itself into position Li and rests on the crest. Explain the reason of this. 40. A sluice 3 ft. square and with a head of 12 ft. over the centre has, from the thickness of the frame, the contraction suppressed on all sides when fully open ; when partially open, the contraction exists on the upper edge, i.e., against the bottom of the gate, which is formed of a thin sheet of metal. Find the discharge in cubic feet when opened i ft., 2 ft. and also when fully open. Ans. 57.77 ; 114.45 ; '75-9. 66 HYDRA ULICS. 41. What quantity of water flows through the vertical aperture of a dam, its width being 36 in. and its depth 10 in. ; the upper edge of the aperture is 16 ft. below the surface. Ans. 50.65 cub. ft. per sec. 42. 264 cubic feet of water are discharged through an orifice of 5 sq. ins. in 3 min. 10 sec. Find the mean velocity of efflux. Ans. 64 ft. per sec. 43. One of the locks on the Lachine Canal has a superficial area of about 12,150 sq. ft., and the difference of level between the surfaces of the water in the lock and in the upper reach is 9 feet. Each leaf of the gates is supplied with one sluice, and the water is levelled up in 2 min, 48 sees. Determine the proper area of the sluice-opening. (Centre of sluice 20 ft. below surface of upper reach.) Ans. Area of one sluice = 43.73 sq. ft. 44. The horizontal section of a lock-chamber may be assumed a rectangle, the length being 360 ft. When the chamber is full, the sur- face width between the side walls, which have each a batter of i in 12, is 45 ft. How long will it take to empty the lock through two sluices in the gates, each 8 ft. by 2 ft., the height of the water above the centre of the sluices being 13 feet in the lock and 4 feet in the canal on the down- stream side. Ans. 594 sec., c being .625. 45. Water approaches a rectangular opening 2 ft. wide with a velocity of 4 ft. per second. At the opening the head of water over the lower edge = 13 ft., and over the surface of the tail-race = 12 ft.; the discharge through the opening is 70 cub. ft. per second. Find the height of the opening. Ans. 1.022 ft. 46. The water in a regulating-chamber is 8 ft. below the level of the water in the canal and 8 ft. above the centre of the discharging-sluice. Determine the rise in the canal which will increase the discharge by 10 per cent. Ans. 1.68 ft. The horizontal sectional area of the chamber is constant and equal to 400 sq. ft.; in what time will the water in the chamber rise to the level of that in the canal, if the discharging-sluice is closed; the sluice between the canal and chamber being 3 sq. ft. in area? Ans. 150.83 sec. 47. A lock on the Lachine Canal is 270 ft. long by 45 ft. wide and has a lift of 8 ft.; there are two sluices in each leaf, each 8f ft. wide by 2 ft. deep ; the head over the horizontal centre line of the sluices is 19 ft. Find the time required to fill the lock. Ans. 164.6 sec. 48. Show that the energy of a jet issuing through a large rectangular orifice of breadth B is i2$B(ff Hi*), Hi , H* being the depths below the water-surface of the upper and lower edges of the orifice, and the coefficient of discharge being .625. 49. A reservoir at full water has a depth of 40 ft. over the centre of the discharging-sluice, which is rectangular and 24 in. wide by 18 in. deep. Find the discharge in cubic feet per second at that depth, and also FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 6? when the water has fallen to 30, 20, and 10 ft., respectively; find the mechanical effect of the efflux in each case. Ans. 94.8 cub. ft.; 82.1 cub. ft.; 67 cub. ft.; 47.4 cub. ft.; 431.2 H.P.; 280 H.P.; 152.5 H.P.; 53.95 H.P. 50. Require the head necessary to give 7.8 cubic feet per second through an orifice 36 sq. in. in sectional area. Aps. 38.9 ft. 51. The upper and lower edges of a vertical rectangular orifice are 6 and 10 feet below the surface of the water in a cistern, respectively ; the width of the orifice is i ft. Find the discharge through it. Ans. 5642 cub. ft. per sec. 52. To find the quantity of water conveyed away by a canal 3 ft. wide, a board with an orifice 2 ft. wide and i ft. deep is placed across the canal and dams it back until it attains a height of 2 ft. above the bottom and if ft. above the lower edge of the orifice. Find the dis- charge, (c = .625.) Ans. 17.59 cub - ft- per sec. 53. Six thousand gallons of water per minute are forced through a line of piping ABC and are discharged into the atmosphere at C, which is 6 ft. vertically above A. The pipe AB is 12 in. in diameter and 12 ft. in length ; the pipe J5C is 6 in. in diameter and 12 ft. in length. Disre- garding friction, find the " loss in shock " and draw the plane of charge. Ans. Loss of head in shock = 57.9 ft. 54. What should be the height of a drowned weir 400 ft. long, to deepen the water on the up-stream side by 50 per cent, the section of the stream being 400 ft. x 8 ft., and the velocity of approach 3 ft. per second ? Ans. 8.396 ft. 55. The two sluices each 4 ft. wide by 2 ft. deep in a lock-gate are submerged one half their depth. The constant head of water above the axis of the sluice is 12 ft. Find the discharge through the sluice, the velocity of approach being 4 ft. per second. Ans. 16626.2 cub. ft. per minute. 56. Find the flow through a square opening, one diagonal being ver- tical and 12 in. in length, and the upper extremity of the diagonal be- ing in the surface of the water. Ans. 1.727 cub. ft. per sec. 57. The locks on the Montgomeryshire Canal are 81 ft. long and 7f ft. wide ; at one of the locks the lift is 7 ft.; a 24-in. pipe leads the water from the upper level and discharges below the surface of the lower level into the lock-chamber ; the mouth of the pipe is square, 2 ft. in the side, and gradually changes into a circular pipe 2 ft. in diameter. Find time of filling the lock, (c = i.) Ans. 130 sees. 58. A canal lock is 115.1 ft. long and 30.44 ft. wide; the vertical depth from centre of sluice to lower reach is 1.0763 ft., the charge being 6.3945 ft. ; the area of the two sluices is 2 x 6.766 sq. ft. Find the time of filling up to centre of sluices, (c = .625 for the sluice, but is reduced 68 HYDRA ULICS. to .548 when both are opened.) Also, find time of filling up to level of upper reach from centre of sluice-doors. Ans. 25 sec.; 298 sees. 59. A reservoir half an acre in area with sides nearly vertical, so that it may be considered prismatic, receives a stream yielding 9 cub. ft. per second, and discharges through a sluice 4 ft. wide, which is raised 2 ft. Calculate the time required to lower the surface 5 ft., the head over the centre of the sluice when opened being 10 feet. Ans. 1079 sees. 60. Show that in a channel of V section an increment of 10 per cent in the depth will produce a corresponding increment of 5 per cent in the velocity of flow and of 25 per cent in the discharge. 61. The angle of a triangular notch is 90. How high must the water rise in the notch so that the discharge may be 1000 gallons per minute? Ans. 1 2 inches very nearly. 62. Show that upon a weir 10 feet long with 12 inches depth of water flowing over, an error of i/iooo of a foot in measuring the head will cause an error of 3 cubic feet per minute in the discharge, and an error of i/ioo of a foot in measuring the length of the weir will cause an error of 2 cubic feet in the discharge. 63. In the weir at Killaloe the total length is noo ft., of which 779 ft. from the east abutment is level, while the remainder slopes i in 214, giving a total rise at the west abutment of 1.5 ft. Calculate the total discharge over the weir when the depth of water on the level part is 1.8 ft., which gives .3 ft. on highest part of weir. (Divide slope into 8 lengths of 40 ft. each, and assume them severally level, with a head equal to the arithmetic mean of the head at the beginning and end of each length.) Ans. 7483 cub. ft. per sec. 64. A watercourse is to be augmented by the streams and springs above its level. The latter are severally dammed up at suitable places and a narrow board is provided in which an opening 12 in. long by 6 in. deep is cut for an overfall ; it was surmised that this would be suf- ficient for the largest streams; another piece attached to the former would reduce the length to 6 in. for smaller streams. Calculate the delivery by the following streams: In No. i stream with the 12-in. notch, depth over crest = .37 ft. " No. 2 " " " 6-in. " " " " = .41 ft. -; " No. 3 " " " 12-in. " " " " = .29 ft. " No. 4 " " " 6 in. " " " " = .19 ft. (Take into account the side contractions.) Ans. No. i, .695 cub. ft. ; No. 2, .3658 cub. ft. ; No. 3, .4904 cub. ft.; No. 4, .1275 cub. ft. 65. The horizontal sectional area of a reservoir is constant and = 10,000 square feet. When the reservoir is full, a right-angled notch 2 ft. deep is opened. Find the time in which the level of the water falls- to the bottom of the notch. Ans. 15.3 min. FLOW THROUGH ORIFICES, OVER WEIRS, ETC. 69 66. A weir passes 6 cubic feet per second, and the head over the crest is 8 inches. Find the length of the weir. Ans. 3.3068 ft. 67. A weir 400 ft. long, with a 9-in. depth of water on it, discharges through a lower weir 500 ft. long. Find the depth of water on the latter. Ans. .6457 ft. 68. A stream 30 ft. wide, 3 ft. deep, discharges 310 cubic feet per second ; a weir 2 feet deep is built across the stream. Find increased depth of latter, (a) neglecting velocity of approach, (b) taking velocity of approach into account. Ans. (a) 1.26 ft. to 1.265 ft-J (6) 1. 19 ft. 69. A weir is 545 ft. long; how high will the water rise over it when it rises .68 ft. upon an upper weir 750 ft. long? Ans. .8413 ft. 70. In a stream 50 ft. wide and 4 ft. deep water flows at the rate of loo ft. per minute ; find the height of a weir which will increase the depth to 6 ft., (i) neglecting velocity of approach, (2) taking velocity of approach into account. Ans. (i) 4.4126 ft; (2) 4.4509 ft. 71. A stream 50 ft. wide and 4 ft. deep has a velocity of 3 ft. per second ; find the height of the weir which will double the depth, (i) neglecting velocity of approach, (2) taking velocity of approach into ac- count. Ans. (i) 5.615 ft.; (2) 5.7688 ft. 72. A stream 80 ft. wide by 4 ft. deep discharges across a vertical section at the rate of 640 cubic feet per second ; a weir is built in the stream, increasing its depth to 6 ft. Find the height of the weir. Ans. 4.233 ft. 73. Salmon-gaps are constructed in a weir ; they are each 10 ft. wide and their crests are 18 in. below the weir crest. Calculate the discharge down three of these gaps, the water on the level rjart of the weir being 8 in. deep. Ans. 238.15 cub. ft. per sec. 74. A pond whose area is 12,000 sq. ft. has an overfall outlet 36 in. wide, which at the commencement of the discharge has a head of 2.8 ft. Find the time required to lower the surface 12 in. Ans. 354.72 sec. 75. How much water will flow in an hour through a rectangular notch 24 in. wide, the surface of still water being 8 in. above the crest of the notch ? (Take into account side contraction.) Ans. 3.386 ft. 76. Show that when the water flowing over has a depth greater than .3874 ft. it is carried completely over the longitudinal opening, .83 ft. in width. At what depth does all the water flow in ? , Ans. .221 ft. FIG. 49. CHAPTER II. FLUID FRICTION. I. Fluid Friction. The term fluid friction is applied to the resistance to motion which is developed when a fluid flows over a solid surface, and is due to the viscosity of the fluid. This resistance is necessarily accompanied by a loss of energy caused by the production of eddies along the surface, and similar to the loss which occurs at an abrupt change of sec- tion, or at an angle in a pipe or channel. Froude's experiments on the resistance to the edgewise motion of planks in a fluid mass, the planks being T \ in. thick, 19 in deep, and I to 50 ft. long, each plank having a fine cut- water and run, are summarized in the following table : Length of Surface in Feet. Nature of Surface 2 Feet. 8 Feet. 20 Feet. 50 Feet. Covering. A B c A B C A B C A B C Varnish 2.OO 41 3QO 1.85 .325 .264 1.8 5 .2 7 8 .240 I.8 3 .250 .226 Paraffine 38 37O 1.94 .314 ?6o 1.93 .271 .237 Tinfoil 2 16 qo 2QC I QQ .278 .263 I .QO ?6^ 2 44 r 83 .246 272 Calico . . . i cn 87 72^ I .Q2 .626 .504 I. 80 . C.T.I 447 T 87 .474 42^ Fine sand 2.00 .81 .690 2.00 .583 450 2.00 .480 384 2.06 405 337 Medium sand 2.OO .90 .7302.00 .625 .488 2.00 534 465 2.OO .488 456 Coarse sand 2.00 I. 10 .880 2.OO .714 .520 2.0O .588 .490 Columns A give the power of the speed (v) to which the re- sistance is approximately proportional. Columns B give the mean resistance in Ibs. 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 70 FLUID FRICTION. 7 1 the heading, each plank having a standard speed of 10 ft. per second. The resistance at other speeds can be easily calculated. An examination of the table shows that the mean resistance per square foot diminishes as the length of the plank increases. This may be explained by the supposition that the friction in the forward portion of the plank develops a force which drags the water along with the surface, so that the relative velocity of flow over the rear portion is diminished. Again, the de- crease of the mean resistance per square foot is .132 Ib. when the length of a varnished plank is increased from 2 to 20 ft., while it is only .028 Ib. when the length increases from 20 to 50 ft. Hence, for greater lengths than 50 ft. the decrease of resistance may be disregarded without much, if any, practical effect. Thus, generally speaking, these experiments indicate tha-t the mean resistance is proportional to the #th power of the relative velocity, n varying from 1.83 to 2.16, and its average value being very nearly 2. Colonel Beaufoy, as a result of experiments at Deptford, also assumed the mean resistance to be proportional to the nth power of the relative velocity, the value of n in three series of observations being 1.66, 1.71 and 1.9. The frictional resistance is evidently proportional to some function of the velocity, F(v), which should vanish when v is nil, as when the surface is level, and should increase with v. Coulomb assumed the function F(v) to be of the form av -f- bv* , a and b being coefficients to be determined by experi- ment. Experiment shows that when v does not exceed 5 ft. per minute the resistance is directly proportional to the veloc- ity, but that it is more nearly proportional to the square of the velocity when the velocity exceeds 30 ft. per minute ; or, F(v) = av when v < 5 ft. per minute, and F(v) = bv^ when v > 30 ft. per minute. Again, observations on the flow of water in town mains indicate that no difference of resistance is developed under 72 HYDRA ULICS. widely varying pressures, and this independence of pressure is also verified by Coulomb's experiment showing that if a disk is oscillated in water there is no apparent change in the rate of decrease of the oscillations, whether the water is under atmos- pheric pressure or not. From the preceding and other similar experiments the fol- lowing general laws of fluid friction have been formulated : (1) The frictional resistance is independent of the pressure between the fluid and the surface over which it flows. (2) The frictional resistance is proportional to the area of the surface. (3) The frictional resistance is proportional to some func- tion, usually the square, of the velocity. To these three laws may probably be added a fourth, viz.: (4) The frictional resistance is proportional to the density of the fluid. A fifth law, viz., that " the frictional resistance is indepen- dent of the nature of the surface against which the fluid flows," has been sometimes enunciated, and at very low velocities the law is approximately true. At high velocities, however, such as are common in engineering practice, the resistance has been shown by experiment, and especially by the experiments carried out by Darcy, to be very largely influenced by the nature of the surface. Let p be the frictional resistance in pounds per square foot of surface at a velocity of I ft. per second. Let A be the area of the surface in square feet. Let v be the relative velocity of the surface and the water in which it is immersed. Let R be the total frictional resistance. Then from the laws of fluid friction R = p . AV*. 2j? Take/ = p, w being the specific weight of the fluid. Then R = fwA. FLUID FRICTION. 73 The coefficient f is approximately constant for any given surface, and is termed the coefficient of fluid friction. The power absorbed by the frictional resistance v* = pAv' X v = pAv* = fwA . o TABLE GIVING THE AVERAGE VALUES OF / IN THE CASE OF LARGE SURFACES MOVING IN AN INDEFINITELY LARGE MASS OF WATER. Surface. Coefficient of Friction (/"). New well-painted iron plate .............. 00489 Painted and planed plank ............... 0035 Surface of iron ships ... ................. 00362 Varnished surface ........................ 00258 Fine sand surface ........................ 00418 Coarse sand surface .... ................. 00503 2. Surface Friction of Pipes. Assuming that the laws of fluid friction already enunciated hold good when water flows through a pipe, it has been shown by numerous experiments that the coefficient of friction /lies between the limits .005 and .01, its average value under ordinary conditions being about .0075. No single value of f is applicable to very different cases. Indeed, /depends not only upon the condition of the surface, but also upon the diameter of the pipe and the veloc- ity of the water. Some authorities have expressed its value by a relation of the form a and b being constants whose values are to be determined by experiment. The following table gives some of the best numerical results obtained for a and b\ 74 HYDRA ULICS. Authority. a b Prony ........ .......... 00021230 .00003466 D'Aubuisson ............ 0002090 .000037608 Eytelwein ............... 00017059 .00004441 In pipes of small diameter in which the velocity of flow is less than 4 in. per second the term a may be disregarded so that In ordinary practice and when the pipes have been in use for some time, the velocity usually exceeds 4 in. per second, and the term may then be disregarded, so that Now Darcy's experiments have shown that it is more cor- rect to assume that a and b, instead of being constant, are variable, and Darcy expressed them as functions of the diam- eter of the pipe. Thus, for pipes in which the velocity exceeds 4 in. per second, Darcy took / , g '' ^d' d being the diameter of the pipe, and a and ft coefficients. Darcy also gave the following values for a and ft : a ' ft For drawn wrought-iron or smooth cast-iron pipes 0001545 .000012973 For pipes with surfaces covered by light incrustations 0003093 .00002598 FLUID FRICTION. 75 These coefficients can be put into the following very simple form without sensibly altering their values : For clean pipes ........... f= .005(1 -[- - ) For slightly incrusted pipes / = .01(1 -f- - J d being the diameter in feet. Darcy proposed to include all cases by expressing /"more generally in the form in which, for new and smooth iron pipes, a = ,00003959, ft .00002603125 ; a f = .000064375, ft' = .000000335625. These values are rarely of any practical use. TABLE GIVING DARCY'S VALUES OF / FOR VELOCITIES EXCEEDING 4 IN. PER SECOND. Diam. Value of/ Diana. Value of f. Diam. Value of/. of of of Pipe Pipe Pipe m New Incrusted in New Incrusted in New Incrusted Inches. Pipes. Pipes. Inches. Pipes. Pipes. Inches. Pipes. Pipes. 2 .0075 .0150 9 .00556 .OITII 27 .00519 .01037 3 .00667 01333 12 .00542 .01083 30 .00517 .01033 4 .00625 .0125 15 00533 .01067 36 .00514 .01028 5 .0060 .OI2 18 .00528 .01056 42 .00512 .OIO24 6 .00583 .01167 2.1 .00524 .01048 4 8 .00510 .01021 7 .00571 .01143 24 .00521 .01042 54 .00509 .OIOI9 8 .00563 .01125 76 HYDRAULICS. Again, Weisbach has proposed the formula -* Vv where a = .003598 and b = .004289. 3. Resistance of Ships. The motion of a ship through water causes the production of waves and eddies, and the total resistance to the movement of a ship is made up of a frictional resistance, a wave-making resistance, and an eddy-making re- sistance. Although there is no theory by which the resistance at a given speed of a ship of definite design can be absolutely determined, Froude's experiments render it possible to make certain inferences and furnish some useful data. According to Froude, the frictional resistance is sensibly the same as that of a rectangular surface moving with the same speed, of the same length as the ship in the direction of motion, and of an area equal to the immersed surface of the ship. Experiments seem to indicate that as the speed increases, the frictional resistance of well-designed ships with clean bottoms is from 90 to 60 per cent of the total resistance, and that the percentage is greater when the bottoms become foul. The wave-making resistance is especially affected by the form and proportions of the ship, depending, for a given length, upon the proportions of the entrance, middle body, and run. For every ship there is a limit of speed below which the resistance is approximately proportional to the square, of the speed, being chiefly due to friction, and beyond which it in- creases more rapidly than as the square. The eddy-resistance in the case of well-formed ships should not exceed about 10 per cent of the total resistance, and is often much less. Froude's law of resistance may be enunciated as follows : Let / /, be the lengths of a ship and its model. Let A lt A^ be the displacements of a ship and its model. Let /? R^ be the resistances of a ship and its model at the .speeds i v l and v t . FLUID FRICTION. 7? Then, if _i _i_ __ \ _ V " / ~~ /f *' 2 *s ^a the resistances are in the ratio of Hence, too, the H. P., and therefore also the coal consumption per hour, is proportional to Rv, that is, to A 1 or / 5 , and the coal consumption per mile is proportional to A or to / 3 . Again, R is proportional to / 3 ; that is, to / X / 3 1 that is, to v* X ^ ; and it is sometimes convenient to express the resistance irt pounds in the form v being the speed in knots, A the displacement in tons, and k a coefficient depending upon the type of ship and varying from .55 to .85 when the bottom is clean. CHAPTER III. FLOW OF WATER IN PIPES. 1. Assumptions. In the ordinary theory of the flow of water in a pipe it is assumed that the water consists of thin plane layers perpendicular to the axis of the pipe, that each layer is driven through the pipe by the action of gravity and by the difference of pressure on its plane faces, and that the liquid molecules in any layer at any given moment will also be found in a plane layer after any interval of time. In such motion the internal work done in deforming a layer may be generally dis- regarded. It is further assumed that there is no variation of velocity over the surface of a layer, and this is equivalent to saying that each liquid molecule in a cross-section has the same mean ve- locity. The disagreement of these assumptions with the results of recent experimental researches will be referred to in a subse- quent article. 2. Steady Motion in a Pipe of Uniform Section. Since the motion is to be steady, the same volume Q cub. ft. of water will always arrive at any given cross-section of A square feet with the same mean velocity v ft. per second. Then Q = Av. But since the pipe is of constant diameter, A is constant, and hence also v is constant, so that the mean velocity is the same throughout the whole length of the pipe. Consider an elementary mass of the fluid AABB, bounded by the pipe and by the two cross-sections AA, BB. Let dl 78 FLOW OF WATER IN PIPES. 79 FIG. 50. be the length AB of the element, the length / ft. of the pipe being measured along the axis from any origin O. Let z, z + dz be the eleva- tions in feet above a datum line of the centres of pressure in the cross-sections A A, BB, respectively. Let p, p + dp be the intens- ities of the pressures on these cross-sections in pounds per square foot. Let P be the perimeter of the pipe. Let w be the specific weight of the water in pounds per cubic foot. Work Done by Gravity. In one second wQ Ibs. of water are transferred from AA to BB, falling through a vertical dis- tance of dz ft. Thus the work done by gravity per second = wQ . dz, a positive quantity if dz is negative, and vice versa. Work Done by Pressure. The total pressure on A A paral- lel to the axis = pA ; the total pressure on BB parallel to the axis = (p + dp) A. Therefore ^the total resultant pressure parallel to the axis in the direction of motion = A . dp, and the work done per second on the volume Q by this pressure = Q . dp. Note. The work done by the pressure at the pipe surface is nil, as its direction is at right angles to the line of motion. Work Absorbed by Frictional Resistance. From the laws of fluid friction this work per second is evidently p P . dl . F(v) X v = -r . Q . F(v) . dl, the sign being negative as the work is done against a resistance. 80 HYDRA ULICS. Since the motion is steady, the work done by the external forces must be equivalent to the work absorbed by the fric- tional resistance, and hence wQ . dz Q . dp - Q . F(v) . dl o, or , dp P F(v} (Jn I * _l V ' /y/ Q w A " w Integrating, ^ + -f-j../ =a constant = H, w A w so that H ft.-lbs. per pound of fluid is the uniformly distributed total constant energy. A is called the hydraulic mean radius of a pipe and will be denoted by m. Take W 2g the value adopted in ordinary practice, f being the coefficient of friction. Then w in 2g Let #, , A l ,/, be the elevation above datum, the area of the cross-section, and the intensity of the pressure at any point X on the axis of the pipe distant / x from the origin (Fig. 51). Let 2 , AS, pi be the elevation above datum, the area of the cross-section, and the intensity of the pressure at any other point Y on the axis distant 4 from the origin (Fig. 51). FLO W OF WATER IN PIPES. Then, from the equation just deduced, 81 ,.+* + *=*=*+> + !. w m 2g w m 2g Hence w m 2g L being the length / 2 ^ of the pipe between the two points K FIG. 51. Let vertical tubes (pressure-columns) be inserted in the.pipe at X and at Y. The water will rise in these tubes to the levels C and Z>, and evidently being the intensity of the atmospheric pressure. 82 HYDRA ULICS. Hence, if CX and D Fare produced meet the datum line in E and F, I A i/^j^iA ri? j A #. -+- = -Sj -h CA -f- = Czi -t- - JL ze; w w and # a + = a + jC>F+ = Z>F+ . w w w Therefore w wi m 2g G being the point in which the horizontal through C meets FD produced. DG is called the " virtual fall " of the pipe, being the fall of level in the pressure-columns; and since there would be no fall. of level if the friction were nil, DG is said te be the head lost in friction in the distance XY. Denote this head by h\ then = m 2g and therefore _/ L m 2g This ratio - is designated the virtual slope of the pipe, and JL/ is the head lost in friction per unit of length It will be denoted by *', so. that If the section of the pipe is a circle of diameter d, or a square with a side of length d, then and FLOW OF WA7^ER IN PIPES. t _ 83 A d __ = L ~ d 2g 3. Influence upon the Flow of the Pipe's Position and Inclination. In Fig. 5 1 join CD. Now since the fall of level (h) is proportional to L y the free surface in any other column between X and Y must also be on the line CD. Thus the pressure/ 7 at any intermediate point M distant x(==. XM) from X is given by w w w Hence, at every point of a pipe laid below CD, the fluid pres- sure (p r ) exceeds the atmospheric pressure (/ ) by an amount w . MN, so that if holes are made in such a pipe the water will flow out and there will be no tendency on the part of the air to flow in. In pipes so placed vertical bends may be intro- duced, care being taken to provide for the removal of the air which may collect in the upper parts of the bends. If the line of the pipe coincides with CD, i.e., with the vir- tual slope or line of free surface level, MN = o, and the fluid pressure is equal to that of the atmosphere. If holes are now made in the pipe it can easily be shown by experiment that there will be neither any tendency on the part of the water to flow out nor on the part of the air to flow in. Next take CC' = DD' = and join CD'. w J If the pipe is placed in any position between CD and C ' D r MN becomes negative, and the fluid pressure in the pipe is less than that of the atmosphere. If holes are made in this pipe, there will be no tendency on the part of the water to flow out, 84 HYDRA ULICS. but the air will flow in. Thus, if a pipe rises above the line of virtual slope, there is a danger of air accumulating in the pipe and impeding, or perhaps wholly stopping, the flow. No verti- cal bends should be introduced, as the air is easily set free and would collect in the upper parts of the bends, with the effect of impeding the flow and of acting detrimentally upon the water itself, which the liberation of the air renders less wholesome. If the line of pipe coincide with CD', then the fluid pressure is nil. Finally, if the pipe at any point rises above CD', the press- ure becomes negative, which is impossible. In fact, the con- tinuity of flow is destroyed, and the pipe will no longer run full bore. Air will be disengaged and will rise and collect at the point in question, so that in order to prevent the flow being wholly impeded, it will be necessary to introduce an air-chamber at this point from which the air can be removed when required. Note. In the preceding it has been assumed that the pipe is straight. If the pipe is curved, so also is the line of virtual slope. In ordinary practice, however, the vertical changes of level in a pipe at different points are small as compared with the length of the pipe, and distances measured along the pipe are sensibly proportional to distances measured along the horizontal projection of the pipe. Hence the line of virtual slope may be assumed to be a straight line without error of practical importance. 4. Transmission of Energy by Hydraulic Pressure. Let Q cub. ft. of water per second be driven through a pipe of diameter d ft. and length L ft. under a total head of H ft. Also let n per cent, of the total head be absorbed in overcoming the frictional resistance in the pipe. Then the head expended in useful work = H h H-h and the efficiency = FLOW OF WATER IN PIPES. 8$ Again, - h - - - 100 : ~d~ 2g ~'' Since Q = z/, and g is assumed to be 32, thus, ^ InHd* - toy T~' ancf the work transmitted in foot-pounds per second 14 If ^V= the number of horse-power transmitted, then jv - _L i ; _ 5 /^ 1 A^ 8 ^ "550 H V /^ "28V "7^~' and this equation also gives the distance L to which TV horse- power can be transmitted with a loss of n per cent of the total head. Again, ffi . // 2fL V* 2/Lw v* the efficiency = I = I : -77 T l -^ H gH d g pd> p( = wH) being the pressure corresponding to the head H. Thus, the efficiency is constant if - is constant. pa Assuming this to be the case, take v* = c* .pd. Then the total energy transmitted = wQH ' = w vH 4 86 HYDRA ULICS. If it be also assumed that the thickness / of the pipe-metal is so small that the formula pd = 2ft holds true, f being the circumferential stress induced in the metal, then the energy transmitted = F being the volume of the pipe per unit of length. Hence, for a given volume (V) of metal and a constant efficiency, the energy transmitted is a maximum when pd is a maximum. If / is increased beyond a certain limit, the ratio -^ is no longer small and the thickness t will have a greater value than that given by the equation pd = 2. ft. Then the cost of the pipe will also increase. On the other hand, if d is increased the ratio -^, and therefore also the pressure/, will remain small, and thus the cost of the pipe will not increase. Hence it is more economical to employ large pipes and low pressures than small pipes and high pressures. Note. The efficiency diminishes as v increases, so that, as far as the efficiency is concerned, it is advantageous to transmit the energy at a low speed. 5. Flow in a Pipe of Uniform Section and of Length Z, connecting two Reservoirs at Different Levels. Let z ft. be the difference of level between the water-surface in the two reservoirs. FLO W OF WATER IN PIPES. 8? FIG. 52. The work done per second is evidently equal to the work done by the fall of wQ pounds of water through the vertical distance z, and is expended (1) In producing the velocity of flow v feet per second which requires a head of z l feet and an expenditure of wQz l foot-pounds of work per second ; (2) In overcoming the resistance at the entrance from the upper reservoir into the pipe, which requires a head of s a feet and an expenditure of wQz^ foot-pounds of work per second. (3) In overcoming the frictional resistance which requires a head of z^ feet and an expenditure of wQz^ foot- pounds of work per second. Thus wQz = wQz l + wQz^ + or z = z * *. Now #, - - feet, and the corresponding energy wQz^ is ultimately wasted in producing eddy motions, etc., in the lower reservoir. v* z^ may be expressed in the form n feet, n being a coeffi- cient whose value varies with the nature of the construction of the entrance into the pipe. If the pipe-entrance is bell-mouth in form, n = .08, but if it is cylindrical, n = .5. Finally, 88 HYDRA ULICS. f , = ft ., , m w d 2g F(v) v* Baking - - =f , as is usual in practice. Hence 2g\ d since Q = -- v, and g is assumed to be 32. 4 For given values of Q and z a first approximate value of d may be obtained from the last equation by neglecting the term Q* rr;(l + ) Call this value d v and substitute it for the d in A/L the term j- within the brackets. A second approximation may now be made by deducing d from the formula and the operation may be again repeated if desired. Generally speaking, I + n is usually very small as compared with , and may be disregarded without error of practical importance. The formula then becomes _ which is known as Chezy's formula for long pipes. In fact, the term I + n need only be taken into account in the case of short pipes and high velocities. FLOW OF WATER IN PIPES. 8 9 6. Losses of Head due to Abrupt Changes of Section, Elbows, Valves, etc. When the velocity or the direction of motion of a mass of water flowing through a pipe is abruptly changed, the water is broken up into eddies or irregular mo- tions which are soon destroyed by viscosity, the corresponding energy being wasted. CASE I. Loss due to a sudden contraction. (Art. 16, Chap. I.) (a) Let water flow from a pipe (Fig. 53), or from a reser- voir (Fig. 54) into a pipe of sectional area A. FIG. 53- FIG. 54. Let c c be the coefficient of contraction. Then the area of the contracted section = c c A, and the loss of head = (--. ,Y 2 V I 2g V, where m = ( 2 I . = m V2 2? The value of m has not been determined with any great degree of accuracy ; but if c c = .64, then m = .316. HYDRAULICS. When the water enters a cylindrical (not bell-mouthed) pipe from a large reservoir, the value of m is about .505. (b) Let the water flow across the abrupt change of section through a central ori- fice in a diaphragm placed as in Fig. 55. Let a be the area of the orifice. Then c,a is the area of the contracted section, and the loss of head = ( W (A y where m = I I J V^z t According to Weisbach, I A Vv* v* = I I J = m , \c e a I 2g 2g' f- = I .2 3 4 5 c c = .616 .614 .612 .610 .607 m = 231.7 50.99 19.78 9.612 5.256 i- .6 7 .8 9 I.OO C c .605 .603 .601 .598 .596 m = 3.077 1.876 1.169 734 .48 central ~ =i orifice of area a. nlaced in a cylin- -.V I drical pipe of 7 1 sectional area A as in FIG. 56. Fig. 56. The " contracted area " of the water = c being the elbow angle. Weisbach deduced this formula from the results of experi- ments with pipes 1.2 in. in diameter. 9 2 HYDRA ULICS. The velocity v l with which the water flows along the length AB may be resolved into a component v with which the water flows along BC and a component u at right angles to the FIG. 58. direction of v. The component u and therefore the corre- sponding head, viz., , is wasted. The component u evidently diminishes with the angle and becomes nil when a gradually and continuously curved bend is substituted for the elbow. CASE IV. Weisbach gives the following empirical formula for the loss of head at a bend in a pipe : h b = m t , d\k where m = .131 -f 1.847 for a circular pipe of diameter d, p being the radius of curvature of the bend, and FIG. 59. m = .124+ 3-104 for a pipe of rectangular section, s being the length of a side of the section parallel to the radius of curvature (p) of the bend. CASE V. Valves, Cocks, Sluices, etc. The loss of head in each of the cases represented by the several figures may be traced to a contraction of the stream similar to the con- FLO W OF WATER IN PIPES. 93 traction which occurs in the case of an abrupt change of sec- v* tion. The loss may be expressed in the form m , and the following tables give the results obtained by Weisbach. (a) Sluice in Pipe of Rectangular Section. (Fig. 60.) Area of pipe = a ; area of sluice = s. = i .9 5 -4 .2 .1 m = .oo .09 .39 .95 2.08 4.02 8.12 17.8 44.5 193 FIG. 60. (b) Sluice in Cylindrical Pipe. (Fig. 61). s = ratio of height of opening to diameter of pipe. s= i .875 -75 -625 .5 .375 .25 .125 m = .00 .07 .26 .81 2.06 5.52 17.00 97.8 (c) Cock in Cylindrical Pipe (Fig. 62). s = ratio of cross-sections; 6 = angle through which cock is turned. FIG. 61. FIG. 62. If = 5 s = ,926 m = .05 If/= 40 .85 .2 9 45 = 385 .315 = 17.3 31.2 15 .772 75 50 .25 52.6 FIG 63. 20 25 30 35 .692 .613 .535 .458 1-56 3-1 547 55 106 60 65 .137 .091 206 486 82 00 oo (d) Throttle-valve in Cylindrical Pipe (Fig. 63) angle through which valve is turned. 94 HYDRA ULICS. If 61 =5 10 15 20 25 30 35 40 ^ = .24 .52 .90 1.54 2.51 3.91 6.22 10.8 If 0=45 50 55 60 65 70 90 #2=18.7 32.6 58.8 118 256 oo CASE VI. The fall of free surface-level, or loss of head, due to sudden changes of section, frictional resistance, etc., may be graphically represented as in Fig. 64. FIG. 64. Let a length of piping AE connect two reservoirs, and let h be the difference of surface-level of the water in the reser- voirs. Let L lt r l be length and radius of portion AB of pipe. T ~ it nr" " <( /- r t ZJCx " L,, r, " " " " " " CD " " L T " " " " " " JD fi 4i " " u t ,u 9 ,u t , u t be the velocities of flow in AB, BC, CD, DE, respectively. FLOW OF WATER IN PIPES. 95 The reservoir opens abruptly into the pipe at A. There is an abrupt change at B from a pipe of radius r v . to one of radius r^. There is an abrupt change at C from a pipe of radius r a to one of radius r a . At D the water flows through an orifice of area A in a dia- phragm. At E the velocity of the water as it enters the lower reservoir is immediately dissipated in eddies or vortices. Draw the horizontal plane amnop at a distance from the water-surface in the upper reservoir equal to the head due to atmospheric pressure. Draw vertical lines at A, B, C, D, E. Take ab =loss of head at the entrance A = .49 - ; = u tt due to faction from A to B =fci-j ; r> ^g r * \a^ a cd=. " " " due to change of section at B=l-^ il -*- V i I -%3 re " " " due to friction from B to C =^- a ; = " " " due to change of section at ^=.316- ; o " due to friction from C to D =^ . ^-Z " " due to change of section at D= (^ - 1) ; tk = " " " due to friction from D to E ^- ^L, ; z/ 2 kl-= " " " corresponding to u - . HYDRA UL1CS. Through / draw a horizontal plane Ix. This plane must evidently be at a distance from the water-surface in the lower reservoir equal to the pressure-head due to the atmosphere. Then the total loss of head = Ip ef + gh + M + C + re + sg+ tk, i , 2g r r % 2g ' r 9 2g 2 3 3 The broken line abcdefghkl is the hydraulic gradient. 7. Remarks on the Law of Resistance. Poiseuille's ex- periments on the flow of water through capillary tubes showed that the loss of head was directly proportional to the ve- locity. In the case of pipes used in ordinary practice the loss is undoubtedly more nearly proportional to the square of the velocity, and must be mainly due to the formation of eddies. These eddies, again, are formed more or less readily according as the water possesses less or greater viscosity. FLO W OF WATER IN PIPES. 97 The experiments of Unwin and others have shown that the surface friction is diminished by about i for every rise of 5 F. in the temperature, and it is also known that the viscosity diminishes as the temperature rises and vice versa. Reynolds has propounded a single law of resistance to the flow through pipes, which embraces the results of Poiseuille and of Darcy, and takes into account the effects of viscosity, temperature, etc. This law may be expressed in the form B n v n slope = * = ____ where d is the diameter of the pipe, A = 67,700,000, B = 396, and P= (i + .0336^ + .00022 1/ 2 ), the units being metres and degrees centigrade (/). Unwin considers that the index of the diameter d is not exactly 3 n, and should be determined independently. For a rough surface n 2, for a smooth cast-iron pipe n = 1.9, and for a lead pipe n = 1.723 ; a limitation which is analogous to that found by Froude in his experiments upon surface fric- tion. Experimenting with glass tubes, Reynolds found for veloc- ities below a certain critical velocity given by the formula that the motion of the water is undisturbed, i.e., that it was in parallel stream-lines. At and above this critical velocity eddies- are formed, and the parallel stream-line motion is completely broken up within a very short distance from the mouth of the tube. In capillary tubes = 43.79. In ordinary pipes = 278. 9 8 HYDRA ULICS. 8. Flow of Water in a Pipe of Varying Diameter. The variation in the diameter is supposed to be so gradual that the fluid filaments may still be assumed to flow in sensible parallel lines. Consider a thin slice of the moving fluid, bounded by the transverse sections AB, CD, dis- tant s and s -f- ds, respectively, from an origin on the axis of the pipe. FIG. 65. Let/ be the mean intensity of pressure, A the water area, P the wetted perimeter for the sec- tion AB. Let these symbols become / + dp, A + dA, P + dP, re- spectively, for the section CD. Let z be the height of the C. of G. of the section AB above datum. Let z -f- dz be the height of the C. of G. of the section CD above datum. Let , u + du be the velocities of flow across the sections AB, CD y respectively. Then The rate of increase of momentum of the slice ABCD in the direction of the axis f momentum generated by the effective forces acting Iupon the slice in the same direction. The acceleration in time dt = Au . dt-j- = Au . du. g dt g The total pressure on AB = p .A, and acts along the axis. The total pressure on CD = (p + dp) (A + dA\ and acts along the axis. The total normal pressure on the surface ACBD of the pipe = 27t[r-\ J \p + j A C = 2nrp . A C, very nearly. FLOW OF WATER IN PIPES. 99 The component of this pressure along the axis = 2nrpAC .sin 6 = 2 npr . dr, nearly, 6 being the angle between AC and the axis. Thus the total resultant pressure along the axis = pA - (p -\-dp\A + dA) + 2npr.dr = p.dA A.dp-}- 27rpr . dr = -A.dp, since A 7tr\ and therefore dA = 27tr . dr. The component of the weight of the slice along the axis dA\ I dA\ w sm i = \A H )/ dz= iv A . dz. The frictional resistance = P.AC. F(u) = P . ds . F(u), very nearly. Hence wAu . du ~ - = A . dp wA . dz P. ds . F(u\ o and therefore dpu.du. PF(u) y Integrating, p ,u' CP F(u} j z + w + ^+J A ~^ ds = a Constant Then 100 HYDRAULICS. The integration can be effected as soon as the relation be- tween r and s is fixed. Example. Take r = a -f- bs t and assume /and Q to be con- stant. Then _L"C__L + -r 3 / = a constant, ' w~ 2g~ b gn J r" and therefore z i <_ _|_ i _ _ a constant. ' w 2 ' - 2 4 9. Equivalent Uniform Main. A water-main usually con- sists of a series of lengths of different diameters. As a first approximation the smaller losses of head due to changes of section, etc., may be disregarded, and the calcula- tions may be further simplified by substituting for the several lengths a single pipe of uniform diameter giving the same fric- tional loss of head. Such a pipe is called an equivalent main. FIG. 66. Let /,,/, / 3 be the successive lengths of the main. Let d t , d^ , d^ be the diameters of these lengths. Let z/ v , v t , v 3 be the velocities of flow in these lengths. Let //, , h^ , h^ be the frictional losses of head in these lengths. Let Z,, d, v, h be the corresponding quantities for the equivalent uniform main. Then h = //, + h, + h, + . . . , and therefore r _ , , , , ~~ ~ 1 FLOW OF WATER IN PIPES. IOI Hence where it is assumed that /is the same for the several lengths of the main and also for the equivalent pipe. But nd* nd? nd? T V = Q = Vi = v , = c . Hence L I, / 2 / 3 an equation giving, the length L of an equivalent pipe having the same total frictional loss of head. 10. Branch Main of Uniform Diameter. In a branch main AB of length L and diameter d, receiving its supply at A. Let Q w be the way-service, i.e., the amount of water given up to the service-pipes on each side. Let Q be the end-service i.e., the amount of water dis- charged at the end B. Then it may be assumed, and it is approximately true, that the way-service per lineal foot, viz., -JT-, is constant. Thus the amount of water consumed in way-service in a length AC of the main, where BC = s, is while the total amount of water flowing across the section of the pipe at C v being the velocity of flow at C. I O2 H YDRA ULICS. Now dh, the frictional loss of head at C for an elementary length ds of the pipe, is given by the equation = 32. Integrating, the total loss of head is SPECIAL CASES. CASE I. Let <2/ be the total discharge for the same fric- lional loss of head, ^, when the whole of the way-service is stopped. Then or = & + Q.Q. + Qj f- and therefore Hence and <2/ lies between g + and Q e + 7=Q W , its mean value V 3 being & CASE II. If there is no end-service, all the water having been absorbed in way-service, Q e = o, and therefore Q' e = r= V $ and FLOW OF WATER IN PIPES. 1 03 CASE III. If Q e = o, fQ dh = vTsT^ds == elementary f fictional loss of head. Integrating between o and s, and the vertical slope, or line of free pressure, becomes a cubical parabola. CASE IV. Let the main receive its supply at A from a reservoir X in which the surface of the water is h l above datum, and let it discharge at the end B into a reservoir Fwith its surface I? above datum. Since (Q e J = Q? + C Q W + ^, therefore If Q w = TsQ.', Q e = o; and if Q w > 3 g/, then the res- ervoir Fwill furnish a portion of the way-service. Suppose that X gives the supply for the distance AO (= /,) and that Ksupplies BO (= / a ). Let z be the height above datum of the surface in a press- ure column inserted at O. Then, neglecting the loss of head at entrance, w i fQ V" = loss of head between A and O = r / a , 3 ^ a L, and J fQ 2/S = loss of head between B and O = rybi 3 n d L Also A +/, = L. 104 HYDRA ULICS. II. Nozzles. Let a pipe AB, of length / and diameter d, lead from a reservoir h ft. above the end B. First, let the pipe be open to the atmosphere at B. FIG. 68. Then (v* = n 2g I -f- head to overcome resistance due to bends, etc. = m V 2 4- head to overcome frictional resistance (= >) \ d 2gl -|- head corresponding to the velocity v in the pipe and at the outlet f= * J 4/A I ^ d } 2g Hence the height to which the water is capable of rising B v\ or, again, is =f=A-- h 4/ ^- t d r ~d Second, let a nozzle be fitted on the pipe at B. Let V be the velocity with which the water leaves the nozzle. FLO W OF WATER IN PIPES. 1 05 Let D be the diameter of the nozzle-outlet. This diameter is very small as compared with the diameter d of the pipe. But T7 V = v, 4 4 and therefore so that Fis very large as compared with i>. Also, h = head to overcome the resistance to entrance at A -\- head to overcome the resistance due to bends, etc. -f- head to overcome the frictional resistance in pipe + head to overcome the frictional resistance in nozzle (=*) V 2g ) -f- head corresponding to the velocity V with which the / V**\ water leaves the nozzle , 4/A ,F 3 . V n -f m + ^- + m' + , 2g\ dl 2g^ 2g and the height to which the water is now capable of rising at j5is v * 7 v*( . . 4/A ,F a = h (n-\- m-4- ^} m' 2g 2g\ d I 2g h Let , = //, be the pressure-head at the entrance to the w nozzle. Then the effective head at the same point Hence io6 HYDRAULICS. It will be observed that the delivery from the nozzle is less than that from the pipe before the nozzle was attached, but that the velocity-head at the nozzle-outlet is enormously in- creased. The actual height to which the water rises on leav- ing a nozzle is less than the calculated height, owing to air- resistance and to the impact of particles of water as they fall back. The force required to hold the nozzle is evidently g 4 If the water flowing through a pipe, or hose, of length / ft., with a velocity of v ft. per second, is quickly and uniformly shut off by a stop-valve t sec., the pressure in the pipe near the valve is increased by an amount - - Ibs. per square foot. <3 Of two forms of nozzle in general use, the one (Fig. 70) is a FIG. 69. FIG. 70. surface of revolution with a section which gradually diminishes to the outlet, while the other (Fig. 69) is a frustum of a cone, having a diaphragm with a small circular orifice at the outlet. Denoting the former by A and the latter by , the following- table gives the results of Ellis's experiments : Height of jet from i-inch Nozzle. Height of jet from i-mch Nozzle. Height of jet from iHnch Nozzle. Pressure in Ibs. Head in per sq. in. feet. A B A B A B 10 23 22 22 22 22 23 22 20 46 43 42 43 43 43 43 30 69 62 61 63 62 63 63 40 92 79 78 81 79 82 80 50 "5 94 92 97 94 99 95 60 , 138 108 104 112 108 "5 no 70 161 121 "5 125 121 129 123 Sc 184 131 124 137 131 142 135 90 207 140 132 148 141 154 146 IOO 230 148 136 157 149 164 155 FLOW OF WATER IN PIPES. IO/ Third, if an engine, working against a pressure of p c Ibs. per square foot, pumps Q cubic feet of water per second through a nozzle at the end of a hose / feet in length, then the pumping H.P. of the engine == . The total head at the engine end of the hose = the head corresponding to the pressure p in the hose -f~ the head re- quired to produce the velocity of flow v W 2g and this head is expended in overcoming the frictional resist- ance of the hose (all other resistances are disregarded) and in producing the velocity of flow Fat the outlet. Hence W W 2g d 2g 2g and therefore W d 2g 2g - _ JL gn* * .- Ttd 41.J-S T - since Q = v = F. 4 4 The pumping H.P. 8wff_(j_ 4/7\ -o7t*\D* d* r 12. Motor Driven by Water from a Pipe. Let the nozzle in the preceding article be replaced by a cylinder hav- ing its piston driven by the water from the pipe. Let u = the velocity of the piston per second. Let p m = unit pressure at the end of the pipe, i.e., in the cylinder. Let d m diameter of cylinder. IO8 HYDRA ULICS. Then, velocity of flow in pipe = ~jp-u* Hence , _ d^ u l , 4// d m < u* p m (other losses of head being disregarded). 13. Siphons. A siphon is a bent tube, ABCD, Fig. 71, and >- r is often employed to convey water from one reservoir to another at a lower level. Let h v // 3 , respectively, be the differences of level be- tween the top of the siphon and the entrance A and outlet D to the siphon. Then, so long as the height k 1 does not exceed the head of water (= 32.8 ft.) which measures the atmospheric pressure, the FlG * 7I> water will flow along the tube in the direction of the arrow, with a velocity v given by the equation /being the length of the tube ABCD, and all resistances, ex- .cept that due to frictional resistance, being disregarded. If //, > 32.8 feet, each of the branches AB and DC becomes a water-barometer, and the siphon will no longer work. Even when the siphon does work, an arrangement must be made for withdrawing the air which will always collect at the upper part of the siphon, 14. Inverted Siphons. The existence of a cutting or a valley sometimes renders it necessary to convey the water from a course AB to a course DE by means of an inverted siphon BCD of length. Let u be the velocity of flow in AB, and h the height of B above a datum line. FLOW OF WATER IN PIPES, 1 09 Let v be the velocity of flow in the siphon, and h t the height of D above datum. FIG. 72. Then h^ 7z a = loss of head at B -\- frictional loss of head in siphon loss of head at D = , zg d 2g "" 2g 4/7 v* = ZL -- , approximately, assuming the entrance and outlet to the siphon formed in u* v* such a manner as to considerably reduce the losses and , zg 2g and to allow of these losses being disregarded without practical error. Find, by chaining along the ground, the length of the siphon from B up to a point F not far from D. Call this length /, , and let \ be the height above datum of F, obtained with a level. Generally speaking, DF is nearly always of uniform slope. Call the slope a. Then, DF = (k^ h^) cosec a. But = h l h^ DF. sin #, an equation from which DF can be found, as /^ h^ can be determined by means of a level. 10 HYDXA ULICS. 15. Air in a Pipe. The effect of an air-bubble in a pipe ABCD may be discussed as follows: Let the air occupy the portion BC of a pipe. Let the surface of the water in the reservoir supplying the pipe be h^ ft. vertically above E, and h z ft. above D. FIG. 73. Also, let h^ be the difference of level between C and D, h t the difference of level between B and C, and / the thickness of the water-layer EF. Let H designate the head equivalent to the elastic resist- ance of the air in BC. Then, approximately, and A /! being length of portion of pipe from A to E, and /, the length from E to D. Adding the two equations, L IL f - 4/ *>* ,, , ... _ 4// v* /*,+;,,-/-. __ ( / l + /,). __, / being total length of pipe. But //! / + h< = //, h^ , very nearly. Hence an equation showing the variation of v with a variation in the height /* 4 of the space occupied by the air. Note. H o>{ course varies with the temperature. FLOW OF WATER IN PIPES. Ill 16. Three Reservoirs at Different Levels connected by a Branched Pipe. Let a pipe DO of length / x ft. and radius r l ft., leading from a reservoir A in which the water stands h l ft. above datum, divide at O into two branches, the one, OE, of length / 2 ft. and radius r 2 ft., leading to a reservoir B in which the water stands // 2 ft. above datum, the other, OF, of length / 3 ft. and radius r 3 ft., leading to a reservoir C in which the water stands h ft. above datum. I -. FIG. 74. Let v lt v^ v z be the velocities of flow in DO, OE, OF, re- spectively. Let Q lt <2 a , Q, be the quantities of flow in DO, OE, OF, respectively. Let z be the height above datum to which the water will rise in a tube inserted at the junction. Two problems will be considered, and all losses of head excepting those due to frictional resistance will be disre- garded. PROBLEM I. Given h,, h^ h z ; r lf r 9 , r 3 ; to find Q lf <2 3 , Q 3 1 z.', , z> 2 , i> 3 , and -S". fo ^ ^ a For the pipe DO, -~ = a- . . (i) and Q^Ttrfv,. . . (2) ** ^\ = V . . (3) " Q.= r: Vf (4) 112 HYDRAULICS. For the pipe OF, Z -^-^=a^*- . . (5) and Q 3 =nr 3 *v s . . . (6) Also, 1= <2,+ <2.. V. ; V ; , . (7) From these seven equations the seven required quantities can be found. In equations (3) and (7), the upper or lower signs are to be taken according as the flow is from O towards or from R towards O. This may be easily determined as follows : Assume z 7z a , and then find v l and v^ by means of equa- tions (i) and (5), and hence Q^ and <2 3 by means of equations (2) and (6). If it is found that Q l > Q 9 , then the flow is from O to E, and equations (3) and (7) become s=^ and = while if <2, < <2 3 , the flow is from E to O, and the equations are -? = <* and (-9- . It is assumed that a - is the same for each pipe. SPECIAL CASE. Fig. 75. Suppose the pipe OE closed at E. Also let r^ = r a = r a = r, and let V be the velocity of flow from A to C. The " plane of charge " for the reservoir A is a horizontal plane MQ distant from the water surface, / being the at- mospheric pressure. The " plane of charge " for the reservoir C is a horizontal plane TS distant from the water-surface. w F a In the vertical line VTQ, take TN and join MN. 2g' Then, neglecting the loss of head at entrance, MN is the FLOW OF WATER IN PIPES. 113 " line of charge," or hydraulic gradient, for the pipe DF, and is approximately a straight line. Let the " plane of charge " KK for the reservoir B, distant from the water-surface, meet MN in G. If the junction O is vertically below G, there is no head ._-- _, ^ ... 7^ >i v v v 7^^ x--* tSJnfa TIT \^ f\- - m. f 3 KTxv ^^ / x\ r*4 , f\ ;s i " / V m f i FIG. 75- available for producing flow either from E towards O or from O towards E, and hydrostatic equilibrium is established. If the junction O is on the left of G, and a vertical line OKHL is drawn intersecting KK, MN, and MQ in the points K, H, and L, there is the head Hf available for producing flow from O towards E. If the junction O is on the right of G, and the vertical line OHKL is drawn, the head HK is now available for producing flow from E towards O. Let the vertical through G meet MQ in P t and take PG = Y. Then, approximately, I, + / ~~ MN~ QN~ h,- H4 HYDRA ULICS. and therefore y - k * ~ k * j ' A+A ' '' If HL < F, the flow is from O towards . If HL > F, " " " " E " O. Again, w 2gl r and therefore, approximately, Next assume the junction O to be on the left of G, and open the valve at E. Then and Q,= Q,+ Q,, or z/ t = v, + v t . Thus y(A+ O = *,-*, = "(/^, 1 + A*,*) = " j A(.+ fJ'+Af ,' } ; and therefore ^.'(A + A) + 2/w, + A^. 1 - (A + A) ^ = o. Hence, assuming z/ a very small as compared with V, or where Q = nr* V. FLOW OF WATER IN PIPES. 115 Thus it appears that if a quantity <2 3 of water is drawn off by means of a branch from a main capable of giving a total end service Q, this end service will be diminished by j-<2 2 , \Q^ \Qv etc. according as the junction O divides the pipe DF into two portions in the ratio of I to I, I to 2, I to 3, etc. Note. The more correct value of v^ is /,+/ and the maximum value of : - L TT- does not exceed . 4 Orifice Fed by Two Reservoirs. Neglect all losses of head except the losses due to frictional resistance. FIG. 76. When the valve at is closed the flow is wholly from A to and the delivery is The line of charge (hydraulic gradient) is -M/V, where . w Il6 HYDRAULICS. Open the valve a little : a volume <2 a will now flow through <9, and a volume (2 3 mto where The " line of charge" becomes the broken line MiN. As the opening of the valve continues, the pressure-head at O diminishes, and when it is equal to /z 3 + the line of charge \sM2N, 2 N being horizontal. Hydrostatic equilibrium is now established between O and C, and the whole of the water from A passes through O, the delivery being given by Opening O still further, both reservoirs will serve the ori- fice, and the line of charge will continue to fall. When the valve is full open the "line of charge" is where 3(9 = , and the discharge is w The supply from A is equal to that from C when - 1 = *. The above investigation shows the advantage of a second reservoir in emergent cases when an excessive supply is sud- denly demanded, as, e.g., on the occasion of a fire. PROBLEM II. Given /z,, // 2 , h^\ Q 2 , Q s , and therefore Q t (= a+G 3 ); tofirtdr I ,r;,r t ,f> l ,9 t ,f'.,jr. As before, let z be the pressure-head at O. Then ... (i) and <2, = >,; ... (2) (3) " e, = *r,' Vt ; ... (4) ... (5) " C. = ^>.. . . . (6> FLOW OF WATER IN PIPES. 1 1/ These six equations contain the seven required quantities, viz., r 1 , r a , r 5 , z\ , v 9 , v t , and z. Thus a seventh equation must be obtained before their values can be found. This equation is given by the condition " that the cost of the piping laid in place should be a minimum/' it being assumed that the cost of a pipe laid in place is proportional to its diameter. Hence l l r l + 4 r a + 4 r a a minimum (7) From equations (i) and (2), - L j-^; (3) (4), - -^ " (5) " (6), *-^ b = ^ 3 rr, Differentiating these three equations, dz _ $aQ* , t But by equation (7) /X^ -|_ l^dr^ + / 3 ^r 3 = O. Hence 6 6 which is the seventh equation required. Il8 HYDRAULICS. This last equation may be written in the forms and a = a v a ^ 3 z/, 1 ^ 3 3 " 17. Mains with any Required Number of Branches. Let there be n junctions and m pipes. Let h l , // a , . . . h m be the m pressure-heads at the end of each successive length of pipe. Let #!,,,...# be the n pressure-heads at the 1st, 2d y 3d, . . . 72th junctions. Let /!,/,,.../, be the lengths of the ;// pipes. PROBLEM I. Given //, , h^ , . . . h m , r l , r 2 , . . . r m , to find ?i;tf 9 ,. *,*, *,.** _i. ^ nz > 2 There are 772 equations of the type - - - = a. Also, the quantity flowing through the first portion of the main is equal to the sum of the quantities flowing through all the branches at the first junction, and an analogous equation will hold for each of the remaining n I junctions. Thus n additional equations are obtained. From these m -f- n equations, v l , v z , . . . v m , z^ , ^ , . . . z n may be found analytically or by the method of repeated ap- proximation. PROBLEM II. Given h^ , h^ , . . . h m , Q l , <2 2 , . . . Q m , to find There are now only m equations of the type -[- h If % _ V* ~T~ a ~r ' involving m -\- n unknown quantities, and the problem admits of an infinite number of solutions. It is therefore assumed that the cost of the piping laid in place is to be a minimum. Thus n new equations are ob- FLO W OF WATER IN PIPES. tained, and the m -\- n equations may be solved analytically or by repeated trial. 18. Variation of Velocity in a Transverse Section. Assumption. That the water in any portion of a pipe is made up of an infinite number of hollow concen- tric cylinders of fluid, each moving parallel to the axis with a certain definite velocity. Let u be the velocity of one of these cyl- inders of radius x and thickness dx. Then the flow across a transverse section is given by the equation FlG ?7 dq = 2nx dx . u, and the total flow Q27tl uxdx, (i) r being the radius of the pipe. If v m be the mean velocity for the whole transverse section of the pipe, nr* (2) Again, assuming with Navier that the surface resistance between two concentric cylinders is of the nature of a viscous resistance and may be represented by k per unit of area at dx the radius x, k being a coefficient called the coefficient of vis- cosity, then the total resistance at the radius x for a length ds of the cylinder , du du = 2nx . ds . k - = 2nk . as . x--. dx dx The total resistance at the radius x -f- dx du , < I2O HYDRAULICS. Hence the total resultant resistance for the length ds of the cylinder under consideration = 2nkds The component of the weight of the slice of the cylinder in the direction of the axis = w . 2nx . dx . ds . sin 0, being the inclination of the axis to the horizon. Let dz be the fall of level in the distance ds. Then dz = ds . sin 6. Therefore, component of weight in direction of axis = w . 2nx dx . dz. The resultant pressure on the slice in the direction of motion = P (P ~\~ d) . znx . dx = 2nx .dx .dp. Then, since the motion is uniform, w . 2nk . ds . r\x-\dx w . 2rtx .dx.dz 2nx .dx .dp o, dx\ dxi and therefore k . ds d f du\ dp /t* _ ,/V ty f. -T-U-H - dz - - = o. x ax\ ax I w Integrating only for the cylinder under consideration, ks d f du\ ( p\ -- \ x -r] (z + ) = a constant. x ax\ ax I \ w' But z + is evidently independent of x, and is a linear w function of s (Art. 2, Chap. III.). Hence I d I du\ -- (x = a constant = A, suppose. x dx\ dx> FLO W OF WATER IN PIPES. 121 Therefore d I du\ -(,-J = ^.. . . . . . .. (3) Integrating, du x* x- = A \- B. dx 2 Assuming that the central fluid filament is the filament of maximum velocity, then when x = o, - is also nil. Therefore B = o and x^ = Ax\ dx and therefore Integrating, u = A- + C. 4 Let w max be the velocity of the central filament, i.e., the value of u when x = o. Then (5) where D = . 4 Again, by equation I, Q = 27tJ (# max Dx^x.dx = 7rr*\2 max J : and by equation 2, Dr* v m = u max (6) 2 If ;/, = surface velocity, then, by equation 5, U, = max - DS (7) Hence, by equations 6 and 7, u s + inax zv m (8) 122 i v ,-r y U '$*' V-M ' f5 T : 5 : < *^ c ( - ^ ULICS. ' -1 -^ I ^\: i^ EXAMPLES. r sec. / 4 ft. in J 1. A water-main is to be laid with a virtual slope of i in 850, and is to give a maximum discharge of 35 cubic feet per second. Determine the requisite diameter of pipe and the maximum velocity, taking/ = .0064. Q V- Ans. 3.679 ft.; 3.2888 ft. per sec. 2. Find the loss of head due to friction in a pipe : diameter of pipe / = 12 in., length of pipe = 5280 ft., velocity of flow = 3 ft. per second ; *^ f = .0064.; Also find the discharge. Ans. 19.008 ft. ; 2.3562 cub. ft. per 3. A pipe has a fall of 10 ft. per mile ; it is 10 miles long and diameter. Find the discharge, assuming/ = .0064. Ans. 54.7 cub. ft. per sec. 4. A pipe discharges 250 gallons per minute and the head lost in fric- tion is 3 ft.. Find approximately the head lost when the discharge is 300 gallons per minute ; also find the work consumed by friction in both cases. Ans. 4.32(1.; 7500 ft.-lbs. ; 12,960 ft.-lbs. 5. What is the mean hydraulic depth fn 1 a circular pipe when the diameter water rises to the height -- -=- above the centre ? 2^2 I0 Ans. x diameter. 6. A 12-inch pipe has a slope of 12 feet per mile; find the discharge. (/=.oo5.) Ans. 2. ii^ cub. ft. per sec. 7. The mean velocity of flow in a 24-in. pipe is 5 ft. per second ; find its virtual slope,/ being .0064. Ans. i in 200. 8. Calculate the discharge per minute from a 24-in. pipe of 4000 ft, length under a head of 80 ft., using a coefficient suitable for a clean iron pipe. Ans. 34.909 cub. ft. per sec. 9. How long does it take to empty a dock, whose depth is 31 ft, 6 ins. and which has a horizontal sectional area of 550,000 sq. ft., through two 7-ft. circular pipes 50 ft. long, taking into account resistance at en- trance ? Ans. 214 min. 6 sec. 10. The virtual slope of a pipe is i in 700; the delivery is 180 cubic -feet per minute. Find the diameter and velocity of flow. Ans. 1.26 ft.; 2.401 ft. per sec. n. Determine the diameter of a clean iron pipe, 100 feet in length, which is to deliver .5 cub. ft. of water per second under a head of 5 feet.. Assume/ = .006. Ans. .326ft. FLOW OF WATER IN PIPES. 123 12. A reservoir has a superficial area of 12,000 ft. and a depth of 60 ft. ; it is emptied in 60 minutes through four horizontal circular pipes, equal in diameter and 50 ft. long. Find the diameter. Ans. 1.75 ft. Explain how the total head is made up, and draw the plane of charge. 13. A 3-inch pipe is very gradually reduced to i inch. If the press- ure-head in the pipe is 40 ft., find the greatest velocity with which the water can flow through. Ans. 1.4 ft. per sec. 14. Water flows through a 24-inch pipe 5000 yards in length. At 1000 yards it yields up 300 cubic feet per minute to a branch. At 2800 yards it yields up 400 cubic feet per minute to a second branch. At 4000 yards it yields up 600 cubic feet per minute to a third branch. The delivery at the end is 500 cubic feet per minute. Find the head absorbed by friction. (/=.oo75.) Ans. 176.801 ft. 1 5. Find the H. P. required to raise 550 gallons per minute to a height of 60 feet, through a pipe 100 feet in length and 6 in. in diameter, the coefficient of friction being .0064. , Ans. 10.74. 1 6. What head of water is required for a $-in. pipe, 150 ft. in length,, to carry off 25 cub. ft. of water per minute ? Ans. 1.56223 ft. What head will be required if the pipe contains two rectangular knees? Ans. 1.84918 ft. 17. Determine the delivery of a 2- in. pipe, 48 ft. long, under a 5-ft. head. Ans. .1349 cub. ft. per sec. What will be the delivery if the pipe has five small curves of 90 cur- vature, the ratio of the radius of the pipe to that of the curves being 1:2? Ans. .1327 cub. ft. per sec. 1 8. The curved buckets of a turbine form channels 12 in. long, 2 in. wide, and 2 in. deep; the mean radius of curvature of the axis is 8 in. the water flows along the channel with a velocity of 50 ft. per minute. What is the head lost through curvature ? Ans. .00138 ft. 19. Find the maximum power transmitted by water in a 36-inch pipe, the metal being \\ inches thick and the allowable stress 2800 Ibs. per square inch. If the pipe is \\ miles in length, find the loss of power. Ans. 576 H. P. ; 720.2 ft. -Ibs. 20. Find the diameter of a pipe \ mile long to deliver 1500 gallons of water per minute with a loss of 20 feet of head. (/ = .005.) Ans .1.0135 ft- 21. Water is to be raised 20 ft. through a 3Or half the length ; (3) by dividing it into three sections of equal length, of which the radii are r, , and , .respectively ? (f = coefficient of friction.) Ans. i. New discharge = r + 64/A1 -4 9' + 4// 4 228//y 25. A 24-inch pipe 2000 ft. long gives a discharge of Q cubic feet of water per minute. Determine the change in Q by the substitution for the foregoing of either of the following systems : (i) two lengths, each of looo ft., whose diameters are 24 in. and 48 in. respectively; (2) four lengths, each of 500 ft., whose diameters are 24 in., 18 in., 16 in., and 24 in. Draw the " plane of charge " in each case. Ans. (i) Discharge is increased 33.2 per cent taking loss at change of section into account; Discharge is increased 35.7 per cent disregarding loss at change of section. (2) Discharge is diminished 45 per cent disregarding losses at change of section. 26. Q is the discharge from a pipe of length / and radius r \ examine the effect upon Q of increasing r to nr for a length ml of the pipe. * Ans. New discharge = Q (n* - i) 2 * 27. A reducer, I ft. in length, discharges at the rate of 400 gallons per minute, and its diameter diminishes from 12 in. to 6 in.; find the total loss of head due to friction. Ans. .0055297. 28. A reservoir of 10,000 square feet superficial area and 100 feet / deep discharges through a pipe 24 in. in diameter and 2000 feet long. * Find the velocity of flow in the pipe. What should be the diameter of the pipe in order that the reservoir might be emptied in two hours ? Ans. 15.36 ft. per sec.; 3.67 ft. 29. Eight cubic feet of ore is to be raised at the rate of 900 ft. per . FLOW OF WATER IN PIPES. 12$ minute by a water-pressure engine with four single acting cylinders of 6 in. diameter and 18 in. stroke, making 60 revolutions per minute. Find the diameters of a supply- pipe 230 ft. long for a head of 230 ft., disregarding friction of machinery, etc. Ans. 4 in. 30. A 2-inch pipe A suddenly enlarges to a 3-inch pipe B, the quan- tity of water flowing through being 100 gallons per minute. Find the loss of head and the difference of pressure in the pipes (i) when the flow is from A to B ; (2) when the flow is from B to A. Ans. (i) Loss of head = 8.639 i n - Gain of pressure-head = 13.83 " (2) Loss of head = 7.428 " Diminution of pressure-head = 29.88 " 31. A 3-inch horizontal pipe rapidly contracts to a i-inch mouih- piece, whence the water emerges into the air, the discharge being- 660 Ibs. per minute. Find the pressure in the 3-inch main. If the 3-inch pipe is 200 ft. in length and receives water from an open tank, find the height of the tank. Ans. 1003.5 Ibs. P er sq. ft.; 19.92 ft. 32. The efficiency of an engine is f ; it burns 8 Ibs. of coal per hour per H.P., and works 8 hours a day for 300 days in the year; the cost of the engine is $12.00 per H.P., and the cost of the coal is $3.00 per ton ; 4500 gallons of water per minute have to be raised a height of 200 ft. through a pipe of which the diameter is to be a minimum. Cost of piping = $> per lineal foot, D being the diameter. Find the value of D. Ans. 2.923 ft. 33. A reservoir is to be supplied with water at the rate of 11,000 gallons per minute, through a vertical pipe 30 ft. high; find the minimum diameter of pipe consistent with economy. Cost of pipe per foot = &/, d being the diameter; cost of pumping = i cent per H.P. per hour; original cost of engine per H.P. = $100.00; add 10 per cent for depreciation. Engine works 12 hours per day for 300 days in the year. Ans. 4.375 ft. 34. A horizontal pipe 4 in. in diameter suddenly enlarges to a diameter of 6 in.; find the force required to cause a flow of 300 gallons of water per minute through the sudden enlargement. Ans. .06 H.P. 35. 1000 gallons per minute is to be forced through a system of pipes AB, BC, CD, of which the lengths are 100 ft., 50 ft., 120 ft., and the radii 4 in., 6 in., and 3 in., respectively. Draw the plane of charge. Ans. Loss in friction from A to B = 111.96 ft.; loss at B 4.499 ft.; " " " " B to C 7.372 " " " C 14.56 " " " " " C to D = 566.17 " 126 HYDRA ULICS. 36. A pipe 4 in. in diameter suddenly contracts to one 3 in. in diameter; find the power necessary to force 250 gallons per minute through the sudden contraction. Ans. 1.23997 H.P. 37. If a pipe whose diameter is 8 in. suddenly enlarges to one whose diameter is 12 ins., find the power required to force 1000 gallons per minute through the enlargement, and draw to scale the plane of charge. Ans. Energy expended = .1377 H.P. 38. 1000 gallons per minute are forced through a system of pipes AB, BC, CD, of which the lengths are 100 ft., 50 ft., and 120 ft., and the radii 6 in., 3 in., and 4 in., respectively. Draw to scale the plane of charge. Ans. Loss in friction from A to B = 14.744 ft.; loss at B = 14.56 ft. " " " " B to c = 235.9 " ; " " c= 8.819" " " " " CtoZ>= 134.36 " 39. Water flows from a 3-inch pipe through a i^-inch orifice in a diaphragm into a 2-inch pipe. What head is required if the delivery is to be 8 cubic feet of water per minute ? Ans. 2.826 ft. 40. 500 gallons of waiter per minute are forced through a continuous line of pipes AB, BC, CD, of which the radii are 3 in., 4 in., 2 in., and the lengths 100 ft., 150 ft., and 80 ft., respectively. Find the total loss of head (a) due to the sudden changes of form at B and C, (b) due to friction. Find (c) the diameter of an equivalent uniform pipe of the same total length. Ans. (a) .1378 ft.; 1.152 ft. (b) 3.688 ft. in AB; 1.313 ft. in BC\ 22.393 ft - in CD. (c) .4212 ft. 41. AB, BC, CD is a system of three pipes of which the lengths are looo ft., 50 ft., and 800 ft., and the diameters 24 in., 12 in., and 24 in., respectively; the water flows from CD through a i-inch orifice in a thin diaphragm, and the velocity of flow in AB is 2 ft. per second. Draw the plane of charge and find the mechanical effect of the efflux. Ans. Loss at B = -& ft.; at C = -/fa ft.; in friction from A to B = .8 ft. ; from B to C = 1.28 ft.; from C to D = .64 ft. ; energy of jet = 14,81 if H.P. 42. looo gallons per minute flows through a sudden contraction from 12 inches to 8 inches at A, then through a sudden enlargement from 8 inches to 12 inches at B, the intermediate pipe AB being 100 ft. long. Draw the plane of charge. Ans. Loss at A = .288 ft. ; at B = .281 ft. ; in friction from A to B = 3.499 ft. 43. Water flows from one tube into another of twice the diameter; the velocity in the latter is 10 ft. Find the head corresponding to the resistance. Ans. 14.0625 ft. FLOW OF WATER IN PIPES. 12 7 44. In a given length / of a circular pipe whose inner radius is r and thickness %BO. (x If AO = 2000 ft'., the pressure-head at O =161 ft.; 2/1 = 14.9 ft. ; 2/2 = 3.02 ft.; v* = 14.18 ft. If AO=4ooo ft., the pressure-head at 0=96 ft.; 2/1=13.8 ft.; 2/2 = 6.7 ft.; 2/3 = 15.4 ft. 64. A pipe 24 in. in diameter and 2000 ft. long leads from a reservoir in which the level of the water is 400 ft. above .datum to a point B, at which it divides into two branches, viz., a 12-in. pipe J3C, 1000 ft. long, leading to a reservoir in which the surface of the water is 250 ft. above datum, and a branch BD, 1500 ft. long, leading to a reservoir in which the surface of the water is 50 ft. above datum. Determine the diameter of BD when the free surface-level at B is (a) 300 ft., (b) 250 ft., and (c) 200 above datum. Ans. (a) 1.454 ft.; (&) 1.783 ft.; (c) 2.096 ft. 65. Two reservoirs A and B are connected by a line of piping MON, 2000 ft. in length. From the middle point O of this pipe a branch OP, looo feet in length, leads to a reservoir C. The reservoirs A and (Tare 200 feet and 100 feet, respectively, above the level of C. The deliveries in MO, OP, ON, in cubic feet per second, are ^-it, ^-TT, and TC, respec- 1 30 HYDRA ULICS. lively. Find (a) the velocities of flow in MO, OP, ON\ (b] the radii of these lengths; (c) the height of the free surface-level at O above C. Ans. (a) 1 1. 121 ft. per sec. in MO; 10.158 ft. per sec. in OP; 14.145 ft. per sec. in OA r . (b) .49976^.; .41831 ft.; .26588 ft. (c) 150.5 ft., very nearly. 66. A main, 1000 ft. long and with a fall of 5 ft. discharges into two branches, the one 750 ft. long with a fall of 3 ft., the other 250 ft. long with a fall of i ft. The longer branch passes twice as much water as the other and the total delivery is 47^ cu. ft. per minute. The velocity of flow in the main is i\ ft. per second. Find the diameters of the main and branches. Ans. .63245 ft.; .288ft.; 488ft. 67. How far can 100 H.P. be transmitted by a 3^ in. pipe with a loss of head not exceeding 25 per cent under an effective head of 750 Ibs. per square inch ? Ans. 5426.3 ft. 68. A city is supplied with water by means of an aqueduct of rect- angular section, 24 ft. wide, running 4 ft. deep, and sloping i in 2400. One-fourth of the supply is pumped into a reservoir through a pipe 3000 ft. long, rising 25 ft. in the first 1500 ft., and 75 ft. in the second 1500 ft. The pumping is effected by an engine burning 2| Ibs. or coal per H.P. per hour, and working constantly through the year. A percentage is to be allowed for repairs and maintenance; the cost of the coal per ton of 20oolbs. is $4 ; the prime cost of the engine is $100 per H.P. ; the effi- ciency of the engine is f ; the coefficient of pipe friction is .0064, the cost of the piping is $30 per ton. Determine the most economical diam- eter of pipe, and the H.P. of the engine. Ans. 4.84 ft. ; 456.455 H.P. CHAPTER IV. FLOW OF WATER IN OPEN CHANNELS. i. Flow of Water in Open Channels. A transverse sec- tion of the water flowing in an open channel may be supposed to consist of an infinite number of elementary areas represent- ing the sectional areas of fluid filaments or stream-lines. The velocities of these stream-lines are very different at different points of the same transverse section, and the distribution of the pressure is also of a complicated character. Generally speaking, the side and bed of a channel exert the greatest retarding influence on the flow, and therefore along these surfaces are to be found the stream-lines of minimum velocity. The stream-lines of maximum velocity are those farthest removed from retarding influences. If the stream-line velo- cities for any given section are plotted, a series of equal velocity-curves may be obtained. In a channel of symmetrical FIG. 78. section, the depth of the stream-line of maximum velocity below the water-surface is less than one fourth of the depth of the water, while the mean velocity-curve cuts the central vertical line at a point below the surface about three fourths of the depth of the water. In the ordinary theory of flow in open channels, the variation of velocity from point to point in a transverse section is disregarded, and it is assumed that all the stream-lines are sensibly parallel and move normally to the section with a common velocity equal to the mean velocity of the stream. With this assumption, it also necessarily follows that the 131 132 HYDRA ULICS. distribution o 4 " pressure over the section is in accordance with the hydrostatic law. Again, it is assumed that the laws of fluid friction already enunciated are applicable to the flow of water in open chan- nels. Thus, the resistance to flow is proportional to some function of the velocity (F(v)}, to the area (S) of the wetted surface, is independent of the pressure, and may be expressed by the term S.F(v). An obvious error in this assumption is that v is the mean velocity of the stream and not the velocity of the stream-lines along the bed and sides of the channel. In practice, however, the errors in the formulae based upon these imperfect hypotheses are largely neutralized by giving suitable values to the coefficient of friction (/). When a constant volume (Q) of water feeds a channel of given form, the water assumes a definite depth. A permanent regime is said to be established and the flow is steady. If the transverse sectional area (A) is also constant, then, since Q = vA, the velocity v is constant from section to section and the flow is said to be uniform. Usually the sectional area A is variable and therefore the velocity v also varies, so that the motion is steady with a varying velocity. Any convenient short stretch of a channel, free from obstructions, may be selected, and treated without error of practical importance, as being of a uniform sectional area equal to that of the mean section for the whole length under consideration. 2. Steady Flow in Channels of Constant Section (A). The flow is evidently uniform ; and since A is constant, the depth of the water is also con- stant, so that the water-surface is parallel to the channel-bed. JH Consider a portion of the stream, of length /, between the two transverse sections aa, bb. Let i be the inclination of the bed (or water-surface) to the horizon. Iff FIG. 79. Let Pbe the length of the wetted perimeter of a cross-section. FLOW OF WATER IN OPEN CHANNELS. 133 Then, since the motion is uniform, the external forces acting upon the mass between aa and bb in the direction of motion must be in equilibrium. These forces are : (1) The component of the weight of the mass, viz., wAl sin i = wAli = wAl = wAk, h being the fall of level in the length /. Note. When i is small, as is usually the case in streams, -j = tan / = sin / = /, approximately. (2) The pressures upon the areas aa and bb, which evi- dently neutralize each other. (3) The frictional resistance developed by the sides and bed, viz., Hence wAh - PlF(v) = o, or FM Ah -^ = w= m *> m being the hydraulic mean depth. It now remains to determine the form of the function F(v). In ordinary English practice it is usual to take W 2g f being the coefficient of friction. Then or jig v = \i ~~f y mi = cy mi. 1 34 H YDRA UL1CS. c being a coefficient whose value depends upon the roughness of the channel surface and upon the form of*its transverse section. Prony and Eytelwein adopted the formula F(v] - = av -j- bv* = mi, w and carried out different experiments to determine the values of a and b. According to Prony, - = 22472.5 and -r 10607.02, " "Eytelwein, - = 41688.02 " -= 8975.43. For a velocity of about 70 ft. per minute Prony's and Eytelwein's results give the same value for mi. For other velocities, Prony's values of mi are greater or less than those of Eytelwein, according as the velocity v is greater or less than 70 ft. If v, however, does not differ very widely from 70 ft., the change of value is small and of no practical importance. For values of v exceeding 20 ft. per minute the term av may be disregarded without practical error, and the formula then becomes mi = bv* t or Hence v = 105 \/mi f according to Prony, and v = 95 ^/mi, according to Eytelwein, giving as a mean v = loo^mz, which is Beardmore's formula. The total head H in a stream is made up of two parts, the FLOW OF WATER IN OPEN CHANNELS. 13$ one required to produce the velocity of flow, and the other absorbed by the frictional resistance. Thus, 2g ?;/ w In long canals, and in rivers with slopes not exceeding 3 ft. v* per mile, the term is very small as compared with the term / Mv) , and may be disregarded without sensible error. m w Note. The retarding effect of the air upon the free surface of a stream or river has yet to be determined by careful observation and experiment. It may, however, be assumed that the resistance offered by calm air per unit of free surface is approximately one tenth of the resistance offered by similar units at the bottom and sides of smooth channels. Thus, in smooth channels, if X is the width of the free surface, the Y wetted perimeter is more correctly P -\- In general, the wetted perimeter may be expressed in the form P -f- -ip ft being 10 for smooth channels and greater than 10 for rough channels. The value of ft is obviously diminished by opposing winds and increased by following winds. 3. On the Form of a Channel. In the formula F(v) mt ' = *T' = J and ilss -yl are similarly related in the deter- mination of v, the mean velocity of flow. If v is constant, the product mi must also be constant, so that if m increases i must diminish, and vice versa. Thus, in a very flat country the flow may be maintained by making m sufficiently large, while again if the channel-bed is steep m is small. 136 HYDRAULICS. The erosion caused by a watercourse increases with the rapidity of flow. At the same time the sectional area (A) of the waterway also increases, so that the velocity of flow v diminishes. Thus there is a tendency to approximate to a " permanent regime " when the resistance to erosion balances the tendency to scour. Hence, throughout any long stretch of a river, passing through a specific soil, the mean velocity of flow will be very nearly constant if the amount of flow (Q) does not vary. Gen- erally speaking, the volume conveyed by a river increases from source to mouth on account of the additions received from tributaries, etc. Since Q increases, A must also increase ; and if mi or v is to remain constant, i must diminish. It is also observed that the surface slopes of large rivers diminish gradu- ally from source to mouth. Again, various problems relating to the proper sectional form of a channel may be discussed by means of the formulae 'A . v = c \mi = and Suppose the slope to be constant. Then A v* is proportional to 75 and A 9 Q* is proportional to p. ' PROBLEM I. The section of the waterway being a rectangle of width x and depth y, and of given area (A xy\ it is required to find the ratio of x to y for which the velocity of flow (v) will be a maximum. Then dv = o, and therefore P.dAA.dP P* ' FLO W OF WATER IN OPEN CHANNELS. 137 Hence PdA-AdP=o. But dA o = xdy -\-ydx, and therefore also ^ dP o = dx + 2dy, J since P=x + 2y. Hence, FIG. 80. and the mean hydraulic depth _ A _ xy _y ~P~x-\-2y~2 = one half of the depth of the water. The same results follow if the discharge Q instead of v is to be a maximum. In such case dQ M'\ - o - d\-p) = .dA - A*.dP and therefore $PdA AdP o. But dA = o, and therefore dP = o. Hence, etc. Note. The same results also follow if, instead of A being given, the wetted perimeter P is to be a minimum, since then dP = o, and therefore also dA = o. PROBLEM II. The waterway being trapezoidal in section, FIG. 81. of bottom width x, depth y, and sides sloping at a given angle Q to the horizontal, it is required to find the ratio of x to y which, for a given wetted perimeter (P^ or area (A), will make the velocity of flow or the discharge a maximum. HYDRAULICS. As in Problem I, dA = o and dP = o. But A = (x +y cot 6)y and P = x + 2y cosec 6. Hence ^4=0= ydx + <^/O + 27 cot 0) and

e \s c being a constant of integration. 1 4 H YDRA ULICS. When x = o, y = a, and therefore o = log, (a + *V - M 2 ) + c = log e & + c, where b = a -f- j/# a ^ 2 . Hence = te or is the equation to the required profile, which, as may be easily shown, is a curve which flattens very rapidly. PROBLEM IV. If water flows through a circular aqueduct, find the angle 6 subtended at the centre by the wetted perim- eter, for which the velocity of flow is a maximum. Let r = radius of aqueduct. Area of waterway = (6 sin 0). Wetted perimeter = r6. FIG. 85. Then r sin r I sin m = -ft 2 C^ sin Now v is to be* a maximum and therefore -^ must be a minimum. Hence cos - sin , /sin 0\ 0COS 4)==- and therefore cos sin = o. Hence 6= tan 0, and the angle in degrees is about 77 27'. FLOW OF WATER IN OPEN CHANNELS. 141 Also, the mean hydraulic depth = ^i --- -r ) = - (i _ cos d) = rsin = . 39 X r. PROBLEM V. A channel of given slope has a given surface- width AC, vertical sides AB (=y l ) and CD (=7,) of given depths, and a curved bed BD (= L) of given length. FIG. 86. The amount and velocity of flow in the channel will be a maximum when the form of the bed BD is a circular arc. This can be easily proved as follows : Since the slope is constant, v oc /~A a \f -p. But P (= L -\- y^ + >0 is a constant quantity, and therefore v and also Q will be a maximum when ^4 is a maximum. Hence, too, the area between the chord BD and the curve must be a maximum, and therefore the curve must be a circu- lar arc. The proof of this by the Calculus of Variations is as follows : Take O in CA produced as the origin, OC as the axis of x, and the vertical through O as the axis of y. Then ydx is to be a maximum. 1 42 H YD RA ULICS. Also, dy is a given quantity, OA being = JT, , OC = x^ , and Hr Let V = y -f- a Vi -\- p\ a being some constant. Then /*; / F. dx is to be a maximum, *X ^i and therefore that is, and thus ^ + ~^=r^ = ^ Therefore ^ / ' Va* - (c, - max v b to be constant. Take IllUX O sy 7 y \ / IS 2 HYDRA ULICS. Then the general equation (4) of the velocity-curve becomes . .... (9) Now if Y' o, i.e., if the filament of maximum velocity is in the free surface, H P =tw-A^. But in such case Bazin's experiments led to the relation Hence ^=36.3 and the general equation of the velocity-curve becomes ^iv- FV ..... (10) This is Bazin's formula, and it agrees well with his experi- ments on artificial channels and also with the results of experiments on the Saone, Seine, Garonne, and Rhine. It was found that *7) - 1.17 in the Rhine at Basle and ranged from i.i to 1.13 ^ in the others; 36.3^ i/ . (h _ KY" y between r 3 and 2 o; Y = .33 in some artificial channels and ranged from O to 0.2 in the other cases ; *W v b ranged from Jz/ max to i^ max . These results are not in agreement with the Mississippi measurements. FLOW OF WATER IN OPEN CHANNELS. 153 Note. When the filament of maximum velocity is in the free surface, Y = o, and therefore, by equation 5, wih* TI 191 u m ^max ~~" z- 7 > and by equation 8, wit? Hence, combining these two equations, Boileau assumes that the velocity-curve is given by the equation .. ..... (12) below the filament of maximum velocity, being MM l in Fig. 91, and by the equation v = a-Bf + Cy (13) above the filament of maximum velocity, being MM 9 in Fig. 92. Let v s be the surface-velocity, i.e., the value of v when y o. Then, by equa- tion 13, v s = a. Also, the two equations (12) and (13) must each give the same value for the maximum velocity (zw), and therefore A - BY* = v max = a BY* + CY, FlG ' 9I> from which A a A v s Again, taking A = z/ max + ^ Boileau deduced experimen- tally that d is sensibly constant for different streams. 1 54 HYDRA ULICS. But A = ?w + d = A - B Y* + d, and therefore B Hence Boileau's equation becomes for points below the filament of maximum velocity, and V = V, - *' + (ZW + d- . for points above the filament of maximum velocity. 8. Relations between Surface, Mean, and Bottom Ve- locities. Bazin deduced from his experiments on canals the relation , v m v m = v s 25.4 Vmt = v s 25.4, where c V -~. Therefore cv s v m = - c+ 25.4 Darcy and Bazin give the relation 10.87 ^wt = v b + 10.87. Therefore v = ~ C I0.8/ A mean value of c is 45.7, which makes v m = 1.312. zv ib Dubuat gives the following table of maximum bottom velocities consistent with stability : FLOW OF WATER IN OPEN CHANNELS. 155 Nature of Canal Bed. Vj,. Soft earth 0.25 Loam 0.50 Sand i.oo Gravel 2.00 Pebbles 3.40 Broken stone, flint 4.00 Chalk, soft shale 5.00 Rock in beds 6.00 Hard rock , 10.00 TABLE OF MAXIMUM VELOCITIES FROM INGENIEURS TASCHENBUCH. Nature of Canal-bed. v s v m vb 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 TABLE OF VISCOSITY OF WATER AND MERCURY. (From Everett's System of Units.) WATER. MERCURY. Temp. (Cent.) o 5 10 15 20 25 30 Viscosity. .Ol8l .0154 0133 .0116 .OIO2 .OOQI .0081 Temp. (Cent.) 35 40 45 50 60 80 90 Viscosity. .0073 .0067 .0061 .0056 .0047 .0036 .0032 Temp. (Cent.) O u IO 18 99 154 197 249 Viscosity. .0169 .0162 .0156 .0123 .0109 .OIO2 . 00964 Temp. (Cent.) 315 340 Viscosity. .00918 .00897 156 HYDRA ULICS. The viscosity is given by _.oi83 and by .0369^ , according to Meyer ; j| .00131, according to Slotte; / being the temperature centigrade. 9. Flow of Water in Open Channels of Varying Cross- section and Slope, Assumptions. (a) That the motion is steady. Thus the mean velocity is constant for any given cross- section, but varies gradually from section to section. (b) That the change of cross-section is also gradual. (c) That, as in cases of uniform motion, the work absorbed by the frictional resistance of the channel-bed and sides is the only internal work which need be taken into consideration. Xy FIG. Q2. Let Fig. 92 represent a longitudinal section of the stream. The fluid molecules which are found in any plane section db at the commencement of an interval will be found in a curved surface dc at the end of the interval, on account of the differ- ent velocities of the fluid filaments. Suppose that the mass of water bounded by the two trans- verse sections ab, ef, comes into the position cdhg in a unit of time. Then the change of kinetic energy in this mass is equal to the algebraic sum of the work done by gravity, of the work done by pressure, and of the work done against the frictional resistance. Change of Kinetic Energy. This is evidently the difference FLOW OF WATER IN OPEN CHANNELS. I 57 between the kinetic energies of the masses efgh and abed, since, as the motion is steady, the kinetic energy of the mass between cd and ef remains constant. Let A 1 be the area of the cross-section ab. " j " " mean velocity across this section. *' v " " velocity at this section of any given fluid, filament of sectional area a. Let v = u l V. Then Aji, = 2(av) and 2(aV) = O. The kinetic energy of the mass abed Since S(aV) o and 3 x V 2, + v. Now 2j + v is evidently positive. Hence the kinetic en- ergy of the mass abed a' being a coefficient of correction whose value depends upon the law of the distribution of the velocity throughout the sec- tion ab. It is positive and greater than unity. Assume that a has the same value for the sections ab and ef. Then if A^ 158 H YDRA ULICS. # a , are the area and mean velocity at the transverse section ef, the kinetic energy of the mass efgh = h. is now negative, and the depth diminishes in the direc- ds tion of flow. Up-stream, h increases and approaches H in value, so that MN is asymptotic to PQ. FLOW OF WATER IN OPEN CHANNELS. 163 Down-stream, h diminishes, u increases, and therefore the value of is more and more nearly equal to unity, gh Thus, in the limit, the denominator in equation 8 becomes zero, and therefore = 00. Hence theory indicates that at a as certain point down-stream the surface line MN takes a direc- tion which is at right angles to the general direction of flow. This is contrary to the fundamental hypothesis that the fluid filaments flow in sensibly parallel lines. In fact, before the FIG. 95. limit could be reached this hypothesis would cease to be even approximately true, and the general equation would cease to be applicable. This form of water-surface is produced when there is an abrupt depression in the bed of the stream. Fig. 96 shows one of the abrupt falls in the Ganges canal as at first constructed. The surface of the water flowing freely FIG. 96. over the crest of the fall took a form similar to MN below the line PQ.oi uniform motion. The diminution of depth in the approach to the fall caused an increase in the velocity of flow, with the result that for several miles above the fall a serious 164 HYDRA ULICS. erosion of the bed and sides took place. In order to remedy this, temporary weirs were constructed so as to raise the level of the water until the surface-line assumed a form MN' cor- responding approximately to PQ. In some cases the water was raised above its normal height and a backwater produced* CASE III. au* > gh and H < h. - is negative and the surface-line of the stream is wholly above PQ. FIG. 97. dk If h gradually increases, u diminishes and j- approximates to i in value. If h gradually diminishes it approximates to H in value, dk and in the limit -T~= o. ds Between these two extremes there is a value of h for which the denominator of equation 8 becomes nil, viz., and the corresponding value of -y- is infinity. Thus one part of the surface line is asymptotic to PQ, the line of uniform motion, another part is asymptotic to a hori- zontal line, while at a certain point at which the depth is the surface of the stream is normal to the bed. FLOW OF WATER IN OPEN CHANNELS. I6 S This is contrary to the fundamental hypothesis that the fluid filaments flow in sensibly parallel lines, and the general equation no longer represents the true condition of flow. In cases such as this, there has been an abrupt rise of the surface of the stream, and what is called a " standing wave " has been produced. In a stream of depth H flowing with a uniform velocity tgr depth to h^ which is > U which is > \ / , construct a weir so as to increase the all* Then in one portion of the stream near the weir the depth aU* aU* is > , while further up the stream the depth is < . o o U* Thus at some intermediate point the depth = a , the cor- o dh responding value of -r- being oo , so that at this point a stand- ds ing wave is produced. Now flT = Mi=-Hi\ and since 1 66 HYDRAULICS. and therefore which condition must be fulfilled for a standing wave. Bazin gives the following table of values of/: Nature of Bed. Slope (A = /) below which stand- ing wave is im- possible. In Metres per Metre. Standing Wave Produced. Slope in Metres per Metre (or Feet per Foot). Least Depth in Metres. Very smooth cemented surface .00147 .00186 .00235 00275 {.002 .003 .004 ( -003 4 .004 ( .006 .004 .006 .010 .006 .010 .015 .08 03 .02 .12 .06 03 .36' .16 .08 I. O6 47 .28 Earth A standing wave rarely occurs in channels with earthen beds, as their slope is almost always less than the limit, .00275. The formation of a standing wave was first observed by Bidone in a small masonry canal of rectangular section. The width of the canal = 0^.325 = x " slope f= -j) of the canal - 02 3 " uniform velocity of flow = 1^.69 = U\ " depth corresponding to U = 0^.064 = H. A weir built across the canal increased the depth of the water near the weir to o w .287 = h^ It was found that the " uniform regime " was maintained up to a point within 4^.5 of the weir. At this point the depth suddenly increased from 0^.064 to about o w .i7O, and between the point and the weir the surface of the stream was slightly convex in form (Fig. 98). FLOW OF WATER IN OPEN CHANNELS. i6 7 With the preceding data and taking a = i.i, is therefore > I at a section ab, Fig. 99. At the section cd, = q H_ h .064 ^87 X 1.69 = 0^.377, and = .055 and is therefore < i. au FIG. 99* Thus the expression I -- is negative for a section ao and positive for a section cd t so that z must change sign between these sections, and will then as become infinite. Consider a portion of a stream bounded by two- trans- verse sections ab, cd, in which a standing wave occurs, Fig. 99. Assume that the fluid filaments flow across the sections in sensibly parallel lines. Let the velocities and area at section ab be distinguished by the suffix i, and those at cd\sy the suffix 2. Then Change of momentum in di- ) rection of flow [ == im P ulse in same direction. Hence w and therefore =A 1 y l - A,v v ... (9) the depths below the surface of the centres of gravity of the sections ab, cd, respectively. 1 68 H YDRA ULICS. Now, v l = u l + V r Therefore Also, as already shown, a ,A, U : = 2av? = AM' and, neglecting F, as compared with 3, , **# =-Arf + Thus and hence ufA-i, = ~^- L ( a + 2 ) = aA * u a 4- 2 where a' = ! , and is 1.033 * !! Similarly it may be shown that Thus equation 9 becomes ~(A^ - Ap?) = ^^ - Aj,. . . . (10) Let the section of the canal be a rectangle of depth H l at ab and H t at ^. Then ff H ufr = u,H, ; - = >, ; -y-= ^. FLOW OF WATER IN OPEN CHANNELS. 169 Therefore, by equation 10, which reduces to //, = H^ satisfies the equation and corresponds to a condition of uniform motion. Also a'u? ^ff.ff. + ff, g H l 2 In Bidone's canal, u 1 = 1^.69, H l = 0^.064. Substituting these values in equation II, the value of H^ is found to be o w . 16, which agrees somewhat closely with the actual measure- ments. N.B. The coefficients a and a' have not been very accu- rately determined, but their exact values are not of great importance. They are often taken equal to unity. H YD RA ULICS. EXAMPLES. 1. What fall must be given to a canal 2600 ft. long, 7 ft. wide at the top, 3 ft. wide at the bottom, \\ ft. deep, and conveying 40 cubic ft. of water per second ? /=^ . Ans. i in 135. 2. Determine the fall of a canal 1500 ft. long, of 2 ft. lower, 8 ft. upper breadth, and 4 ft. deep, which is to convey 70 cubic feet of water per second. Ans. i in 1365.4. 3. For a distance of 300 ft. a brook with a mean water perimeter of 40 ft. has a fall of 9.6 in.; the area of the upper transverse profile is 70 sq. ft., that of the lower 60 sq. ft. Find the discharge. Ans. 662.87 cub. ft. per sec. 4. In a horizontal trench 5 ft. broad and 800 ft. long it is desired to carry off 20 cub. ft. discharge and to let it flow in at a depth of 2 ft. ; what must be the depth at the end of the canal ? (/ = .008.) Ans. 1.64 ft. 5. Water flows along an open channel 12 ft. wide and 4 ft. deep, at the rate of 2 ft. per second. What is the fall? A dam 12 ft. by 3 ft. high is formed across the channel; how high will the water rise over the crest of the dam ? Ans. i in 48o,/ being .08 ; .899 ft. 6. A stream is rectangular in section, 12 ft. wide, 4 ft. deep, and falls i in 100. Determine the discharge (i) with an air-perimeter; (2) without air-perimeter. Ans. (i) 645.398 cub. ft. per sec. (2) 665.088 cub. ft. per sec. 7. A canal 20 ft. wide at the bottom and having side slopes of i to i has 8 ft. of water in it; find the hydraulic mean depth. Ans. 5.24 ft. 8. The water in a semicircular channel of 10 ft. 'radius, when full flows with a velocity of 2 ft. per second ; the fall is i in 400. Find the co- efficient of friction. Ans. .2. 9. Calculate the flow per minute across a given section of a rectarw- gular canal 20 ft. deep, 45 ft. wide, the slope of the bed being 22 in. per mile and the coefficient of friction per square foot = .008. Ans. 279,229 cub. ft. 10. Why does the water of the St. Lawrence rise on the formation of the ice ? 11. Find the depth and width of a rectangular stream of 900 sq. ft. sectional area, so that the flow might be a maximum ; also find the flow, f being .008 and the slope 22 in. per mile. Ans. 21.21 ft.; 42.42 ft.; 4885 cub. ft. per second. FLOW OF WATER IN OPEN CHANNELS. \7\ 12. Water flows along a symmetrical channel, 20 ft. wide at top and 8 ft. wide at bottom ; the friction at the sides varies as the square of the velocity, and is i Ib. per square foot for a velocity of 16 ft. per second. Find the proper slope, so that the water may flow at the rate of 2 ft. per second when its depth is 6 ft. Arts, i in 3445. 13. Calculate the flow across the vertical section of a stream 4 ft. deep, 1 8 ft. wide at top, 6 ft. wide at bottom, the slope of the surface being 18 in. per mile. (/= .008.) Ans. 110.9376 cub. ft. per second. 14. The sewers in Vancouver are square in section and are laid with one diagonal vertical. To what height should the water rise so that (a) the velocity of flow may be a maximum ; (b) the discharge may be a maximum ? (A side of the square = 12 in.)' Ans. (a) .292 ft. above horizontal diameter. (b) .5797 ft. " 15. The sides of an open channel of given inclination slope at 45* and the bottom width is 20 ft. Find the depth of water which will make the velocity of flow across a vertical section a maximum. Ans. 6.73 ft. 17. The banks of a channel slope at 45 ; the flow across a transverse section is to be at the rate of 100 cubic feet at a maximum velocity of 5 ft. per second. Determine the dimensions of the transverse profile. Ans. 11.05 ft. wide at bottom ; 2.28 ft. deep. 1 8. What dimensions must be given to the transverse profile of a canal whose banks slope at 40, and which has to conduct away 75 cubic feet with a mean velocity of 3 ft. per second ? Ans. Depth = 3.6 ft. ; width at bottom = 2.62 ft. 19. The section of a canal is a regular trapezoid ; its slope is i in 500 ; its width at the bottom is 8 ft.; the sides are inclined at 30 to the vertical. On one occasion when the water was 4 ft. deep a wind was blowing up the canal, causing an air-resistance for each unit of free sur- face equal to one fifth of that for like units at the bottom and sides, where the coefficient of friction may be taken to be .08. Determine the discharge. How will the discharge be affected when the canal is frozen over? Ans. 75.34 cub. ft. per sec. 20. The section of a channel is a rhombus with diagonal vertical. How high must the water rise in the channel (a) to give a maximum of flow, and (b) to give a maximum discharge? Ans. If D is the length of the horizontal diameter, and if & is the inclination of a side to the vertical, the water must rise above the horizontal diameter to the height Z)cot0 x .207 in (a) and to the height Z>cotfl x .4099 in (b). 21. In the transverse section ABCD of an open channel with a verti- cal slope of i in 300, the bottom width is 20 ft., the angle ABC 90* 1 72 H YDRA ULICS. and the angle BCD = 45. Find the height to which the water will rise so that the velocity of flow may be a maximum ; also find the dis- charge across the section,/ being .008. Ans. 11.715 ft.; 1584 cub. ft. per second. 22. A canal is 20 ft. wide at the bottom, its side slopes are i| to i, its longitudinal slope is i in 360; calculate H.M.D. and the flow per minute across any given vertical section when there is a depth of 8 ft. of water in the canal. (Coeff. of friction = .008.) Ans. 5.24 ft.; 2762.7776 cub. ft. per second. 23. If a weir 2 ft. high were built across the canal in the preceding question, what would be the increase in the depth of the water? Ans. 2.79 ft. 24. For a small tachometer the velocities are .163, .205, .298, .366, ,61 metre; the number of revolutions per second are .6, .835, 1.467, 1.805, 3.142. Find the constants corresponding to the wheel. Ans. ,162; .202; .309; .367; .595. 25. If the head of water in a channel increase by one tenth, show that the velocity and discharge, respectively, increase by -$ and ^. approximately. If the depth diminish by 8$, show that the velocity and discharge, respectively, diminish by 4% and 12%, approximately. 26. Assuming (i) that a river flows over a bed of uniform resistance to source ; (2) that to maintain stability the velocity is constant from source to mouth ; (3) that the river sections at all points are similar ; (4) that the discharge increases uniformly in consequence of the supply from affluents determine the longitudinal section of such a river. Ans. A parabola. CHAPTER V. METHODS OF GAUGING. I. Gauging of Streams and Watercourses. The amount of flow Q in cubic feet per second across a transverse section of A sq. ft. in area is given by the expression Q - Au, u being the mean velocity of flow in the section in feet per second. Various methods are employed for the determination of u. METHOD I. The most convenient method for gauging small streams, canals, etc., is by means of a temporarily constructed weir, which usually takes the form of a rectangular notch. The greatest care should be exercised to ensure that the crest of the weir is truly level and properly formed and that the sides are truly vertical. The difference of level between the crest of the weir and the surface of the water at a point where it has not begun to slope down towards the weir is best es- timated by means of Boyden's hook gauge, Fig. 100. This gauge consists of a carefully grad- uated rod, or of a rod with a scale attached, having at the lower end a hook with a thin flat body and a fine point. The rod slides in vertical supports, and a slow vertical movement is given by means of a screw of fine pitch. In an experiment, the hook FIG. TOO. point is set truly level with the crest of the weir, and a read- ing is taken. The gauge is then moved away from the weir, HYDRA ULICS. about 2 to 4 ft. for small weirs and about 6 to 8 ft. for large weirs. The hook is then slowly raised, until a capillary eleva- tion of the surface is produced over the point. The hook is now lowered until this elevation is barely perceptible, and a second reading is taken. The difference between the two readings is the difference of level required. In ordinary light, differences of level as small as the one- thousandth of a foot, can be easily detected by the hook gauge, while with a favourable light it is said that an experi- enced observer can detect a difference of two ten-thousandths of a foot. METHOD II. A portion of the stream which is tolerably straight and of approximately uniform section is defined by two transverse lines O^B, OfD, at any distance 5 ft. apart. FIG. 101. The base-line O,O^ is parallel to the thread EF of the stream, and observers with chronometers and theodolites (or sextants) are stationed at (9, , <9 2 . The time T and path EF taken by a float between AB and CD can now be determined. At the moment the float leaves A B the observer at O l signals the observer at (9 2 , who measures the angle O^O^E, and each marks the time. On reaching CD the observer at O. t signals the observer at O l , who measures the angle O^Of, and each again marks the time. Experience alone can guide the observer in fixing the dis- METHODS OF GAUGING. 175 tance 5 between the points of observation. It should be remembered that although the errors of time observations are diminished by increasing S, the errors due to a deviation from lines parallel to the thread of the stream are increased. A number of floats may be sent along the same path, and their velocities UsJ are often found to vary as much as 20 per cent and even more. Having thus found the velocities along any required num- ber of paths in the width of the stream, the mean velocity for the whole width can be at once determined. Surface-floats are small pieces of wood, cork, or balls of wax, hollow metal and wood, colored so as to be clearly seen, and ballasted so as to float nearly flush with the water-surface and to be little affected by the wind. Subsurface-floats. A subsurface-float consists of a heavy float with a comparatively large intercepting area, maintained at any required depth by means of a very fine and nearly vertical cord attached to a suitable surface-float of minimum immersion and resistance. Fig. 102 shows such a combina- tion, the lower float consisting of two pieces of galvanized iron soldered together at right angles, the upper float being merely a wooden ball. FIG. 102. FIG. 103. Another combination of a hollow metal ball with a piece of cork is shown by Fig. 103. The motion of the combination is sensibly the same as that HYDRA ULICS. of the submerged float, and gives the velocity at the depth to which the heavy float is submerged. Twin-floats. Two equal and similar floats (Fig. 104), one denser and the other less dense than water, 1 are connected by a fine cord. The velocity (v t ) of the combination is approximately the mean of the surface-velocity (v s ) and of the velocity (v^) at the depth to which the heavier float is submerged. Thus FIG. 104. and therefore d> ~~"~ ^ t */s 9 so that v d can be determined as soon as the value of v t has been observed and the value of v s found by surface-floats. Velocity-rod. This is a hollow cylindrical rod of ad- justable leiigth and ballasted so as to float nearly vertical. It sinks almost to the bed of the stream, and its velocity (v m ) is approximately the mean velocity for the whole depth. Francis gives the following empirical formula connecting the mean velocity (v m ) with the observed velocity (v r ) of the rod : ...*/), =zv(i.oi2 d being depth of stream, and d' the depth FlG - I0 5- of water below bottom of rod ; but d' should not exceed about one fourth of d. METHOD III. Pitot Tube and Darcy Gauge. A Pitot tube (Figs. 106 to 108) in its simplest form is a glass tube with a right-angled bend. When the tube is plunged vertically into the stream to any required depth z below the free surface, with its mouth pointing up-stream and normal to the direction of METHODS OF GAUGING. 177 flow, the water rises in the tube to a height h above the out- side surface, and the weight of the column of water z -f- h FIG. i 06. FIG. 107. FIG. 108. high, is balanced by the impact of the stream on the mouth. Hence, (Chap. VI.), wA(z -f- k) = wAz -f- kwA , and therefore A being the sectional area of the tube, u the velocity of flow at the given depth, and k a coefficient to be determined by experiment. A mean value of k is 1.19. With a funnel-mouth or a bell- mouth, Pitot found k to be 1.5. This form of mouth, however, interferes with the stream-lines, and the velocity in front of the mouth is probably a little different from that in the unob- structed stream. The advantages of tubes of small section are that the dis- turbance of the stream-lines is diminished and the oscillations of the column of water are checked. Darcy found by careful measurement that the difference of level between the surfaces of the water-column in a tube of small section placed as in Fig. 106, and of the water-column placed as in Fig. 107 with HYDRA ULICS. FIG. 109. its mouth parallel to the direction of flow, is almost exactly equal to -. When the tube is placed as in Fig. 108 with its mouth pointing down- stream and normal to the direction of flow, the level of the surface of the water in the tube is at a depth ti below the outside surface, and where k f is a coefficient to be determined by experi- ment and a little less than unity. In this case the tube again obstructs the stream- lines. Pitot's tube does not give measurable indi- cations of very low veloc- ities. A serious objection to the simple Pitot tube is the difficulty of obtaining accurate readings near the surface of the stream. This objection is removed in the case of Darcy's gauge, shown in the accompany- ing sketch, Fig. 109. A and B are the water- inlets; C and D are two double tubes ; E is a brass METHODS OF GAUGING. tube containing two glass pipes which communicate at the bottom with the water-inlets and at the top with each other, and with a pump F by which the air can be drawn out of the glass, pipes thus allowing the water to rise in them to any convenient height. Thus Darcy's gauge really consists of two Pitot tubes con- nected by a bent tube at the top and having their mouths at right angles or pointing in opposite directions. If h is the difference of level between the water-surfaces in the tubes when the mouths are at right angles, then and Darcy's experiments showed that k does not sensibly differ from unity. When the mouths point in opposite directions, let h^ h^ be the differences of level between the stream-surface and the surfaces of the water in the tube pointing up-stream and the tube pointing downstream, respectively. Then u* ** = k{ 2g'> U* and therefore u* h j. h - (k , + k ) *>2g where k = k v -\- k^ k having been determined experimentally once for all, the difference of level (= h^ -\- h^) between the columns for any given case can be measured on the gauge and then u can be at once found. 1 80 H YDRA ULICS. A cock may be inserted in the bend connecting the two tubes, and through this cock air may be exhausted and a partial vacuum created in the upper portion of the gauge. The water-columns will thus rise to higher levels, but the dif- ference between them will remain constant. Thus the surface of the column in the down-stream tube may be brought above the level of the outside surface, and the reading is then easily made. Sometimes the gauge is furnished with cocks at the lower parts of the tubes, and if these cocks are closed when the measurement is to be made, the gauge may be removed from the stream for the readings to be taken. METHOD IV. Current-meters. The velocity of flow in large streams and rivers is most conveniently and most ac- curately ascertained by means of the current-meter. The earliest form of meter, the Woltmann mill, is merely a water- mill with flat vanes, similar in theory and action to the .wind- mill. When the Woltmann is plunged into a current, a counter registers the number of revolutions made in a given interval of time, and the corresponding velocity can then be deter- mined. This form of meter has gone out of use and has been replaced by a variety of meters of greater accuracy, of finer construction, and much better suited to the work. In its sim- plest form the present meter consists of a screw-propeller wheel (Fig. 1 10), or a wheel with three or more vanes mounted on a spindle and connected by a screw-gearing with a counter which registers the number of revolutions. The meter is put ' in or out of gear by means of a string or wire. When a cur- rent velocity at any given point is to be found, the reading of the counter is noted, the meter is sunk to the required position, and is then set and kept in gear for any specified interval of time. At the end of the interval the meter is put out of gear and is raised to the surface when the reading of the counter is again noted. The difference between the readings gives the number of revolutions made during the interval, and the veloc- ity is given by an empirical formula connecting the velocity and the number of revolutions in a unit of time. METHODS OF GAUGING. The vane Fis introduced to compel the meter to take its proper direction. In order to prevent the mechanism of the meter from being FIG. 1 10. FIG. in. injuriously affected by floating particles of detritus, Revy en- closed vthe counter in a brass box, Fig. ill, with a glass face, FIG. 112. FIG. ri3. and filled the box with pure water so as to ensure a constant coefficient of friction for the parts which rub against each other. In the best meters, however, the record of the number 1 82 HYDRAULICS. of revolutions is kept by means of an electric circuit, Fig. 112, which is made and broken once, or more frequently, each revolution, and which actuates the recording apparatus. The time at which an experiment begins and ends is noted, and the revolutions made in the interval are read on the counter, which may be kept in a boat or on the shore, as the circumstances of the case may require. The meter is usually attached to a suit- ably graduated pole, so that the depth of the meter below the water-surface can be directly read. The mean velocity for the whole depth at any point of a stream may be found by moving the meter vertically down and then up, at a uniform rate. The mean of the readings at the two surface positions and at the bottom position will be the number of revolutions corre- sponding to the mean velocity required. The mean velocity for the whole cross-section may also be determined by moving the meter uniformly over all parts of the section. Before the meter can be used it must be rated. This is done by driving the meter at different uniform speeds through still water. Experiment shows that the velocity v and the number of revolutions n are approximately connected by the formula v = an + b, where a and b are coefficients to be determined by the method of least squares or otherwise. Exner gives the formula V Q being the velocity at which the meter just ceases to re- volve. OTHER METHODS. Many other pieces of apparatus for the measurement of current velocities have been designed. PerrodiTs hydrodynamometer, for example, gives the ve- locity directly in terms of the angle through which a vertical torsion-rod is twisted, and in this respect is superior to the current-meter. METHODS OF GAUGING. 183 FIG. 114. The hydrometric pendulum (Fig. 114), again, connects the velocity with the angular devia- tion from the vertical of a heavy ball suspended by a string in the current. Hydrometric and torsion bal- ances have also been devised, but they must be regarded rather as curiosities than as being of any real practical use. 2. Gauging of Pipe Flow. A variety of meters have been designed to register the quantity of water delivered by a pipe. The principal requisites of such a meter are : 1. That it should register with accuracy the quantity of water delivered under different pressures. 2. That it should not appreciably diminish the effective pressure of the water. 3. That it should be compact and adaptable to every situation. 4. That it should be simple and durable. The Venturi Meter (Fig. 115) is so called from Venturi, who first pointed out the relation between the pressures and veloci- ties of flow in converging and diverging tubes. FIG. 115. As shown by the longitudinal section, Fig. 116, this meter consists of two truncated cones joined at the smallest sections by a short throat-piece. At A and B there are air-chambers with holes for the insertion of piezometers, by which the fluid 1 84 HYDRA ULICS. pressure may be measured. By Art. 5, Chap. I, the theoretical quantity Q of flow through the throat at A is a t , # a being the sectional areas at A and B, respectively, and ff t H l the difference of head in the piezometers, or the "head on Venturi," as it is called. FIG. 116. Introducing a coefficient of discharge C, the actual delivery through 'A is An elaborate series of experiments by Herschel gave C values varying between .94 and 1-04, but the great majority of the values lay between .96 and .99. The piezometers may be connected with a recorder, and thus a continuous register of the quantity of water passing through the meter may be obtained at any convenient position within a radius of 1000 ft. This distance may be extended to several miles by means of an electric device. Other meters may be generally classified as Piston or Re- ciprocating Meters and Inferential or Rotary Meters. They are all provided with recorders which register the delivery with a greater or less degree of accuracy. The piston meter (Fig. 1 1 8) is the more accurate and gives a positive measurement of the actual delivery of water as METHODS OF GAUGING. 185 recorded by the strokes of the piston in a cylinder which is filled from each end alternately. Thus an additional advan- l_ ..: PIG. 118. SCHONHEYDER'S POSITIVE METER. FIG. 119. THE UNIVERSAL METER. FIG. 120. THE BUFFALO METER. FIG. 121. THE UNION ROTARY PIS- TON METER. tage possessed by a water-engine is that the working cylinder will also serve as a meter. In inferential meters, a drum or turbine is actuated by the force of the current passing through the pipe, but it often happens that when the flow is small the force is insufficient to cause the turbine to revolve, and there is consequently no register of the corresponding quantity of water passing through the meter. CHAPTER VI. IMPACT. Note. The following symbols are used : z/, = the velocity of the jet before impact ; z> 2 = " " " " " after leaving the vane ; u " " " " vane ; V " " " water relatively to the vane ; A = sectional area of the impinging jet ; m = mass of the water reaching the vane per second. i. Impact of a Jet upon a Flat Vane oblique to the Direction of the Jet. Let 6 be the angle between the nor- mal to the vane and the direction of the impinging jet,

. . . ( 23 ) The maximum efficiency= - (24) 2. Reaction Jet Propeller. The term reaction is em- ployed to denote the pressure upon a surface due to the di- rection and velocity with which the water leaves the surface. Water, for example, issues under the head h and with the IMPACT. IQI velocity v (at contracted section) from an orifice of sectional area A in the vertical side of a vessel, Fig. 124. Let R be the reaction on the opposite vertical side of the vessel, and let Q be the quantity of water which flows through the orifice per second. Then R = horizontal change of momentum wQ w = v CcAv* 2wc c c v Ah 2wAh, . . . (i) o e disregarding the contraction and putting c v I. Thus the reaction is double the corresponding pressure when the orifice is closed (Special Case I, Art. i). Again, let the vessel be propelled in the opposite direction with a velocity u relatively to the earth. Then v l u is the velocity of the jet at the contracted section relatively to the earth and R = horizontal change of momentum = ^Q( Vl -u) . . (2) o The useful work done by the jet IV = Ru = Qu(v l -u) (3) o The energy carried away by the issuing water Hence w w (v. uY the total energy = Qu(v, -u) + Q (5) IQ2 HYDRAULICS. , and w g 2U the efficiency = 5 r = . . . . . (6) w v, u v, -4- u Thus the more nearly v l is equal to u, and therefore the larger the area A of the orifice, the greater is the efficiency. If the vessel is driven in the same direction as the jet, then 77, -f- u is the relative velocity of the jet with respect to the earth, and the reaction is R horizontal change of momentum -G& + u ) = c^Av^v, + u) (7) disregarding the contraction and putting c, = I. 3. A Jet of Water impinging upon a Surface of Rev- olution moving in the Direction of its Axis and also in the Line of the Jet's Motion. Disregarding friction, the water flows over the surface without any change in the magni- tude of the relative velocity v t u, but the stream-lines are deviated from their original direction through an angle /?. (N.B. The sign before u is plus if the surface and jet are moving in opposite directions.) Let the water leave the surface at D, and in the direction of the tangent at D take DE to represent v l u in direction and magnitude. Also draw DF parallel to the axis of the sur- face and take DF to represent //, Complete the parallelogram EF. The diagonal DG evidently represents in direction and magnitude the absolute velocity v^ with which the water leaves IMPACT. 193 the surface. Hence, from the triangle DFG, since the angle DFG = n ft, v* = (v l uj + u* - 2(z/ 1 *) cos (?r /?), from which /? z/j 3 v* = 2^(^j w)(i cos /?) = 4(z/ u) sin 2 . (i) Also, -A(v l u) = the quantity of water reaching the . Au(v l u) sm sm^. . (4) W -V v* 2 A 194 HYDRAULICS. This is a maximum when , = 3, > - " ... (5) and therefore the maximum efficiency = sin 2 -. . . . (6) If, instead of one surface, a series of surfaces are succes- sively introduced at short intervals at the same point in the path of the jet, the quantity of water reaching each surface per second becomes w m= " (7) and hence the useful work, pressure, and efficiency also respec- tively become w ft 2~A^u(^-u)sm 9 ~', (8) Avfa w)sin a ; (9) u(v l u} . 2 ,/? 4 " V* 2 The efficiency is a maximum when v. (ii) Q its value then being sin a . 2 It will be observed that the results given by equations 2 to II are identical with those given by equations 17 to 20 and 21 to 24, Art. I, except that in each case there is an additional ft factor 2 sin 8 or I cos ft. This factor is greater than unity, and therefore the pressure, useful work, and efficiency are each IMPACT. 195 increased, if ft > 90, i.e., in the case of a concave vane ; while in the case of a convex vane, ft being < 90, the factor is also less than unity and they are each diminished. SPECIAL CASE. Let fi = 180, i.e., let the vane be of the cup type and in the form of a hemisphere. 1 80 The maximum efficiency is sin" = unity, and is per- fect. The water should therefore leave the surface without velocity; and this is the case ; for, by equation I, Hence v* v* = ^ii(v^ u), and u = . 2 v* v* = v*, and therefore ^ a = o. 4. Impact of a Jet of Water upon a Vane with Borders. Let the vane in Art. i be provided with borders, Figs. 126, 127, so as to produce a further deviation of the stream- lines, and let the water finally flow off with a velocity v* in a direction making an angle 0' with the normal to the vane. FIG. 126. FIG. 127. Then the normal pressure = N = mv t cos T mv^ cos tf 3= mu cos = m(v l cos ^F z> a cos 0' =F u cos 0), the sign of the second term being plus or minus according to the direction in which the stream-lines are finally deviated. 196 HYDRAULICS. The effect of the borders is therefore to increase or diminish the normal pressure, and hence also the useful work and the efficiency. SPECIAL CASE. Let the vane be at rest, i.e., let u = o, and let the final and initial directions of the jet be parallel. Also, let v 1 = Vf Then N = m(v^ cos 6 -\- v l cos 6) w = 2Av? cos 6 o = 4wAH cos 6. Hence, if fl o, the normal pressure N= qwAH = four times the weight of a column of water of height H and sec- tional area A. 5. Pressure of a Steady Stream in a Uniform Pipe against a Thin Plate AB Normal to the Direction of Motion. The stream-lines in front of the plate are deviated and a contraction is formed at Cf^ They then converge, leaving a mass of eddies behind the plate. Consider the mass bounded by the transverse planes C l C l> 3 , where the stream-lines are again parallel. At Ci let A , A l , v l , z l be the mean intensity of the press- ure, the sectional area of the waterway, the velocity of flow, and the elevation of the C. of G. of the section above datum. Let / 2 , AS, z> 3 , z^ be corre- sponding symbols at Cf v Let / 3> A 19 v lt * be corre- sponding symbols at C 9 C 3 . Let a be the area of the plate. Let c c be the coefficient of contraction. Neglect the skin and fluid friction between C l C l and Then by Bernoulli's theorem, + + ' ' ' ' W 2g W 2g W 2g 2g IMPACT. 197 ( v v\ the term - representing the loss of head due to the bending of the stream-lines between Cf^ and C 3 C 3 . Hence A -A (v* - v>Y Again, let R be the total pressure on the plane. Then x . x A ( fluid pressure in the direction A -M, = (A - AK = | of the axis _ 2 * = component of the weight in the direction of the axis. Thus ^ __ j> s )A l + wA l (z l ^,) R = change of motion in direction of axis = 0, since the motion is steady. Hence R A -A (",-".)' wA l w 2g But A^, = AM = c c (A l a)Vr Therefore =-*${&&>- >} v? ( m } a = wa m \ , r I [ , 2g \ c c (m - i) j A where m = , or a R = r m ) where K in \ , r I > \ c c (m - i) f 198 HYDRAULICS'. 6. Pressure of a Steady Stream in a Uniform Pipe on a Cylindrical Body about Three Diameters in Length. The stream-lines in front of the body are deviated and a contraction is formed at C 9 C t . They then converge, flow in parallel lines, and converge a second time at C 3 C 3 , leaving a mass of eddies behind the body. Consider the mass bounded by the planes C^ C t C 4 . As in the previous article, let />,, A lt v l , z l be the intensity of pressure, sectional area of the waterway, velocity of flow, and elevation of C. of G. above datum at r A, ^ 2 , z> 2 , ? . {^s - *$ , ! * 4++ " t " ^"" being the loss of head between <7 a , and C,C 3 and being the loss of head between C 3 C 9 and C t C t . o Hence * i A A _ (^. ^) a I (^ - O a J " 4-f ~^^ "IF IF" But A& = ^ 3 e; a = ^ 3 ^,, and A 3 = A, a. IMP A CT. 199 Therefore , y t A_ _A^n a I \7JJ^-a) A,- a] J where m = *. Also, as in the preceding article, (A- Hence f 2g (m i) 2 (m - i) a V, where m = -, and a This value of K is always less than the value of K for the plate in the preceding article for the same values of m, a, and c f Hence the pressure on the cylinder is also less than the corresponding pressure on the plate. In every case K should be determined by experiment. 7. Jet impinging upon a Curved Vane and deviated wholly in one Direction Best Form of Vane. Let the jet, of sectional area A, moving in the direction AB with a velocity v^ , drive the vane AD in the direction AC with a velocity u. 200 HYDRAULICS. Take AB to represent v^ in direction and magnitude. " AC " " u " " ". Join CB. Then CB evidently represents F, the velocity of the water relatively to the vane, in direction and magnitude. If CB is parallel to the tangent to the vane at A, there will be no sud- FIG. 130. den change in the direction of the water as it strikes the vane, and, disregarding friction, the water will flow along the vane from A to D without any change in the magnitude of the rela- tive velocity V (= CB). The vane is then said to "receive the water without shock."* Again, from the triangle ABC, denoting the angles BA C, ABC, ACB, byA,, C, respectively. sin B u _ AC _ sin B __ ^ = " ~AB ~ sin C ~ sin (A + B)' ' ' . . (I) and therefore cot B = cosec A cot A, .... (2) IMPACT. 201 a formula giving the angle between the lip and the direction of the impinging jet, which will ensure the water being received " without shock." In the direction of the tangent to the vane at D, take DE = CB (= V). Draw DF parallel and equal to AC(= u). Complete the parallelogram EF. Then the diagonal DG evidently represents in direction and magnitude the absolute velocity v^ with which the water leaves the vane. Draw AK equal and parallel to DG (= z/ a ). Join BK. Then BK represents the total change of velocity between A and D in direction and magnitude. Thus, if R is the resultant pressure on the vane, then R = m. BK. Let ML be the projection of BK upon AC. Then ML represents the total change of velocity in the direction of the vane's motion. Let P be the pressure upon the vane in this direction. Then P=m. LM. (3) The useful work = Pu = mu . LM = m V * ~ V * . . . (4) w A v? The total available work = - A -- (5) ,, ~ . mu. LM v? v* The efficiency -- = img -- r- ...... (6) w A v? * wAv? Again, join CK.- Then, since A C is equal and parallel to DF, and AK to DG, the line CK is equal and parallel to DE, and is therefore equal to CB. Thus in the isosceles triangle CBK, CB is equal and parallel to the relative velocity Fat A, CK is equal and parallel to the 2O2 HYDRA ULICS. relative velocity Fat D, and the base B K represents the total change of motion. Let 8 be the angle through which the direction of the water is deviated, i.e., the angle between AB and AK. Then = V* -\- U* 2V Ji COS (A + #), ...... (7) and also F 3 = CK* = CB* = AB* + AC* - 2AB . AC cos A = v* -\-u* 2v ji cos A .......... (8) Hence L = u \ v t cos (A + 6) v l cos A } . . . (9) If BH is drawn parallel to the tangent at D, BK evidently bisects the angle between BC and BH, and this angle is equal to the angle between the tangents to the vane at A and D. Let a be the sttpplcmcnt-^f the angle between the normals at A and D. Then the angle KCB a, and the angle CBK = -(180 - ) = 90 - 2 2 Therefore BK = 2CB (cos 00 - ] = 2Fsin -. \ 21 2 Hence ;in- (10) IMP A CT. 2O3 Let X, Fbe the components of R in the direction of the normal at A and at right angles to this direction. Then Y=R cos- = mVsm or; .... (n) X = R sin = 2m V sin 3 - = m V( i cos a). ( 1 2} 2 2 The efficiency is a maximum when dP The efficiency is nil when Pu = o, i.e., when u = o or P = o. . . . (14) In the latter case, since P m. LM, the projection LM must be nil, and therefore BK must be at right angles to A C, as in Fig. 131. FIG. 131. FIG. 132. 204 HYDRA ULICS. The angle ACB is now = 180 -- , and therefore u_ sin ABC v l ~~ sin A CB sn in (180 -^ (IS) sm 2 If BK is parallel to AC (Fig. 132), then the angle ACB = -(180 -) + = 90 + - 2 2 .and therefore sin (90 + - + A\ cost- ~4- A] u_ _ sin ABC V r 2 ) _ \2 1 sm I Qcr + - 1 cos - SPECIAL CASE. Let the direction of the impinging jet be tangential to the vane at A, and let the jet and vane move in the same direction. Then V v. u y m = A(v. 11) ; g P = Y= A(v t u)\i cos a) = 2 A(v^ u) sin 2 -; 5 o W & useful work = Pu = 2 Ati(v, uY sin 2 ; g 2 U(V. U}" OL efficiency = 4 sin . IMP A CT. 20$ This is a maximum and equal to sin 2 when v l = $u. 27 2 These results are identical with those for a concave cup when a = 180. Instead of one vane let a series of vanes be successively introduced at short intervals at the same point in the path of the jet. Then w m = Av^ and hence the pressure P, useful work, and efficiency respec- tively become A o w A Av, . S and 8. Friction. The effect of friction has been disregarded, and nothing definite is known as to its action or law of distri- bution. It has been suggested to assume that the loss of head due to friction is a fraction of the head due to the velocity of the jet relatively to the surface over which it spreads. Thus in Art. 7 V* the loss of head due to friction =/ V* and the corresponding loss of energy = wQ*f 9. Resistance to the Motion of Solids in a Fluid Mass. The preceding results indicate that the pressure due to 2O6 HYDRA ULICS. the impact of a jet upon a surface may be expressed in the form A being the sectional area of the jet, V the velocity of the jet relatively to the surface, and K a coefficient depending on the position and form of the surface. Again, the normal pressure (N) on each side of a thin plate, completely submerged in an indefinitely large mass of still water, is the same. If the plate is made to move hori- zontally with a velocity F, a forward momentum is developed in the water immediately in front of the plate, while the plate tends to leave behind the water at the back. A portion of the water carried on by the plate escapes laterally at the edges and is absorbed in the neighboring mass, while the region it originally occupied is filled up with other particles of water. Thus the normal pressure N, in front of the plate, is increased by an amount n, while at the back eddies and vortices are pro- duced, and the normal pressure N at the back is diminished by an amount n' . The total resultant normal pressure, or the normal resistance to motion, is n-\- n', and this increases with the speed. In fact, as the speed increases, n' approximates more and more closely to N, and in the limit the pressure at the back would be nil, so that a vacuum might be main- tained. Confining the attention to a plate moving in a direction normal to its surface, the resistance is of the same character as if the plate is imagined to be at rest and the fluid moving in the opposite direction with a velocity V. So, if both the water and the plate are in motion, imagine that a velocity equal and opposite to that of the water is impressed upon every particle of the plate and of the water. The resistance is then of the same character as that of a plate rrioving in still water, the velocity of the plate being the velocity relatively to the water. Thus, in general, the resistance to the motion of such a plane moving in the direction of the normal to its IMPACT. 207 surface, with a velocity V relatively to the water, may be ex- pressed in the form R - KwA - , A being the area of the plate, and K a coefficient depending upon the form of the plate and also upon the relative sectional areas of the plate and of the water in which it is submerged. According to the experiments of Dubuat, Morin, Piobert, Didion, Mariotte, and Thibault, the value of K may be taken at 1.3 for a plate moving in still water, and at 1.8 for a current moving on a fixed plate. Unwin points out the unlikelihood of such a difference between the two values, and suggests that it might possibly be due to errors of measurement. Again, reasoning from analogy, the resistance to the motion of a solid body in a mass of water, whether the body is wholly or only partially immersed, has been expressed by the formula R = KwA, V being the relative velocity of the body and water, A the greatest sectional area of the immersed portion of the body at right angles to the direction -of motion, and K a coefficient de- pending upon the form of the body, its position, the relative sectional areas of the body and of the mass of water in which it is immersed, and also upon the surface wave-motion. The following values have been given for K\ K = i.i for a prism with plane ends and a length from 3 to 6 times the least transverse dimension ; K = i.o for a prism, plane .in front, but tapering towards the stern, the curvature of the surface changing gradu^ ally so that the stream-lines can flow past without any production of eddy motion, etc.; 208 HYDRA ULICS. K .5 for a prism with tapering stern and a cut-water or semi-circular prow ; K = .33 for a prism with a tapering stern and a prow with a plane front inclined at 30 to the horizon ; K = .16 for a well-formed ship. Froude's experiments, however, show that the resistance to the motion of a ship, or of a body tapering in front and in the rear, so that there is no abrupt change of curvature lead- ing to the production of an eddy motion, is almost entirely due to skin-friction (see Art. i, Chap. II). IMPACT. 209 EXAMPLES, 1. A stream with a transverse section of 24 square inches delivers y 10 cubic feet of water per second against a flat vane in a normal direc- ^ tion. Find the pressure on the vane. Am. 1171! Ibs. 2. If the vane in question i moves in the same direction as the im- ./ pinging jet with a velocity of 24 ft. per second, find (a) the pressure on the vane ; (b) the useful work done ; (c) the efficiency. Am. (a) 4211 Ibs.; (ff) 10,125 ft.-lbs.; (c) .288. 3. What must be the speed of the vane in question 2, so that the J efficiency of the arrangement may be a maximum ? Find the maximum ^ efficiency. Ans. 20 ft. per sec.; ^V % 4. Find (a) the pressure, (b) the useful work done, (c) the efficiency, when, instead of the single vane in question 2, a series of vanes are intro- duced at the same point in the path of the jet at short intervals. Ans. (a) 703^ Ibs.; (b} 16,875 ft.-lbs.; (c) .48. What must be the speed of the vane to give a maximum efficiency ? What will be the maximum efficiency? Ans. 30 ft. per sec.; .5. 5. A stream of water delivers 7,500 gallons per minute at a velocity of 15 ft. per second and strikes an indefinite plane. Find the normal pres- sure on the vane when the stream strikes the vane (a) normally; (d) at an angle of 60 to the normal. Ans. (a) 585.9 Ibs.; 292.9 Ibs. 6. A railway truck, full of water, moving at the rate of 10 miles an hour, is retarded by a jet flowing freely from an orifice 2 in. square in the front, 2 ft. below the surface. Find the retarding force. Ans. 7.97 Ibs. 7. A jet of water of 48 sq. in. sectional area delivers 100 gallons per Q% second against an indefinite plane inclined at 30 to the direction of the- ( jet ; find the total pressure on the plane, neglecting friction. How will the result be affected by friction ? Ans. 750 Ibs. ' 8. If the plane in question 7 move at the rate of 24 ft. per second in a direction inclined at 60 to the normal to the plane, find the useful work done and the efficiency. Ans. 2250 ft.-lbs.; T V At what angle should the jet strike the plane so that the efficiency might be a maximum? Find the maximum efficiency. Ans. sin 1 ; -.. 9. A stream of 32 square inches sectional area delivers 32 cub. feet of water per second. At short intervals a series of flat vanes are intro- 210 HYDRA ULICS. duced at the same point in the path of the stream. At the instant of impact the direction of the jet is at right angles to the vane, and the vane itself moves in a direction inclined at 45 to the normal to the vane. Find the speed of the vane which will make the efficiency a maximum. Also find the maximum efficiency and the useful work done. Ans. 15.08 ft. per sec.; / T ; 2io6f|-f ft.-lbs. 10. In a railway truck, full of water, an opening 2 in. in diameter is made in one of the ends of the truck, 9 ft. below the surface of the water. Find the reaction (a) when the truck is standing; (b) when the truck is moving at the rate of 10 ft. per second in the same direction as the jet ; (c) when the truck is moving at the rate of 10 ft. per second in a direction opposite to that of the jet. If this movement of the truck is produced by the reaction of the jet, find the efficiency. Ans. (a) 24.55 Ibs. per sq. in.; (b) 34.78 Ibs. per sq. in.; (c) 14.3 Ibs. per sq. in.; .588. 11. From a ship moving forward at 6 miles an hour a jet of water is sent astern with a velocity relative to the ship of 30 feet per second from a nozzle having an area of 16 square inches; find the propelling force and the efficiency of the jet as a propeller without reference to the man- ner in which the supply of water may be obtained. Ans. i 12. A stream of 64 sq. in. section strikes with a 40- ft. velocity against a fixed cone having an angle of convergence = 100 ; find the hydraulic pressure. Ans. 492.1 Ibs. 13. A jet of 9 sq. in. sectional area, moving at the rate of 48 ft. per second, impinges upon the convex surface of a paraboloid in the direc- tion of the axis and drives it in the same direction at the rate of 16 ft. per second. Find the force in the direction of motion, the useful work done, and the efficiency. The base of the paraboloid is 2 ft. in diameter and its length is 8 inches. Ans. 25 Ibs.; 400 ft.-lbs.; r y. 14. A stream of water of 16 sq. in. sectional area delivers 12 cubic feet of water per second against a vane in the form of a surface of revolu- tion, and drives in the same direction, which is that of the axis of the vane. The water is turned through an angle of 120 from its original direction before it leaves the vane. Neglecting friction, find the speed of vane which will give a maximum effect. Also find impulse on vane, the work on vane, and the velocity with which the water leaves the vane. Ans. 36 ft. per sec.; 562^ Ibs.; 20,250 ft.-lbs.; 95.24 ft. per sec. 15. At 8 knots an hour the resistance of the Water-witch was 5500 Ibs.; the two orifices of her jet propeller were each 18 in. by 24 in. Find (a) the velocity of efflux; (b) the delivery of the centrifugal pump; IMPACT. 211 (V) the useful work done ; (d) the efficiency; ( 24. Two cubic feet of water are discharged per second under a press- ure of loo Ibs. per sq. in. through a thin-lipped orifice in the vertical side of a vessel, and strike against a vertical plate. Find the pressure on the plate and the reaction on the vessel. Ans. 475.82 Ibs. 25. A stream moving with a velocity of 16 ft. per second in the direc- tion ABC, strikes obliquely against a flat vane and drives it with a velocity of 8 ft. per second in the direction BD, the angle CBD being 30. Find {a) the angle between ABC and the normal to the plane for which the efficiency is a maximum ; (b) the maximum efficiency ; (c) the velocity with which the water leaves the vane; (d} the useful work per cubic foot of water. Ans. (a) 21 44'; (b) .25664; (c) 12.6 ft. per sec.; (d) 256.64 ft.-lbs. CHAPTER VII. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. I. Hydraulic Motors are machines designed to utilize the energy possessed by a moving mass of water in virtue of its position, pressure, and velocity. The motors may be classified as follows : (1) Bucket Engines. In this now antiquated form of motor weights are raised and resistances overcome by allowing water to flow into suspended buckets, thus causing them to descend vertically. (2) Rams and Jet-pumps, in which the impulsive effect of one mass of water is utilized to drive a second mass of water. (3) Water-pressure Engines are especially adapted for high pressures and low speeds, and necessarily have very heavy moving parts. With low pressures the engine becomes un- wieldy and costly. Pressure-engines are either reciprocal or rotative. The latter are very convenient with moderately high pressures and -especially when they are to drive machinery which is to be used intermittently. They also give an exact measurement of the water used. Direct-acting pressure-engines are of great advantage where a slow and steady motion is required, as, for example, in work- ing cranes, lifts, etc. (4) Vertical Wat er-iv heels, in which the water acts almost wholly by weight, or partly by weight and partly by impulse, or wholly by impulse. (5) Turbines, in which the water acts wholly by pressure or wholly by impulse. 214 HYDRAULICS. 2. Hydraulic Rams. By means of the hydraulic ram a quantity of water falling through a vertical distance h l is made to force a smaller weight of water to a higher level. The water is brought from a reservoir through a supply- pipe 5. At the end B of this pipe there is a check- or clack- valve opening into an air-chamber A, which is connected with a discharge-pipe D. At C there is a weighted check- or clack- valve opening inwards, and the length of its stem (or the stroke) is regulated by means of a nut or cottar at E. When the waste- valve at C is open the water begins to escape with a velocity due to the head h l and suddenly closes the valve. The momentum. FIG. 133. of the water in the pipe opens the valve at B, and a portion of the water is discharged into the air-vessel. From this vessel it passes into the discharge-pipe in consequence of the reaction of the compressed air. At the end of a very short interval of time the momentum of the water has been destroyed, the valve at B closes, the waste-valve again opens, and the action com- mences as before. It is found that the efficiency of the ram is increased by introducing a small air-vessel at F, supplied with a check- or clack-valve opening inwards at G. The wave-motion started up in the supply-pipe by the opening and closing of the valve at B has been utilized in driving a piston so as to pump up water from some independent source. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 21$ Let v be the velocity of flow in the supply-pipe at the mo- ment when the valve at C is closed. " W l be the weight of the mass of water in motion. W v 2 Then - - is the energy of the mass, and this energy is expended in opening the valve at B, forcing the water into the air-chamber, compressing the air, and finally causing the eleva- tion of a weight W^ of the water through a vertical distance k '. Let h f be the head consumed in frictional and other hy- draulic resistances. Then W,(h' + h f } = the actual work done = ' -. This equation shows that, however great h' may be, W^ has a definite and positive value, and therefore water may be raised to any required height by the hydraulic ram. WJt' The efficiency of the machine = 2 , and may be as much 11 as 66 per cent if the machine is well made. 3. Pressure-engines. The energy required to drive a press- ure-engine is usually supplied by means of steam-pumps, but an accumulator is often interposed between the pumps and the motor in order to store up the pressure energy of the water. Indeed, it is perhaps to the introduction of the accumulator that the success of hydraulic transmission is especially due. Its cost, however, only allows of its use in cases where the demand for energy is for short intervals of time. In its simplest form the accumulator is merely a vertical cylinder into which the water is pumped and from which it is then discharged by the descent of a heavily loaded piston. The water-pressure thus developed in ordinary hydraulic ma- chinery is from 700 to 800 Ibs. per square inch, but in riveting and other similar machinery pressures of 1500 Ibs. per square inch and upwards are often employed. Fig. 134 represents an accumulator designed by Tweddell for these higher pressures. 216 HYDRAULICS. The loaded cylinder A slides upon a fixed spindle B. The water enters near the base, passes up the hollow spindle, and fills the annular space surrounding the spindle. Thus FIG. 134- the whole of the weight is lifted by the pressure of the water upon a shoulder C. The water section being small, any large demand for water will cause the loaded cylinder to fall rapidly, so that when it is brought to rest there will be a considerable HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 2 1/ increase of pressure which is of advantage in punching, rivet- ing, etc. Let Wbe the weight of the loaded cylinder. Let /'"be the friction of each of the two cup-leathers. Let T-J be the radius of the cylinder, r t the radius of the spindle. Let h be the height of the column of water above the pipe D. Let w be the specific weight of the water. Then/j, the intensity of the pressure in D when the cylinder is rising, W+2F = Wk -f- ( a __ 5r t and /, , the intensity of the pressure in D when the cylinder is falling, W-2F Hence an approximate measure of the variation of the pressure is p l p^ , ^ r . , which ordinarily varies from about ifo of the pressure for a i6-in. ram to 4$ for a 4-in. ram. In a direct-acting pressure-engine let A be the sectional area of the working cylinder (Fig, 135). Let a be the sectional area of the supply- pipe. Let A = na. Let IV be the weight of the water, piston, FJG - '35- and other reciprocating parts in the working cylii.der. Let / be the length of the supply pipe. Let f be the acceleration of the piston. Then nf is the acceleration of the water in the supply-pipe. The force required to accelerate the piston 218 HYDRAULICS. and the corresponding pressure in feet of water W f ~~wAg' The force required to accelerate the water in the supply pipe wal : = ^ nf ' and the corresponding pressure in feet of water A. Similarly, if /' is the length of the discharge-pipe and its sectional area, the pressure-head due to the inertia of the discharge-water Hence the total pressure in feet of water required to over- come inertia in the supply-pipe and cylinder W The quantity - ;-)-#/ has been designated the length of working cylinder equivalent to the inertia of the moving parts. Let the engine drive a crank of radius r, and assume that the velocity V of the crank-pin is approximately constant. Then the acceleration of the piston when it is at a distance x from its central position F 2 HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 21 9 and the pressure due to inertia wA^ Let v be the velocity of the piston in the working cylinder. Let u be the velocity of the water in the supply-pipe. Let h be the vertical distance between the accumulator- ram and the motor. Let/ be the unit pressure at the accumulator-ram. Let/ be the unit pressure in the working cylinder. Then / & a _ / V* ( losses due to friction, sudden changes w 2g ~~ w 2g \ of section, etc. Thus A t v - -11 + losses. W 2g V U The term 1- losses may be approximately expressed o v 1 in the form K , AT being the coefficient of hydraulic resistance. Hence w 2g the term h being disregarded as it is usually very small as compared with . w Thus the total pressure-head in feet required to overcome inertia and the hydraulic resistances and is represented by the ordinate between the parabola ced 220 HYDRA ULICS. and the line ab in Fig. 136, in which afgb is a rectangle, ab representing the stroke 2r, ac = oa the pressure due to inertia at the end of the stroke, and F 2 the pressure required to overcome the hydraulic resistances at the centre of the stroke. 9 FIG. 136. The ordinate between the parabola fmg and the line fg represents the back pressure, which is necessarily proportional F a to the square of the piston-velocity, i.e., to (r* x*}. Hence the effective pressure-head on the piston, transmitted to the crank-pin, is represented by the ordinate between the curves amg and ced. The diagram shows that the pressure at the end of the stroke is very large and may become excessive. It is therefore usual to introduce relief-valves or air-vessels to prevent violent shocks. In certain cases, however, as, e.g., in a riveting-machine, a heavy pressure at the end of the stroke, just where it is most needed to close the rivet, is of great HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 221 advantage, and therefore the inertia effect is increased by the use of a supply-pipe of small diameter and an accumulator with a small water section (Fig. 134). The effective pressure should be as great as possible, and therefore the pressures due to inertia and frictional resistance, and the back pressure, which are each proportional to v*, should be as small as possible, and hence it is of importance to fix a low value for the speed of the piston, which in practice rarely exceeds 80 ft. per minute. The exhaust port should also be made of large area, as the back pressure diminishes as the area of the port increases. By equation I, (3) This speed v can be regulated at will by the turning of a cock, as in this manner the hydraulic resistances may be in- definitely increased. Let the engine be working steadily under a pressure P t and let v be the speed of steady motion. Then and _ j useful resistance overcome by the piston ( + friction between piston and accumulator-cylinder. If P is diminished, the speed V Q will be slightly increased, but in no case can it exceed, 4. Losses of Energy. The losses may be enumerated as follows : (a) The Loss L^ due to Piston-friction. It may be assumed that piston-friction consumes from 10 to 20 per cent of the total available work. 222 HYDRA ULICS. (b) The Loss Z, due to Pipe-friction. The loss of head in the supply-pipe of diameter and it may be assumed that nearly the whole of this is wasted. Hence the corresponding loss of head is ~" _/ /'W _W_ll_ ^_\^__ X n ' ri)~2 ~~ ^2r\n " ~n'} ^~~ ~2 % A2r \n ' ri2g ~~ 2rn n' g~~ 2g HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 22$ (d) The Loss L 4 due to Curves and Elbows. The losses due curves and elbows may be expressed in the form A =/ 4 (Chap. Ill, Art. 6). (e) The loss L 6 due to sudden Changes of Section. The loss of head in the passage of the water through the ports may be expressed in the form/' . The loss occasioned by valves may also be expressed by /f/ . Thus the total loss is The coefficient/" may be given any desired value between O and oo by turning a valve, so that any excess of pressure may be destroyed and the speed regulated at will. (/) The Loss L t due to the Velocity with which the Water leaves the Discharge-pipe. A = Hence the effective head ==-- (L^ + A - A + A + L 6 + ), and the efficiency = I - (L, + A + A + L< + L>). The volume of water used per stroke is a constant quan- tity, and the efficiency, which may be as great as eighty per cent when the engine is working under a full load, may fall below forty per cent when the load is light. 5. Brakes. Hydraulic resistances absorb energy which is proportional to the square of the speed. This property has 224 H YDRA ULICS. been taken advantage of in the design of hydraulic brakes for arresting the motion of a rapidly moving mass, as a gun or a train, of weight W. In Fig. 137 the fluid is allowed to pass from one side of the piston to the other through orifices in the piston. Let m be the ratio of the area of the piston to the effective area of the orifices. Let v be the velocity of the -piston when moving under a force P. Let A be the sectional area of the cylinder. FIG. 137. Then the work done per second = Pv = the kinetic energy produced and therefore P= wA(m i) 2 , and is the force required to overcome the hydraulic resistance at the speed v. Let V be the initial value of v, and P, the maximum value of P. Then P l = wA(m i) 2 *g Let F be the friction of the slide. Then o HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 22$ and P l -}- F is the maximum retarding force. It would cer- tainly be an advantage if the retarding force could be constant. In order that this might be the case (m i)v must be con- stant, and therefore as v diminishes m should increase and con- sequently the orifice area diminish. Various devices have been adopted to produce this result. Assuming the retarding force to be constant, let x be the piston's distance from the end of the stroke when its velocity is v. Then and therefore ^ 2 is proportional to x. But (m i)v is constant. Therefore (m i) is inversely proportional to 6. Water-wheels. Water-wheels are large vertical wheels which are made to turn on a horizontal axis by water falling from a higher to a lower level. These wheels may be divided into three classes : (a) Undershot Wheels, in which the water is received near the bottom and acts by impulse. (b) Breast Wheels, in which the water is received a little below the axis of rotation and acts partly by impulse and partly by its weight. (c) Overshot Wheels, in which the water is delivered nearly at the top and acts chiefly by its weight. 7. Undershot Wheels. Wheels of this class, with plane floats or buckets, are simple in construction, are easily kept in repair, and were in much greater use formerly than they are now. They are still found in remote districts where there is an abundance of water-power, and are also employed to work floating mills, for which purpose they are suspended in an open current by means of piles or suitably moored barges. They are made from 10 to 25 feet in diameter, and the floats, which are from 24 to 28 in. deep, are fixed either normally to the periphery of the wheel, or with a slight slope towards the supply-sluice, the angle between the float and radius being 226 HYDRA ULICS. from 1 5 to 30. Generally from one half to one third of the total depth of float is acted upon by the water. Let Fig. 138 represent a wheel with plane floats working in an open current. FIG. 138. Let v l be the velocity of the current. Let u be the velocity of the wheel's periphery. Let Q be the delivery of water in cubic feet per second. The water impinges upon a float, is reduced to relative rest, and is carried along with the velocity u. Thus the impulse = (#, u), o and wQ the useful work per second = - u(v l u). o Hence wQ . u(y. u) , x ^ /*= 2u(v. U) the efficiency = ^-^ - = v * a - '-, which is a maximum and equal to when u = v.. ^ l HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 22/ Theoretically, therefore, the wheel works to the best advan- tage when the velocity of its periphery is one half of the cur- rent velocity. Even then its maximum theoretic effect is only 50$, and in practice this is greatly reduced by frictional and other losses, so that the useful effect rarely exceeds 30$. Undershot wheels with plane floats are cumbrous, have little efficiency, and should not be used for falls of more than 5 feet. Again, let A be the water-area of a float, and w be the specific weight of the water. wQ is somewhat less than wAv^ , as there will be an escape of water on both sides of the float. Let wQ = kwAv lt k being some coefficient (< i) to be oletermined'by experiment. Then ^ the useful work per second = kAw l (y l u), o kA and its maximum value = - v.w. According to Bossut's and Poncelet's experiments a mean A *y value of k is , and the best effect is obtained when u = -v l , the corresponding useful work being - - - and the effi- 48 ciency , 125 8. Wheels in Straight Race. Generally the water is let on to the wheel through a channel made for the purpose, and closely fitting the wheel, so as to prevent the water escaping without doing work. For this reason also, the space between the ends of the floats in their lowest positions and the channel is made as small as is practicable and should not exceed 2 in. Hence /&, and therefore also the efficiency, will be increased. Assume the channel to be of a uniform rectangular section and to have a bed of so slight a slope that it may be regarded as horizontal without sensible error. 228 HYDRA ULICS. The wheel is usually from 24 to 48 ft. in diameter, with 24 to 48 floats, either radial or inclined. The floats are 12 to 20 inches deep, or about 2\ to 3 times the depth of the approach- ing stream. The fall should not exceed 4 ft. Let the floats be radial, Fig. 139. FIG. 139. Let h l be the depth of the water on the up-stream side of the wheel. Let //, be the depth of the water on the down-stream side of the wheel. Let , be the width of the race. The impulse = impulse due to change of velocity -|- impulse due to change of pressure g 2 and the useful work per second = impulse X u = ^u(v, - u) + ^ - *), g 2 Vtf, -Ul The second term is negative, since h^ > /i, , and tne maxi- mum theoretic efficiency may be easily shown to be <.5. Three losses have been disregarded, viz. : HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 229 (i) The loss of Q l cubic feet of the deeper fluid elements which do not impinge upon some of the foremost floats. According to Gerstner, o --= c -2( cQ ( *' V *,'U - u) ' .72, being the number of the floats immersed, and c being -J or v according as the bottom of the race is straight or falls .abruptly at the lowest point of the wheel. (2) The loss of <2 2 cubic feet of water which escape between the wheel and the race-bottom. Approximately, the play at the bottom may be said to vary from a minimum, s l = BC, when a float AB is in its lowest position, Fig. 140, to a maximum, B l C l = CD=^C t , when FIG. 140. two floats A l B l , A^Bs are equidistant from the lowest position, Fig. 140. Thus the mean clearance = J(25, + BD) = 5, +-, nearly, r l being the wheel's radius. 230 HYDRA ULICS. But - - = distance between two consecutive floats ft = 2 . B^D, very nearly, n being the total number of floats. Hence a and therefore the mean clearance = S l -\ --- *. Again, the difference of head on the up-stream and down stream sides and the velocity of discharge, v d , through the clearance is given by the equation Hence Introducing .7 as a coefficient of hydraulic resistance, ^ . / I TrVA a =.7,+--^ If the depth of the stream is the same on both sides of the wheel, i.e., if h, = & t , then (3) The loss of 03 cubic feet of water which escape between the wheel and the race-sides. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. Let J a be the clearance on each side. Then .7 being a coefficient of hydraulic resistance. Finally, if f^lbs. is the weight on the wheel-journals, the loss due to journal friction /* being the journal coefficient of friction, and p the journal radius. Thus the actual delivery of the wheel in foot-pounds These wheels are most defective in principle, as they utilize only about one third of the total available energy. They may be made to work to somewhat better advantage by introducing the following modifications: (a) The supply may be so regulated by means of a sluice- board, that the mean thickness of the impinging stream is about 6 or 8 inches. If the thickness is too small, the relative loss of water along the channel will be very great. If the thickness is too great, the floats, as they emerge, will have to raise a heavy weight of water. The sluice-board is inclined at an angle of 30 to 40 to the vertical, so that the sluice-opening may be as near the wheel as possible, thus diminishing the loss of head due to channel friction, and is rounded at the bottom to pre- vent a contraction of the issuing fluid. Neglecting frictional losses, etc., f i re /->/rr . v ? u *\ ( loss of energy the useful effect = wQ[H-\--^ -- J , _ f 7 \ 2 " 2 gl ( due to shock g 232 HYDRA ULICS. H being the difference of level between the point at which the water enters the wheel and the surface of the water in the tail- race, i.e., the fall. H is usually very small and may be negative. If the vanes are inclined, the resistance to emergence is not so great, and the frictional bed resistance between the sluice and float is practically reduced to nil. With a straight bed and small slope (i in 10) the minimum convenient diameter of wheel is about 14 ft. (b) The bed of the channel for a distance at least equal to the interval between two consecutive vanes may be curved to the form of a circular arc concentric with the wheel, with the view of preventing the escape of the water until it has exerted its full effect upon the wheel. When the bed is curved, the mini- mum convenient diameter of wheel is about 10 ft. An undershot wheel with a curb is in reality a low breast- wheel, and its theory is the same as that described in Arts. 13 and 14. (c) The down-stream channel may be deepened so that the velocity of the water as it flows away becomes > v r The im- pulse due to pressure is then positive, which increases the useful work and therefore also the efficiency. (d) The down-stream channel may be widened and a slight counter-inclination given to the bed. What is known as a standing-wave is then produced, in virtue of which there is a sudden rise of surface-level on the down-stream side above that on the up-stream side. This allows of the wheel being lowered by an amount equal to the difference of level between the sur- faces of the standing-wave and of the water-layer as it leaves the wheel, thus giving a corresponding gain of head. (e) The introduction of a sudden fall has been advocated in order to free the wheel from back-water, but it must be borne in mind that all such falls diminish the available head. Thus undershot wheels with plane floats have little effect because of loss of energy by shock at entrance and the loss of energy carried away by the water on leaving the floats. These losses have been considerably modified in Poncelet's wheel, which is often the best motor to adopt when the fall does HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 233 not exceed 6 ft., and which, in its design, is governed by two principles which should govern every perfect water-motor, viz. : (1) That the loss of energy by shock at entrance should be a minimum. (2) That the velocity of the water as it leaves the wheel should be a minimum. The vanes are curved and are comprised between two crowns, at a slightly greater distance apart than the vane- width ; the inner ends of the vanes are radial, and the water acts in nearly the same manner as in an impulse turbine. First. Assume that the outer end of a vane is tangential to the wheel's periphery, that the impinging layer is infinitely thin, and that it strikes a float tangentially. Let #/(Fig. 141) be a float, and aq the tangent at a. The velocity of the water relatively to the float = v l u. The water, in virtue of this velocity? ascends on the bucket to a height (" - V" pq , then falls back and FlG I4I < leaves the float with the relative velocity V 1 u and with an absolute velocity v l 2u. This absolute velocity is nil when the speed of the wheel is such that u = %i\, and the theoreti- i v 3 cal height of a float is/0 = -. The total available head is 42- thus changed into useful work, and the efficiency is unity, or perfect. Taking R as the mean radius of the crown and u l as the corresponding linear velocity, the mean centrifugal force on each unit of fluid mass is -~ and acts very nearly at the direc- tion of gravity, so that the height pq of a float may be approximately expressed in the form 'R 234 HYDRA ULICS. V being the velocity with which the water commences to rise on the float. Practically, however, the float is not tangential to the pe- riphery at a, as the water could not then enter the wheel. Also the impinging water is of sensible thickness, strikes the periph- ery at some appreciable angle, and in rising and falling on the floats loses energy in shocks, eddies, etc. Let the water impinge in the direction ac, Fig. 142, and take ac = v^ Take ad in the direction of and equal to , the velocity of the wheel's periphery. Complete the parallelogram bd. Then cd = ab = V is the velocity of the water relatively to the float. That there may be no shock at entrance, ab must be a tan- gent to the vane at a. FIG. 142. Again, the water leaves the vane in the direction of ba pro- duced, and with a relative velocity ae ab = V. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 23 $ Complete the parallelogram de. Then ag(=. v^ is the absolute velocity of the water leaving the wheel. Evidently cdg is a straight line. Let the angle cad = y, and the angle bad = n a. From the triangle adc, V* = v* -f u* 2v ji cos Y I (i) v? = V* -j~ u* 2 Vu cos a ; .... (2) V sin Y v, sin OL **/ From the triangle adg, By equations I, 2, and 4, ^, 8 ^ rr rra / = 2 Vu cos # = v l V u = 2u(v l cos Y u\ 2 Therefore the useful work per second = ^ 2U fa cos y - u ) (s> wQ v? cos 8 Y This is a maximum and equal to when V. COS Y rr u -, and the maximum emciency is cos y, Hence^ too, by equations I and 3, tan (n a) = 2 tan y (6) Also, V R sin , by equation 6. u sin (a -\- y} cos (n a] The efficiency is perfect if y is nil, and therefore a = 1 80. Practically this is an impossible value, but the preceding cal- culations indicate that ; should not be too large (usually < 30), and that the speed of the wheel should be a little less than one half of the velocity of the inflowing stream. 236 HYDRA ULICS. Take y = 15 as a mean value. Then u = v t X .484, and the efficiency = .993. Actually the efficiency does not exceed 68 per cent. In- deed it must be borne in mind that the theory applies to one elementary layer only, say the mean layer, and that all the other layers enter the wheel at angles differing from 15, thus giving rise to " losses of energy in shock." The losses of energy in frictional resistance, eddy motion, etc., in the vane passages, have also been disregarded. The layers of water, flowing to the wheel under an adjustable sluice and with a velocity very nearly equal to that due to the total head, may be all made to enter at angles approximately equal to 15, and the corresponding losses in shock reduced to a minimum by forming the course as follows : The first part of the course FG, Fig. 143, is curved in such a manner that the normal pqr at any point/ makes an angle of 15 with the radius^. The water moves sensibly parallel to the bottom FG, and therefore in a direction at right angles FIG. 143. to/r. Hence at q the direction of motion makes an angle of 15 with the tangent to the wheel's periphery. If or is drawn perpendicular to/r, then or = oq sin 15 = a constant. Thus the normal pqr touches at r a circle concentric with the wheel and of a certain constant diameter. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 2$? The initial point F of the curve FG is the point in which the tangent to this circle, passing through the upper edge of the sluice-opening, cuts the bed of the supply-channel. If t is the thickness (or depth of sluice-opening) and b the breadth of the layer of water as it leaves the sluice, then Q = btv, , and according to Grashof H being the available fall. The thickness should not exceed 12 to 15 inches, and is generally from 8 to 10 inches. Neglecting float thickness, the capacity of the portion of the wheel passing in front of the entering stream per second = bdu^ , very nearly. Only a portion of this space can be occupied by the water, so that Q mbdu l , m being a fraction whose value may be taken to be J or f Hence mbdu l btv^ , and therefore u. md u. t = md = cos y V l 2 r U md R = cos v . 2 r r, According to Morin, r, = 2d to $d. The mean velocity at entrance = c v < 2g(H /), an aver- age value of c v being .9. Thus \it = , HYDRAULICS. The diameter of the wheel is often taken to be The area of the sluice-opening is usually from \\bt to i.^bt. The inside width of the wheel is about (b + J) ft. The water should not rise over the top of the buckets, and in order to prevent this the depth of the shrouding is from J// to \H. If A is the angle subtended at the centre O of the wheel by the water-arc between the point of entrance A and the lowest point , Fig. 144, of the wheel, and if Aq' is drawn horizontally, then Aq' is approximately the height of the float, and the theoretic depth d of the crown is given by ' + OC - Oq' = AC = Aq f +Cq' = In practice it is usual to increase this depth by /, the thick- ness of the impinging water-layer. Again, 2 V" 1 d s -f r,(i cos A) -f- a few inches, approximately. The buckets are usually placed about I ft. apart, measured along the circumference, but the number of the buckets is not a matter of great importance. There are generally 36 buckets HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 239 in wheels of 10 to 14 ft. diameter, and 48 buckets in wheels of 20 to 23 ft. diameter. It may be assumed that the water-arc is equally divided by the lowest point C of the wheel, so that the length of the water-arc = 2\r = 2uT, T being the time of the ascent or descent of the water in the bucket. In the middle position, the upper end of the bucket should be vertical, and if the float is in the form of a circular arc, its radius r' = d sec (it a\ a being the angle between the bucket's lip and the wheel's periphery. The time of ascent or descent is also given by where sin fy = I/cos (it a). 9. Efficiency corresponding to a Minimum Velocity of Discharge (V 2 ). From Fig. 142, ao (= \ag) _ sin y __ Q a ) ad sin aod u Hence for any given values of u and y, v z is a minimum when sin aod is greatest, that is, when aod = 90, or ag is at right angles to de. Then also ad = ae = ab, or u = V, and ac bisects the angle bad. Thus, i7 1 = 2u cos y and v^ 2u sin y. The useful work W v? v? W WV/cos 2y = . -' - '- = 2u* cos 2y = -- 5- - , g 2 g g 2 COS' Y The total available work 240 H YDRA ULICS. Therefore the efficiency cos 2v - Ex. If y = 15, the efficiency = .928 and u = . In practice the best value of u is found to lie between. and .60^. The horse-power of the wheel rf being the efficiency with an average value of 60$. Although, under normal conditions of working, the effi- ciency of a Poncelet wheel is a little less than that of the best turbines, the advantage is with the former when working with a reduced supply. 10. Form of Bucket The form of the bucket is arbitrary, and may be assumed to be a circular arc. In practice there are various methods of tracing its form. METHOD I (Fig. 145), The tangent am to the bucket at a FIG. 145. makes a given angle a with the tangent at a to the wheel's outer periphery. The radius of\s also a tangent to the bucket HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 24! at/! If the angle aof\s known the position of f on the inner periphery is at once fixed, and the form of the bucket can be easily traced. Let the angle aofx. Join af and let the tangents to the bucket at a and /meet in m. Then the angle oam = a 90. " oma 1 80 oam aom = 270 a x. " m fa = the angle maf (180 fmd) = "+-*- 45- ' Let r lt r^ be the radii of the outer and inner peripheries of the wheel. Then sin (f!L _ 45) r l oa sin of a sin mfa \ 2 / of sin oaf sin oaf sin (^-45*) since the angle oaf ' = oam maf '= - 45. Hence r. X tan - 2 tan - an equation giving ;tr. The point o' in which the perpendicular o'f to 0/" meets the perpendicular o'a to am is the centre of the circular arc required and o'f(^o'd) is the radius. METHOD II (Fig. 146). Take mad = 150, and in ma pro- duced take ak = of. With k as centre and a radius equal to 242 HYDRAULICS. ao describe the arc of a circle intersecting the inner periphery in the point f. Join kf, of, and af. The two triangles aof and akf are evidently equal in every respect, and therefore the angle kaf is equal to the angle of a. Drawing ao' at right angles to ak and fo' tangential to the periphery at f, the angle 0'af(= kaf 90) is equal to the angle o'f a (= of a 90), and therefore o'a = o'f. Thus o' is the centre of the circular arc required and o'a (= o'f) is the radius. FIG. 146. 9- METHOD III (Fig. 147). Let the bed with a slope of, say, i in 10 extend to the point C, and then be made concentric with the wheel for a distance CC subtending an angle of 30 at the centre of the wheel. Let the mean layer, half way between the sloping bed and the surface of the advancing water, strike the outer periphery at the point /. Draw fk making an angle of 23 with of, and take fk equal to one half or seven tenths of the available fall, k is the centre of the circular arc required and /is its radius. II. Breast-wheels. These wheels are usually adopted for falls of from 5 to 15 feet, and for a delivery of from 5 to 80 cubic feet per second. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 243 The diameter should be at least 1 1 ft. 6 in., and rarely ex- ceeds 24 ft. The velocity (u) of the wheel's periphery is gen- erally from 3^ ft. to 5 ft. per second, the most useful average velocity being about 4^ ft. per second. The width of the wheel should not exceed from 8 to 10 ft. It is of great importance to retain the water in the wheel as long as possible, and this is effected by surrounding the water-arc with an apron, or a curb, or a breast, which may be constructed of timber, iron, or stone. Hence, too, the buckets may be plane floats, but they should be set at an angle to the periphery of the wheel, so as to rise out of the water with the least resistance (Art. 8). The depth of a float should not be less than 2.3 ft., and the space between two consecutive floats should be filled to at least one half, and even to two thirds, of its capacity. The head (measured from still water) over the sill or lip should be about 9 in. The play between the outer edge of the floats and the curb varies from in. in the best constructed wheels to 2 inches. The distances between the floats is from i^ to if times the head over the sill. 244 HYDRA ULICS. Breast-wheels are among the best of hydraulic motors, giving a practical efficiency which may be as large as 80 per cent. 12. Sluices. The water is rarely admitted to the wheel without some sluice arrangement, which may take the form of an overfall sluice (Fig. 148), an underflow sluice (Fig. 149), or a bucket or pipe sluice (Fig. 150). The pipe sluice is espe- cially adapted for a varying supply, being provided, for a certain vertical distance, with a series of short tubes, so in- clined as to ensure that the water enters the wheel in the right direction. Taking .85 as the mean coefficient of hydraulic resistance for these tubes, the head k l required to produce the velocity of entrance z> is and if H is the total available fall, = remainder of fall available for pressure-work. The profile AB in an overfall and an underflow sluice, should coincide with the parabolic path of the lowest stream- lines of the jet. The crest of the overfall should be properly curved, and the inner edges of the underflow opening should be carefully rounded so as to eliminate losses due to con- traction HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 24$ The underflow sluice-opening should also be normal to the axis of the jet. Let h^ be the head above the crest of an overfall sluice. Then 2 T. ' * Q = -cb, b^ being the width of the crest and c the coefficient of dis- charge. The width b l is usually 3 or 4 inches less than the width b of the wheel. From this equation and the depth of water over the crest or lip is usually about 9 inches. Again, the head h^= CD) required to produce the velocity v l at the point of entrance B is 10 10 per cent being allowed for loss due to friction. Thus the height of the crest A above B, the point of entrance, = AD = CD - CA = h, - ii *;/ 36 V 10 2g \2cb^2g)' But BA is a parabola with its vertex at A, and therefore, if B is the angle between the horizontal BD and the tangent the parabola at B, n f\ A V, sm u 1 1 v* 2g ~ 10 2g y ) 246 HYDRA ULICS. Also v. sin 26 The head available for pressure work = DE = FG = H - h,. Let a be the angle between BT and the tangent to the wheel's periphery at B. Then a _f = the angle EOF, BO being the radius to the centre of the wheel and OFG' vertical. % If the lowest point G' of the wheel just clears the tail- race, the head available for pressure work = H - h, = FG' OG' - OF = rfr _ cos BOF) = 2r, si r, being the radius to the outer periphery of the wheel. If, again, the water enters the wheel tangentially, a = o, and the angle BOF = B, so that H - h, = 2r, sin 2 -. If the sluice-opening is not at the vertex of the parabola, the axis of the opening should be tangential to the parabola. 13. Speed of Wheel. The water leaves the buckets and flows away in the race with a velocity not sensibly different from the velocity u of the wheel's periphery. Let b be the breadth of the wheel (Fig. 151). Let x be the depth of the water in the lowest bucket. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 247 FIG. 151. Allowing for the thickness of the buckets, the play between the wheel and curb, etc., Q = cbxu, c being an empirical coefficient whose average value is about .0. Hence 10 Q u = jr. 9 ox In practice b is often taken to be to . It is impor- tant that b should be as small as possible and hence x should be as large as possible, its value being usually ij ft. to 2 ft. It must be borne in mind, however, that any increase i-n the value of x will cause an increase in the weight of water lifted by the buckets as they emerge from the race, and will therefore tend to diminish the efficiency. 14. Mechanical Effect. Theoretically, the total mechan- ical effect 248 HYDRA ULTCS. H being the fall from the surface of still water in the supply- channel to the surface of the water in the tail-race. This, however, is reduced by the following losses: (a) Owing to frictional resistance, etc., there is a loss of v 3 head in the supply-channel which may be measured by ^-7- v being approximately JL to T L. The head required to produce the velocity at entrance, v l9 (b) Let af, Fig. 152, represent in direction and magnitude v, the velocity of the water entering the bucket. FIG. 152. Let ad, in the direction of the tangent to the wheel's periphery, represent the velocity u of the periphery in direction and magnitude. Complete the parallelogram bd. Then ab evidently repre- sents the velocity V of the water relatively to the wheel. This velocity V is rapidly destroyed, the corresponding loss of head being F 2 U*-\-V? 2UV^ COS y being the angle daf. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 249 Assuming that the water enters the race with the velocity u of the wheel, the theoretical useful work per pound per second due to impact u. = -(v l cos y u). g V^ If the loss is to be a minimum for a given speed of o wheel, v,dv^ u cos y . dv l = o, or ^ u cos y. . . (2) Hence, by equation I, V = u sin 7, and therefore V df tan y = - = 2 v, af so that for a velocity of entrance v t = u cos y the angle afd should be 90. But this value is inadmissible, as the water would arrive tangentially and consequently would not enter the buckets. In Order that the loss in shock at entrance may be as small as possible, ab, the direction of the relative velocity F, should be parallel to the arm xy of the bucket, and should therefore be approximately normal to the wheel's periphery. This is equivalent to the assumption that the water arrives in a given direction (y) with a given velocity (^), and that the speed (?/) of the wheel is to be such as will make V a mini- mum. Thus, by equation I, o udu v^ cos y . du, or u = v l cos y, and therefore V = v l sin y. Hence tan y = -, and therefore the angle adf = 90. u ad 250 HYDRAULICS. In practice y is generally 30, and the corresponding loss of F a v? . v> i if i head = = sin 2 y = -. - = . - At point of entrance x falls below y, the water flows up the inclined plane xy, and F, instead of being wholly destroyed in eddy motion, is partially destroyed by gravity. This velocity, destroyed by gravity, is again restored to the water on its return, and thus adds to the efficiency of the wheel. It will be found advantageous to use curved or polygonal buckets and not plane floats. A bucket, for ex- ample, may consist of three straight portions, ab, be, cd, Fig. 153. Of these the inner portion cd shoud be radial ; the outer portion ab is nearly normal to the periphery of the wheel, and the central portion be may make angles of about 135 with ab and cd. Disregarding all other losses, the theoretical delivery of the wheel in foot-pounds where h^ = total fall fall (h^ required to produce the veloc- ity v,. If 77 be the efficiency, then, according to the results of Morin's experiments, rf = .40 to .45 if h^ = -//"; 4 rf = .42 to .49 if h l = H\ rj = .47 if h, = -H; 3 if h, = ff. 4 HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 2$ I (c) There is a loss of head due to frictional resistance along the channel in which the wheel works. Let / = length of the channel (or curb). Let t thickness of water-layer leaving the wheel. Let b = breadth of wheel. The mean velocity of flow in this curb channel is approxi- mately -u, and the loss of head due to channel friction bt 2g where/ = coefficiency of friction, b -f- 2t = wetted perimeter, bt = water area, and y being 30. (d] There is a loss of head due to the escape of water over the ends and sides of the buckets. Let s 1 be the play between the ends of the buckets and the channel. Let s^ be the play at the sides. (^, = J a , approximately.) Let z l , # 2 , . . . z n be the depths of water in a bucket corre- sponding to n successive positions in its descent from the receiving to the lowest points. Let / a , / a , ... l n be the corresponding water-arcs measured along the wheel's periphery. The orifice of discharge at end of a bucket = bs^ The mean amount of water escaping from a bucket over its end c being the coefficient of discharge. The water escapes at the sides as over a series of weirs, and the mean amount of water escaping from a bucket over the sides 252 HYDRAULICS. Hence the total loss of effect from escape of water per sec., ^ being the vertical distance between the point of entrance and the surface of the water in the tail-race __. (e) There is a loss of head due to journal friction. Let W = weight of wheel. Let w l = weight of water on the wheel. Let r l = radius of wheel's outer periphery. Let r 1 radius of axle. Loss per second of mechanical effect due to journal friction r being the coefficient of journal friction. There is a loss of mechanical effect due to the resistance of the air to the motion of the floats (buckets), but this is prac- tically very small, and may be disregarded without sensible error. A deepening of the tail-race produces a further loss of effect, and should only be adopted when back-water is feared. Hence the total actual mechanical effect, putting Z=b Sl ( V^ cos ,s = HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 253 =wQ ff- (i + v) + fa cos r - ) --*(", cos y-u) Hence, for a given value of z/,, the mechanical effect (omit- ting the last term) is a maximum when = ^ C S Y (= -433 X ^ , if r = 30). In practice the speed of the wheel is made about one half of the velocity with which the water enters the wheel. For a given speed of wheel, and disregarding the loss of effect due to curb friction, which is always small, the mechani- cal effect is a maximum for a value of z/, given by I ^ t/ w 'Z\ l + v i W Q \wQ c V2g 1 ! v l H -u cos Y = o, or U COS Y The loss by escape of water, viz., c V2g, varies, on an average, from 10 to 15 per cent of the whole supply, so that c V2g- varies from to 2s, d n 10 20 254 JfYDRA ULICS. 15. Sagebien Wheels have plane floats inclined to the radius at from 40 to 45 in the direction of the wheel's rota- tion. The floats are near together and sink slowly into the fluid mass. The level of the water in the float-passages grad- FIG. 154. ually varies and the volume discharged in a given time may be very greatly changed. The efficiency of these wheels is over 80 per cent, and has reached even 90 per cent. The action is almost the same as if the water were transferred from upper to lower race, without agitation, frictional resistance, etc., flowing away without obstruction, into the tail-race. 16. Overshot Wheels. These wheels are among the best of hydraulic motors for falls of 8 to 70 ft. and for a delivery of 3 to 25 cub. ft. per second. They are especially useful for falls of 12 to 20 ft. The efficiency of overshot wheels of the best construction is from .70 to .85. If the level of the head-water is liable to a greater variation than 2 ft., it is most advantageous to employ a pitch-back or high breast-wheel, which receives the water on the same side as the channel of approach. 17. Wheel-velocity. This evidently depends upon the work to be done, and upon the velocity with which the water HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 255 arrives on the wheel. Overshot wheels should have a low circumferential speed, varying from 10 ft. per sec. for large wheels to 3 ft. per sec. for small wheels, and should not be less than 2-J ft. per sec. In order that the water may enter the buckets easily, its velocity should be greater than the peripheral velocity of the wheel. 18. Effect Of Centrifugal Force. Consider a molecule of weight W in the " unknown" surface of the water in a FIG. 155. bucket (Fig. 155). At each moment there is a dynamical equilibrium between the " forces" acting on m, viz.: (i) its 256 HYDRA ULICS. IV weight w\ (2) the centrifugal force coV; (3) the resultant T o of the neighboring reactions. 2V Take MF = w, MG = coV, and complete parallelogram o FG. Then MH = T. The direction of T is, of course, normal to the surface of the water in the bucket. Let HM produced meet the vertical through the axis O of the wheel in E. Then w_ a MG z** r FH OM r MF~ w ~MF~OE"OE' and therefore OB =*, = GO taking g = 32 ft. and n being the number of revolutions per minute. Thus the position of E is independent of r and of the position of the bucket, so that all the normals to the water- surface in a bucket meet in E, and the surface is the arc of a circle having its centre at E, or, rather, a cylindrical surface with axis through E parallel to the axis of rotation. 19. Weight of Water on Wheel and Arc of Discharge. Let Q = volume supplied per sec., and N = number of buckets. Noo Then - - = number of buckets fed per sec., 27T and = volume of water received by each bucket per sec. Hence the area occupied by the water until spilling com- mences = , ., , b being the bucket's width (= width of wheel between the shroudings). The water flows on to the wheel through a channel (Fig. 156), usually of the same width b as the wheel, and the HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. supply is regulated by means of an adjustable sluice, which may be either vertical, inclined, or horizontal. When the water springs clear from the sluice, as in Fig. 156, the axis of the sluice should be tangential to the axis of the FIG. 156. jet, and the inner edges of the sluice-opening should be rounded so as to eliminate contraction. Let y, z be the horizontal and vertical distances between the sluice and the point of entrance. Let T be the time of flow between the sluice and entrance. Let v , 2\ be the velocities of flow on leaving the sluice and on entering the bucket. Then 258 H YDRA ULICS. and V? = V* + 2gZ, d being angular deviation of point of entrance from summit, and y the angle between the direction of motion of the water and the wheel at the point of entrance. Assume the bed of the channel to be horizontal, and the sluice vertical and of the same, width b as the wheel. The sluice is also supposed to open upwards from the bed. Then x being the depth of sluice-opening and h^ the effective head over the sluice. This effective head is about T Vths of the actual head. Thus, taking g=. 32, = %xh$ gives the delivery per foot width of wheel. Taking .6 ft. and 3.6 ft. as the extreme limits between which h l should lie, and .2 ft. and .33 ft. as the extreme limits between which x should lie, then ~ must lie between the o limits 1.24 and 5, and an average value of ^ is 3. Thus the width of the wheel should be on the average ^ . Again, neglecting the thickness of the buckets, the capacity of the portion of the wheel passing in front of the water-sup- ply per second = b<*> \ - - ! = Mfafr, -- J = bdrja, approximately, , , Lj = bdu. bd 30 r, being the radius and u l the velocity of the outer circumfer- ence of the wheel, d the depth of the shrouding, and n the number of revolutions per minute. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 259 Only a portion, however, of the space can be occupied by the water, so that the capacity of a bucket is mubd, m being a fraction less than unity and usually -J or J. For very high wheels m may be \. Hence , , 27tQ mbdu. = ~=. NGO Again, since the thickness of the buckets is disregarded, Nu Therefore mdu. = ^. b The delivery \^j per foot of width must not exceed a certain limit, otherwise either d or u will be too great. In the former case the water would acquire too great a velocity on entering the buckets, which would lead to an excessive loss in eddy motion and a corresponding loss of efficiency ; while if the speed u of the wheel is too great the efficiency is again diminished and might fall even below 40$. The depth of a bucket or of the shrouding varies from 10 to 1 6 in., being usually from 10 to 12 in., and the buckets are spread along the outer circumference at intervals of 12 to 14 inches. The number of the buckets is approximately $r or 6r, r being the radius of the wheel in feet. The efficiency of the wheel necessarily increases with the number of the buckets, but the number is limited by certain considerations, viz. : (a) the bucket thickness must not take up too much of the wheel's periphery ; (b) the number of the buckets must not be so great as to obstruct the free entrance of the water; (c) the form of the bucket essentially affects the number. Let the bucket, Fig. 157, consist of two portions, an inner portion be, which is radial, and an outer portion cd\ c being a point on what is called the division circle. The length be is usually one half or two thirds of the depth d of the shrouding. 260 HYDRA ULICS. Take be = \d. It may also be assumed without much error that the water- surface ad is approximately perpendicular to the line ed t so that the angle edc is approximately a right angle. The spilling evidently commences when the cylindrical sur- face, having its axis at e and cutting off from the bucket a water-area equal to -~, passes through the outer edge d of Noo the bucket. FIG. 157. Let /3 be the bucket angle cOd. Let be the inclination of Od to the horizon. Let be the inclination of ad to the horizon. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 26 1 Let r l be the radius of the outer periphery. Let R be the radius of the division circle. Let r a be the radius of the inner periphery. Then od^ __ r l _ sin _ sin oe ""^"sin j 90 0+0} ~ cos (0+0)' and therefore Again, Therefore sin af = fd tan (0 -|- 0), approximately. the area dfa=< tan (0 + 0) = tan (8 + 0), 2 2 where d = r l r 2 . Hence the area abed = area cod area bof area ^/iz Equations (i) and (2) give and 0, and therefore the posi- tion of the bucket when spilling commences. The bucket will be completely emptied when it has reached a position in which cd is perpendicular to a line from e to middle point of cd, or, approximately, when edc is a right angle. Let 0,, 0, be the corresponding values of and 0, and let 262 HYDRA ULICS. y t be the angle between cd and the tangent at d to the wheel's periphery. Then and = 90 - sn r, ._. g sin r two equations giving 0, and 0^ Also, if ^ is drawn perpendicular to od, de r R cos tan y = cot <:# . 9 Jt ^Sb*.^^ Z, .. (4) W These two last equations give N and 0. The number of buckets may also be approximately found from the formula In practice the bucket may be delineated as follows : Let dd r = distance between two buckets. 56 d Take dd" = ~ dd' to - dd'\ also take fo = -, and join dc. This gives the form of a suitable wooden bucket. FIG. 158. If the bucket is of iron, a circular arc is substituted for the portions be, cd. Again, let/w, Fig. 159, be the thickness of the stream just before entering the bucket. Let dn be the thickness of the stream just after entering the bucket. Let \ be the angle between the bucket's lip and the wheel's periphery. 264 HYDRA ULICS. Then mbdu l capacity of bucket = bv^ . pm = bV. dn = bv^dp sin y = b V. dp . sin A, and therefore ~ v.smr" FsinA' Now overshot wheels cannot be ventilated, and it is conse- FIG. 159. quently necessary to leave ample space above the entering stream for the free exit of air. Thus, neglecting float thick- ness, ' = distance between consecutive floats = and N, the number of buckets, 2 Try, F sin \ mdu, HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 26$ For efficient action the number of the buckets is much less than the limit given by this relation, often not exceeding one half of such limit. If y is very small, V=v l u^ approximately, and therefore The capacity of a bucket depends upon its form ; and the bucket must be so designed that the water can enter freely and without shock, is retained to the lowest possible point, and is finally discharged without let or hindrance. Hence flat buckets, Fig. 160, are not so efficient as the curved iron bucket in Fig. 163 and as the compound bucket made of three or two FIG. 1 60. FIG. 161. FIG. 162. FIG. 163. FIG. 164. pieces in Figs. 161, 162, 164. The resistance to entrance is least in the curved bucket, as there are no abrupt changes of direction due to angles. The capacity of a compound bucket may be increased, without diminishing the ease of entrance, by making the inner portion strike the inner periphery at an 266 HYDRA ULICS. acute angle, Fig. 164. The objection to this construction, especially if the relative velocity V is large, is that the water tends to return in the opposite direction and escape from the bucket. Let bed, efg, Fig. 165, represent two consecutive buckets of an overshot wheel turning in the direction shown by the arrow. FIG. 165. Water will cease to enter the bucket-space between efg, and impact will therefore cease, when the upper parabolic boundary of the supply-stream intersects the edge b. The last fluid elements will then strike the water already in the bucket at a point M, whose vertical distance below b may be desig- nated by z. The velocity v' with which the entering particles reach M is given by the equation (0 HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 267 Again, while the fluid particles move from b to M let the buckets move into the positions b'c'd' , e'f'g'. Let arc bb' = s 1 = ee r . Let arc bM = s t . Let T be the time of movement from b to b' (or b to M\ Then s. = uT and assuming that the mean velocity from b to M is an arithmetic mean between the initial and final velocity of entrance. Thus l -f- ^i Also, since the angle between bM and the wheel's periphery is small, it may be assumed that the arc bM ' = be -\- ef-\- ee' y approximately, 27tr, N N u , +**' /, T r^ 7 , V i U 27Cr i V i U \ (Note.ef eb = eb- - = -^T. - - , nearly.) \ J u u N u J i Thus and by equations 2 and 3, ( v i + v *' 2U \ _ 27tr i !!L S \ 2u I ~ N u> 268 HYDRA ULICS. an equation giving approximately the distance s l passed through by a float during impact. The buckets can now be plotted in the positions they occupy at the end of the impact. The amount of water in each bucket being also known, the water-surface can be delineated, and hence the vertical distance x can be at once found. 20. Useful Effect (a) Effect of Weight. The wheel should hang freely, or just clear the tail-water surface, and the total fall is measured from the surface of the water in the tail-race to the water-surface just in front of the sluices through which the water is brought on to the wheel. FIG. 1 66. Let h lt Fig. 166, be the vertical distance between the cen- tres of gravity of the water-areas of the first and last buckets before spilling commences. Then //, = R cos d -\- r l sin 0, very nearly. Let h^ be the vertical distance between the centres of gravity of the water-area of the bucket which first begins to spill, and the point at which the spilling is completed. Then h^ r,(sin 0, sin 0), very nearly. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 269 The useful work per sec. = '^Q(h l + kh^ k being a frac- tion < I and approximately = .5. Let A be the water-area in the bucket which first begins to spill. Between this bucket and the one which is first emptied, i.e., in the vertical distance /z 2 , insert an even number s of buckets, and let their water-areas A l , A 9 , A 3 , . . . A s be care- fully calculated. Let Q m be the mean amount of water per bucket in the discharging arc. Let A m be the mean water-area per bucket in the discharg- ing arc. Then The value of k can now be easily found, since Q m _A m ~-~" Let q be the varying amount of water in a bucket frorrr which spilling is taking place, and at any moment let y be the vertical distance between the outer edge of the bucket and the surface of the water in the tail-race. q is a function of y and depends upon the contour of the water in the bucket. Let Y be the mean value of y between the points where spilling begins and ends, i.e., for values^, and j/ a of y. Then y\ since Jy .dq=yq Jq . dy. 2/O HYDRA ULICS. Again, the elementary quantity of water, dq, having an initial velocity equal to that of the wheel, viz., &, falls a dis- tance y and acquires a velocity = useful work _ v V The reaction = linear ve i oci t y of rotation = g For a maximum efficiency = o = HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 2?$ Hence f 2Z/F+ *.?" = o, and therefore v = V(i + Vi - c,') ....... (4) Experience indicates that the greatest efficiency corresponds to a speed of rotation equal to the velocity due to a head h, i.e., to a value of V given by . - , ..... (5) By equations (i) and (5) f = 4V** ....... (6) and therefore, by equations (4), (5), and (6), o c* = ~ or c, = .94 ....... (7) Hence, by equations (3), (5), (6), and (7), the maximum efficiency = . o Thus one third of the head is lost, and of this amount the ( v F) 2 / h\ portion --- ^= -j is carried away by the effluent water. The portion - -- (= -kj is lost in frictional resistance, etc. Again, = j t | cjj* + -yT terms cont'g higher powers of -~\ i | . 276 HYDRA ULICS. The efficiency therefore increases with F, but the value of V is limited by the practical consideration that, even at moderately high speeds, so much of the head is absorbed by friction as to sensibly diminish the efficiency. The serious practical defects of this wheel are that its speed is most unstable and that it admits of no efficient system of regulation for a varying supply of water. The Scotch or Whitelaw's tur- J 73. bine, Fig. 173, excepting in the curved arms, does not differ essentially from the reaction wheel just considered. 23. Reaction and Impulse Turbines. All turbines be- long to one of two classes, viz., Reaction Turbines and Impulse Turbines, and are designed to utilize more or less of the avail- able energy of a moving mass of water. In a reaction turbine a portion of the available energy is converted into kinetic energy at the inlet surface of the wheel. The water enters the wheel-passages formed by suitably curved vanes, and acts upon these vanes by pressure, causing the wheel to rotate. The proportions of the turbine are such that there is a particular pressure (hence the term pressure- turbine) at the inlet surface corresponding to the best normal condition of working. Any variation from this pressure, caused, e.g., by the partial closure of the passages through which the water passes to the wheel, changes the working con- ditions and diminishes the efficiency. In order to avoid such a variation of pressure, it is essential that there should be a continuity of flow in every part of the turbine ; the wheel- passages should be kept completely filled with water, and therefore must receive the water simultaneously; Such turbines are said to have complete admission. The admission is partial when the water is received over a portion of the inlet surface only. In an impulse (Girard) turbine, Figs. 174, 175, the energy HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. of the water is wholly converted into kinetic energy at the inlet surface. Thus the water enters the wheel with a velocity due to the total available head and therefore without pressure, is received upon the curved vanes, and imparts to the wheel the whole of its energy by means of the impulse due to the FIG. 174. FIG. 175. Girard Turbine for Low Falls. Girard Turbine for High Falls. gradual change of momentum. Care must be taken to ensure that the water may be freely deviated on the curved vanes, and hence such turbines are sometimes called turbines with free deviation. For this reason the water-passages should never be completely filled, and the water should flow through under a pressure which remains constant. In order to ensure an un- broken flow through the wheel-passages and that no eddies are formed at the backs of the vanes, ventilating holes are arranged in the wheel sides, Fig. 177. Figs. 176 and 177 also show the relative path AB and the absolute path CD traversed by the water in an inward-flow and a downward-flow turbine. If there is a sufficient head, the wheel may be placed clear 2 7 8 HYDRAULICS. above the tail-water, when the stream will be at all times under atmospheric pressure. With low falls the wheel may be placed in a casing supplied with air from an air-pump by which the surface of the water may be kept at an invariable level below the outlet orifices, which is essential for per- fectly free deviation. While the wheel-passages of a reaction tur- bine should be kept completely 'filled with water, no such restric- tion is necessary with an impulse turbine. The supply may be par- tially checked and the water may be received by one or more vanes without affecting the efficiency. ' Thus the dimen- sions of an impulse turbine may vary between very wide TAIL WATER FIG. 177. limits, so that for high falls with a small supply, a compara- tively large wheel with low speed may be employed. The speed of a reaction turbine under similar conditions would be disadvantageously great, and any considerable increase of the diameter would largely increase the fluid friction and would also render the proper proportioning of the vane-angles almost impracticable. Impulse turbines may have complete or partial admission, while in reaction turbines the admission should be always complete, as in Fig. 178, which shows the HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. relative path AB and absolute path CD traversed by the water. When there is an ample supply of water the reaction turbine is usually to be preferred, but on very high falls its speed FIG. 178. becomes inconveniently great and it is then better to adopt a turbine of the impulse type. The diameter of the wheel can then be increased and the speed proportionately diminished. The Hurdy-gurdy is the name popularly given to an impulse wheel which was introduced into the mining districts of California about the year 1865. Around the periphery of the wheel is arranged a series of flat iron buckets, about 4 to 6 in. in width, which are struck normally by a jet of water often not more than three eighths of an inch in diameter. Theoretically, the efficiency of such an arrange- ment cannot exceed 50 per cent (Art. 7), while in prac- tice it rarely reaches 40 per cent. The best speed of the wheel, in accordance with both theory and practice, is one half of that of the jet. Although the efficiency is so low, the wheel found great favor for many reasons. Any required speed could be obtained by a suitable choice of diameter ; the plane of the wheel could be placed in any convenient position ; the wheel could be cheaply constructed and was largely free from liability to accident. Hence it was of the utmost importance to increase, if possible, the efficiency of a wheel possessing such advantages. Obviously a first step was to substitute cups for the flat buckets, the immediate result necessarily being a very large increase in the efficiency. This was increased still further by the adoption of double 2 80 H YDRA ULICS. buckets, Fig. 179, that is, curved buckets divided in the middle so that the water is equally deflected on both sides. Thus developed, the wheel is widely and most favorably known as the Pelton wheel, Fig. 179. Its efficiency is at least 80 per cent, and it is claimed that it often rises above 90 per cent. The power of the wheel does not depend upon its diameter, but upon the available quantity and head of water. The water passes to the wheel through one or more nozzles, FIG. 179. having tips bored to suit any required delivery. These tips are screwed into the nozzles and can be easily and rapidly replaced by others of larger or smaller size, so that the Pelton is especially well adapted for a varying supply of water. It is claimed that in this manner the power may be varied from a maximum down to 25 per cent of the same without appreci- able loss of efficiency. The character of the construction of turbines has led to their being classified as (i) Radial-flow turbines; (2) Axial- flow turbines ; (3) Mixed-flow turbines. In Radial-flow turbines the water flows through the wheel in a direction at right angles to the axis of rotation and approximately radial. The two special types of this class are the Outward-flow turbine, invented by Fourneyron, and the Inward-flow or Vortex turbine, invented by James Thomson. In the former, Figs. 180 and 181, the water enters a cylindrical chamber and is led by means of fixed guide-blades outwards HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 28 1 from the axis. It is distributed over the inlet-surface, passes through the curved passages of an annular wheel closely sur- FIG. 1 80. FIG. 181. rounding the chamber, and is finally discharged at the outer surface. The wheel works best when it is placed clear above 282 HYDRA ULICS. the tail-water. A serious practical defect is the difficulty of constructing a suitable sluice for regulating the supply over the inlet-surface. Fourneyron was led to the design of this turbine by observing the excessive loss of energy in the ordi- nary Scotch turbine, or reaction wheel, and introduced guide- blades in order to give the water an initial forward velocity and thus cause a diminution of the velocity of the water leav- ing the outlet-surface. In the Inward-flow or Vortex turbine, Figs. 182 and 183, the wheel is enclosed in an annular space, into which the water flows through one or more pipes, and is usually dis- tributed over the inlet-surface of the wheel by means of four guide-blades. The water enters the wheel, flows towards the space around the axis, and is there discharged. This turbine possesses the great advantage that there is ample space outside the wheel for a perfect system of regulating-sluices. Axial- flow turbines, Figs. 184, are also known as Parallel and Downward-flow turbines and are sometimes called by the names of the inventors, Jonval and Fontaine. In these the water passes downward through an annular casing in a direction parallel to the axis of rotation, and is distributed by means of guide-blades over the inlet-surface of an adjacent wheel. It enters the wheel-passages and is finally discharged vertically, or nearly so, at the outlet-surface. The sluice regulations are worse even than in the case of an outward-flow turbine, but there is this advantage, that the turbine may be placed either below the tail-water, or, if supplied with a suction-pipe, at any point not exceeding 30 ft. above the tail-water. If a turbine is designed so that the pressure at the clear- ance between the casing and the wheel is nil, and with curved passages in the form of a freely deviated stream, it becomes what is called a Limit turbine. In its normal condition of working it is an Impulse turbine, but when drowned, it is a Reaction turbine, with a small pressure at the clearance. For moderate falls with a varying supply its average efficiency is higher than that of a pressure turbine. The Mixed- or Combined-flow (Schiele) turbine is a combi- HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 283 nation of the radial and axial types. The water enters in a nearly radial direction and leaves in a direction approximately FIG. 182. X ///////////////^^^^^ 1 --T FIG. 183. parallel to the axis of rotation. This type of turbine admits of a good mode of regulation and is cheap to construct. 24. Theory of Turbines (Figs. 185 to 188). Denote in- 284 HYDRA ULICS. ward-flow, outward-flow, and axial-flow turbines by I. F., O. F., and A. F., respectively. FIG. 184. Let r,, r a be the radii of the wheel inlet and outlet surfaces or an I. F. or O. F. Let r lt r t be the outer and inner radii of the wheel inlet- surface of an A. F. Let R be the mean radius \== r * "^ r *J of an A. F., assumed constant throughout. FIG 185. Section of an inward-flow turbine. Let A lf A, be the areas of the wheel inlet and outlet orifices. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 285 FIG. 186. Enlarged portion of the section through XY, Fig. 185. FIG. 187. Enlarged portion of a section through XY, Fig. 180, of an outward- flow turbine. FIG. 188. Enlarged portion of a cylindrical section XY, Fig. 184, of a down- ward-flow turbine developed in the plane of the paper. 286 HYDRAULICS. Let d lt d t be the depths of the same in an I. F. or O. F. Let d lt d^ be the widths of the same in an A. F. Let h be the thickness of the wheel in an A. F. Let H l be the effective head over the inlet-surface of the wheel. This is the total head over the inlet- surface diminished by the head consumed in frictional resistance in the supply-channel, and by the head lost in bends, sudden changes of section, etc. Let HI be the fall from the outlet-surface to the surface of the water in the tail-race. If the turbine is submerged, then H 9 is negative. Let v lt v t be the absolute velocities of the water at the inlet- and outlet-surfaces. Let u lt #, be the absolute velocities of the inlet- and outlet- surfaces. Let V^ Vi be the velocities of the water relatively to the wheel, at the inlet- and outlet-surfaces. Let GO be the angular velocity of the wheel. Let the water enter the wheel in the direction ac t making an angle y with the tangent ad. Take ac to represent v l and ad to represent u lt Complete the parallelogram bd. The side ab represents V lt and in order that there may be no shock at entrance, ab must be tangential to the vane at a. Again, at/ drawy^-, a tangent to the vane, and//, a tangent to the wheel's periphery. Take fg and fk to represent V^ and u^ respectively. Com- plete the parallelogram gk. The diagonal /$ must represent in direction and magnitude the absolute velocity v^ with which the water leaves the wheel. Let the angle hfk = d. The tangential component of the velocity of the water as it enters or leaves the wheel is termed the velocity of whirl, and the radial component the velocity of flow. Denote these components respectively by vj, v r ' at the inlet-surface ; v' i v r " at the outlet-surface. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 28/ Let the angle bad = 1 80 a, Let the angle gfk = 180 ft. Draw cm perpendicular to ad, and hn to fk. Then at the inlet-surface, vj=. v^ cos y ac cos y = #;# = ad^dm = #, P, cos a ; (i) ?>/ = z/, sin ^ ;# = V l sin or ; (2) and at the outlet-surface vj' = z> a cos 6 =fn =.fk kn = u 9 V^ cos /? ; . (3) v r " = ^ 2 sin 6 = /* F 2 sin /? (4) Let g be the volume of water passed per second. Then in an I. F. or O. F. Vr'Ai = VrZTtridi = Q (5) in an A. F. i = Q (5) In equations (5) the thickness of the vanes has been disre- garded. If is the angle between the vane, of thickness BC, A / and the wheel's periphery AB, then the space ^f^j occupied by the vane along the wheel's periph- / / ery is AB = BC cosec 0. / Let n be the number of the guide-vanes and / FlG - I8 9- their thickness. Let #, be the number of the wheel-vanes and /, , / 2 their thickness at the inlet- and outlet-surfaces, respect- ively. Then, in a radial-flow turbine, A l -fad l \2nr l nt cosec y n v t l cosec a\ . . (6) and ^. = TWi 2 ^.- *i** cosec ft\> ...... (7) T 9 being a fraction depending on practical considerations. 288 HYDRAULICS. In an axial-flow turbine R is to be substituted for r l ind r y in the values of A l and A 9 . n l may be made equal to n -f- I or n -f- 2. Again, as the water flows through the wheel its angular momentum relatively to the axis of rotation is changed from rjsj at the inlet- to rj)J' at the outlet-surface. o o Hence, if T is the effective work done by the water on the turbine, and GO the angular velocity of the turbine, in an I. F. or O. F. in an A. F. T - ^(vv'n - v w "r t ) since since Ui 2 r*= = > - - (9) '1 * 2 and the hydraulic efficiency T v w 'ui - z>"w a and the hydraulic r/ f \Viv efficiency -^")i / I0 x wQH, gff, ' ( wQ(H, + A) g(i yi + A) ' Equation 10 is the fundamental equation upon which the whole design of turbines depends. From the triangle abc, V* = v* + u* 2v l u l cos y, . . . . (ii) and sn y sin of (12) From the triangle ./M, , 1 = , 1 +F, 1 -2,r,COS/ (I 3 ) HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 289 Again, if , -- are the pressure-heads at the inlet- and outet- w w surfaces of the wheel of a REACTION TURBINE, A -A =ff. - w (14) In an IMPULSE TURBINE the water is under atmospheric pressure only, and therefore 05) To make allowance for hydraulic resistances k^ may be o v* substituted for in equations 14 and 15, a mean value of k l o ^ ' I0 being . ' 9 Applying Bernoulli's theorem to the filament from a to/, % * _ u and taking account of the head - - due to centrifugal force In a reaction I. F. or O. F. W t 2g W and therefore VJ- V? _/i In words, the change of en- ergy from atof = work due to pressure -|- work due to centrifugal force. In an impulse I. F. or O. F. V ^~ r '* = ***~" 1 *. (18) In a reaction A. F. 2g and therefore In words, the change of en- ergy from a to f = work due to pressure -f- work due to gravity. The work due to A centrifugal force is evidently nil. In an impulse A. F. ^ ~ V ^ = h - - ( I8 > 290 HYDRA ULICS. To make allowance for hydraulic resistances , F, 2 may be substituted for V 9 in equations 17 and 18, a mean value of a being i.i. For a maximum effect the water should leave the wheel without velocity, i.e., v t should be nil. But this value of v^ is impracticable, as no water could then pass through the wheel. It is usual either to make the velocity of whirl (v m ") at the outlet-surface equal to nil, or to make the relative (F 2 ) and circumferential (u 9 ) velocities at the outlet-surface, equal and opposite. In each case v 9 is small. First let -" = <>, d9) so that the water leaves the wheel with a much-reduced ve- locity in a direction normal to the out- let-surface. Thus (Fig. 194), &\*)** fy = 90; *.=?*/', and Aj(j ^ = Z> 2 COt /3 = V 9 COS ft. (2O) \v 2 -v' r V 2 ' / Also, by equations 2, 4, 5, and 20 FIG. 189. In an I. F. or O. F. ~= vi sinyridi = V* sin/J>v/ 2 211 (21) In an A. F. = v\ sin ydi = V* sin fid* = 2 tan fidi. (21) The following results are now easily obtained : In an I. F. or O. F. : Relation between the Vane- angles. By equations 9 and 21, and from the triangle acd, r\di sin y 3 r* u\ tan sin a. In an A. F. : Relation between the Vane- angles. By equations 9 and 21, and from the triangle acd, d\ sin y 2 l , x H l = cos y. . (25) Hence, by equations 20, 22, and 25, COt /J tan ft 4- 2 cot^ Note. If the water is to have no velocity of whirl (vj) relatively to the wheel at the inlet-surface, then i - v w ' = o, . . . (27) and therefore a = 90 and Vi Vr tan y = ,. Also, the efficiency and thus W Q\H 1 + h =-J= effective work wQ , wQ . . =v-wUi = UM cos y, (24) S and therefore ^i + h - r 1 = ^~ cos r- Hence, by equations 20, 22, and 25, 4- A) cot -.. (26) tan ft -\- i-j- cot . If the water is to have no velocity of whirl (v w f ) relatively to the wheel at the inlet-surface, then Ul - v w ' = o, . . (27) and therefore a = 90 and Also, the efficiency an thus (28) uS = g(Hi 4- . (28) 2 9 2 HYDRA ULICS. if the efficiency is perfect. Usually the efficiency of good turbines is about .85. Velocity of Efflux. Z'a 2 = z/a 5 tan 2 ft 2,07/1 tan ft (20) if the efficiency is perfect. Usually the efficiency of good turbines is about .85. Velocity of Efflux. z/ 2 2 = 2 2 tan 2 ft 2g(ffi 4- A) tan ft tan ft -j- 2 cot y Useful Work 2-^- cot y -wQfft li . ( 3 o) tan /?-(- 2-^- cot^ 2 cot ^ ^>/ TT \ r\ ** / V tan /3-}- 2-f- cot y Efficiency 2 ^ co. r -J- 2 ~r cot X Amount Q of water passing through turbine tan y#-f- 2-^ cot v i Amount Q of water passing through turbine 1 zgVi tan ft . , /ig(Hi -\- h} tan /5 - 27rr 2 y tan ft-{-2co\.y The pressure-head at the in- let-surface 2g ,.) r a V, Hl< I ~ OV 9 tan^ 2g HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 293 When the turbine is work- ing freely in space above the surface of the tail-water, there will be no inflow of air if p^ > A f - e -> if I > o , tan ft '* sin 2 ;r(tan/?-]-2- 2 -cot;K) d\ If the turbine is drowned with a head h' of water over the outlet-surface, there will be no back-flow of water if that is, if 7i ^ o tan ft When tne turbine is work- ing freely in space above the surface of the tail-water, there will be no inflow of air if p l > At Le -' if ffi -- h h , w w that is, if tan ft IN IMPULSE TURBINES. In an I. F. or O. F.: Speed of Turbine. Since V? = 2gff 1 , . . (35) by equation 22, riVi 8 sin 2 y _ uj _ rj u^ and therefore Velocity of Efflux. = 2 tan p ~nW tan ft -f- 2-^ cot y d\ In an A. F. : Speed of Turbine. Since , - (35) by equation 22, dS sin 2 y _ uf_ _ uf_ d- (39) Work -/r,g-.in'r). (38; Efficiency \ -ij- sin a y = n. (39) Second, let so that the water again leaves the wheel with a much-reduced velocity. Evidently also J - a z= 2& 2 cos = 22/ a sn 2 sn . 2 . (42) Also, by eqs. 2, 4, 5, and 42 In an I. F. or O. F. Q_ zit = a sin/? r a / 2 , i.e., if When the turbine is drowned, with a head h' of water over the outlet-surface, Velocity of Efflux. By equations 42 and 47 ft sin 2 - (50) Useful Work = Q(ffi + h){ i - ~ tan ^ tan y \ (51) Efficiency =I * 2 i<&) =I -| tan ^ an7 '- (52) Amount Q of Water passing through Turbine = inRdiVr" = 2itRd/, i.e., if Hi d sin ft H\ -\-k d\ sin 2y' When the turbine is drowned, with a head h' of water over the outlet-surface, HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. there will be no back-flow of water if * ^ ^ _u h' > r " IV IV ' that is, if Ti W_ r-^d-j sin ft r^di Sin 2*y' HI there will be no back-flow of water if A . A , ,, > -- h h , ' that is, if H, - h 1 d* sin ft Hi -j- A' di sin 2y IN IMPULSE TURBINES. In an I. F. or O. F. : Speed of Turbine. Since t . . (55) Velocity of Efflux. jn U-? = 4 2 2 sin 2 -- 2. *-$# ft - . cos 2 2 Efficiency . (59) In an A. F. : Speed of Turbine. Since (55) ^ ^. . . (56) Velocity of Efflux. . (57) Useful Work -^ cog2 r Ms8) ( 3S 2 ) Efficiency H\ d^ sin 2 v = I ~ LJ- i /. T^ *' (59) 298 H YDRA UL ICS. The great advantages possessed by turbines over vertical wheels on horizontal axes are shown by a consideration of the expressions for the useful work and efficiency. The former involves the available head only, while the latter is independent even of that. Thus a turbine will work equally well under water or above water, while its efficiency remains the same, whatever the available head may be. The efficiency, also, increases as the ratio diminishes. a, The value of d l , however, must not be too small, as there might be a loss of energy due to a contracted section at entrance, while if d z is made too large, the vane-passages will no longer run full bore. Finally, the efficiency -increases as the angles /? and y diminish. In practice y usually ranges from 10 to 30 in an I. F., and from 20 to 50 in an O. F. and A. F., an average value being 20 for an I. F., and 25 for an O. F. and A. F. In an I. F. ft generally ranges from 135 to 150 if ?/ 2 F 2 , or from 30 to 45 if vj' o, and in an O. F. and P. F. from 20 to 30, an average value being 145 or 35 for an I. F., accord- ing as # 2 = F 2 , or vj r = o, and 25 for an O. F. and A. F. 25. Remarks on the Centrifugal Head From equations 14 and 17 In an I. F. w a < u, , and the term L is negative. Hence the velocity v l diminishes as the speed of the tur- bine increases and vice versa. The centrifugal head - J - therefore tends to secure a steady motion in the case of an I. F., and also to diminish the frictional loss of head. For this rea- son it should be made as large as possible consistent with practical requirements, and is usually made equal to 2. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 299 In an O. F., on the other hand, u^ > u l and the centrifugal head is positive. The velocity v l will, therefore increase and diminish with the speed of the turbine (). Thus the cen- trifugal head is adverse to a steady motion, and tends both to augment a variation from the normal speed and to increase u * _ u a the frictional loss of head. It follows that - should be tg as small as possible consistent with practical requirements, and a common value of is 1.25. ^i Again, eq. 5 shows that the velocity of flow v r (and there- fore also ,) increases as the size of the wheel diminishes, and is accompanied by a corresponding increase in the frictional loss of head. Hence it would seem advisable to employ large wheels ; but if the size of a wheel is increased, it must be borne in mind that the skin-friction (if the turbine works under water), the weight, and consequently the journal friction, will all increase. Belanger has suggested that the efficiency of an A. F. may be increased by so forming the vane-passages that the path of a fluid particle gradually approaches the axis of rotation. 26. Practical Values of the Velocities, etc. Let v be the theoretical velocity due to the head H\ i.e., let v* = 2gH. Experience indicates that the following values will give good results in reaction turbines : Inl.R, Vr ' = Vr " = ; In O. F., v r ' = - ; v r " = .2iv to .172; ; u, = -u^ = .$6v. 4 r i In A. F., v r r = v r " = .i$v to .2v ; u, = u 9 = -v to -v. Again, in reaction and impulse turbines the thickness of. the vanes varies from -J inch to f inch if of wrought iron, and 3OO HYDRAULICS. from \ inch to f inch if of cast iron. In the latter case the vanes are usually tapered at the ends. In axial-flow turbines the mean radius R is often made to vary o . _ . _ from - yA J sin y to 2 yA t sin y if A^ sin y < 2 square feet ; from --'\fA 1 sin y to -\A4,sin y\i A 1 s\ny > 2sq. ft.< l6sq. ft.; 4 2 from \/ ' A l sin ;/ to ^\/A 1 sin ^ if ^4, sin y > 16 square feet. 4 In axial-impulse turbines the mean radius R is often made to vary from --v/^sin ;/ to 2<\fA 1 s'my. 4 Also, the depth h of the wheel varies from - r to - - but o II must be determined by experience. Again, For a delivery of 30 to 60 cubic feet and a fall of 25 ft. to 40 ft. y should be 15 to 18, and (3 should be 13 to 16. For a delivery of 40 to 200 cubic feet, and a fall of 5 ft. to 30 ft. y should be 1 8 to 24, and fi should be 16 to 24. For a delivery of more than 200 cubic feet, and lower falls, y should be 24 to 30, and 24 to 28. In axial-impulse turbines it may also be assumed as a first approximation that . ?A vju. work per pound = - = _^L_J 2T g and therefore V l = 2#, cos y = 2 Vi cos y. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 30 1 27. Theory of the Suction (or Draught) Tube. Vortex and axial-flow turbines sometimes have their outlet orifices opening into a suction (or draft) tube which extends down- wards and discharges below the surface of the tail-water. By such an arrangement the turbine can be placed at any conven- ient height above the tail-water and thus becomes easily acces- sible, while at the same time a shorter length of shafting will suffice. The suction tube is usually cylindrical and of constant diameter, so that there is an abrupt change of section at the outlet surface of the turbine, producing a corresponding loss of energy by eddies, etc. This loss may be prevented by so form- ing the tube at the upper end that there is no abrupt change of section, and by gradually increasing the diameter downwards. The cost of construction is greater, but the action of the tube is much improved. Let h' be the head above the inlet orifices of the wheel. Let h" be the head between the inlet orifices and the sur- face of the tail-water. Let L l be the loss of head up to the inlet surface. Let L^ be the loss of head between the wheel and the tube outlet. Let v^ be the velocity of discharge from the outlet at bottom of tube. Let P be the atmospheric pressure. Then, assuming that there is no sudden change of section at the outlet surface, h ' ~~ = L ' and therefore w 2g v* - 2 gi K + J** 302 HYDRA ULICS. where H = h' + h" = total head above tail-water surface ; and -^ a a ,_^ 4 2 , Z-j-, Z a are expressed in the forms 2 l ' 4 1 ' *2g' *2g* * 3 > /*4> A* 6 A<6 being empirical coefficients. Again, the effective head and is entirely independent of the position of the turbine in the tube. Also, if A i is the area of the outlet from the suction-tube, A^VI = Q = A l v l sin y, so that v. can be expressed in terms of z/ 4 , and hence ** 1 ~ ^ is w also independent of the position of the turbine in the tube. Suppose the velocity of flow to be so small that ^ 4 , v L 9 may be each taken equal to nil. Then W and since the minimum value of /, is also nil, the maximum theoretical height of the wheel above the tail-water surface is equal to the head due to one atmosphere. Again, V 3 = v l cos yu l u^u, F, cos ft) + L - But A l v l sin y = Q = A^ sin d = A^ sin ft = Apt ; HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 303 and hence, taking gH = z/,(i cos Y + **** u * cos ft) " + ^-^ and therefore - w a _j_ ? 7 , i cos r + ^ - , cos /? = v? + 2v ^ . cos y + i/^, . cos ( cos ^ + V/* 8 ( \ ^ where B cos V* 8 cos Hence it follows that z/, increases with a , i.e., with the speed of the turbine, if A suction-tube is not used with an outward-flow turbine, but a similar result is obtained by adding a surrounding sta- tionary casing with bell-mouth outlet. A similar diffusor might be added with effect to a Jonval working without a suction-tube below the tail-water. The theory of the diffusor is similar to that of the suction-tube. 28. Losses and Mechanical Effect. The losses may be enumerated as follows: I. The loss (Z,) of head in the channel by which the water is taken to the turbine. L -/-^ *' " 7l m 2g> fi being the coefficient of friction with an average value of 304 HYDRA ULICS. .0067, / the length of the channel of approach tn its mean hydraulic depth, and v the mean velocity in the channel. L l is generally inappreciable in the case of turbines of the inward- and axial-flow types, as they are usually supplied with water from a large reservoir in which V Q is sensibly nil. If A Q is the sectional area of the supply-channel, then A v = Q = A 1 v 1 sin y y and , = /, - A, II. The loss (Z a ) of head in the guide-passages. This loss is made up of : (a) The loss due to resistance at the entrance into the passages ; (b) The loss due to the friction between the fluid and the fixed blades; (c) The loss due to the curvature of the blades ; (d) The loss of head on leaving the guide-passages. These four losses may be included in the expression / a being a coefficient which has been found to vary from .025 to .2 and upwards. An average value of f 9 is .125, but this is somewhat high for good turbines. Note. In Impulse turbines / a has been found to vary from .11 to .17. III. The loss (Z, 3 ) due to shock at entrance into the wheel. In order that there may be no shock at entrance, the relative velocity ( F,) must be tangential to the lip of the vane. For any other velocity (z// = ac'} and direc- tion (dad = y f ) of the water at en- trance, evidently L 3 = the loss of head FIG. 191. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 3O$ (v' sin y' ^ sin y) 3 (v' cos y' v l cos y) 9 _ (v f sin ;/ V l sin <*) a (z/ cos y' z\ V l cos a) 8 Generally a? is small, and L 3 is always nil when the turbine is working at full pressure and at the normal speed. This loss of head in shock caused by abrupt changes of sec- tion, and also at an angle, may be avoided by causing the sec- tion to vary gradually, and by substituting a continuous curve for the angle. IV. The loss (Z, 4 ) of head due to friction, etc., in passing through the wheel-passages, including the loss due to leakage in the space between the guides and the inlet-surface. This loss is expressed in the form V: sn ftl where f^ varies from .10 to .20. Note. The loss of head due to skin-friction often governs the dimensions of a turbine, and renders it advisable, in the case of high falls, to employ small high-speed turbines. V. The loss of head (L b ) due to the abrupt change of sec- tion between the outlet-surface and the suction-tube. As in III, v 9 (=ffy is suddenly changed into v t ' (= fh'\ and loss of head is 2g 2g 2g since h ' x is very small and may be disre- garded. Thus, ( FiG. 192. 4 = #/ being the component of vj (fh f ) in the direction of the axis of the suction-tube. 3O6 HYDRA ULICS. If there is no abrupt change of section between the outlet- surface and the tube, Z & is nil. VI. The loss of head (L 6 ) due to friction the in suction-tube. Assume that the velocity v^ of flow in the tube is equal to v^ the velocity with which the water leaves the turbine. Also let A be the sectional area of the tube. Then / f - f 6 ~~ /6 m' 2g ~ /6 m' \ A, I 2g ' / 6 ( =/ t ) being the coefficient of friction with an average value of .0067, I' the length of the tube, and m' its mean hydraulic depth. VII. The loss (Z 7 ) of head due to entrance to sluice at base of tube. This loss may be expressed in the form A the average value of/ 7 being about .03. VIII. The loss (Z 8 ) of head due to the energy carried away by the water on leaving the suction-tube. and z> 4 usually varies from | V2gH to f V2gH. In good turbines the loss should not exceed 6#. It might be reduced to 3$, or even to i$, but this would largely increase the skin-friction. IX. The loss of head (L 9 ) produced by the friction of the bearings. being the coefficient of journal friction, Wthe weight of the HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. turbine and of the water it contains, and p the radius of the journal. Hence the total loss of head and the total mechanical effect Note. If there is no suction-tube, 6 = O = L 6 = L, = and the total loss becomes fall from outlet-surface tail-water surface. to 29. Centrifugal Pumps. If an hydraulic motor is driven in the reverse direction, and supplied with water at the point from which the water originally proceeded, the motor becomes a pump. All turbines are reversible, and may, therefore, be converted into pumps, but no pump has yet been constructed of an inward-flow type. The ordinary centrifugal pump, Fig. 193, is an outward-flow machine. It is more economical and less costly for low falls than a recip- rocating pump, and has been known to give good and eco- nomic results for falls as great as 40 feet. With compound centrifugal pumps very much greater lifts are economically possible. There are three main differ- ences between centrifugal pumps and turbines: ist. The gross lift with a pump is greater, on account FIG. 193. 308 HYDRAULICS. of frictional resistances, than the fall in the case of a tur- bine. 2d. The water enters the pump-fan without any velocity of whirl (vj o) and leaves the fan with a velocity of whirl (v w ") which should be reduced to a minimum in the act of lifting, but which is by no means small. In a turbine, on the other hand, the water has a considerable velocity of whirl (v w '} at entrance, while at exit the velocity of whirl (v w ") is reduced to a minimum, and is generally nil. 3d. In a turbine the direction of the water as it flows into the wheel is controlled by guide-blades ; whereas in the case of a pump, the direction of the water, as it flows towards the discharge-pipe, is controlled by a single guide-blade, which forms the outer surface of the volute, or chamber, into which the water flows on leaving the fan. FIG. 194. Experimental Centrifugal Pump in the Hydraulic Laboratory, McGill University. Before the pump can be put into action it must be filled, and this can be effected through an opening (closed by a plug) in the casing when the pump is under water, or, if the pump is above water, by creating a vacuum in the pump-case by means of an air pump or a steam-jet pump, when the water must necessarily rise in the suction-tube. At first the water rotates as a solid mass, and delivery com- mences when the speed is such that the head due to centrifugal force r u *\ exceeds the lift. This speed may be after- \ 2g I wards reduced, providing a portion of the energy is utilized at exit. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 309 As soon as the pump, which is keyed on to a shaft driven by a belt or by gearing, commences to work, the water rises in the suction-tube and divides so as to enter the eye of the pump-disc on both sides. As in turbines, the revolving pump- disc is provided with vanes curved so as to receive the water at the inlet-surface, for a given normal condition of working, without shock. Experiment has also tended to show that the angle between the tangents to a vane and the disc circumfer- ence at the outlet-surface, may be advantageously made as small even as 15, but manufacturers hold different opinions on this point. The water leaves the disc with a more or less con- siderable velocity, and impinges upon the fluid mass flowing round the volute, or spiral casing surrounding the disc, towards the discharge-pipe. This volute should have a section gradu- ally increasing to the point of discharge, in order that the delivery across any transverse section of the volute may be uniform. This volute is also so designed as to compel rotation in one direction only, with a velocity corresponding to the velocity of whirl (v w ff ) on leaving the fan. There are exam- ples of pumps in which the delivery is effected in all direc- tions, and the water is guided to the outlet by a number of spiral blades. In these pumps an important advantage is gained by the addition of a vortex or whirlpool chamber surrounding the pump-disc. The water discharged from the disc then contin- ues to rotate in this chamber, and a portion of the kinetic energy is thus converted into pressure energy, which would otherwise be largely wasted in eddies in the volute or discharge- pipe. The water leaves the vortex chamber with a diminished whirling velocity which cannot be very different in direction and magnitude from the velocity of the mass of water in the volute. The vortex chamber is provided with guide-blades following the direction of free vortex stream-lines (equiangular spirals) so as to prevent irregular motion. A conical suction- pipe is advantageous, as it allows of a gradual increase of velocity, and a still greater advantage is to be found in the use of a conical discharge-pipe. The velocity in the dis- charge-pipe should not be too great, as it leads to a waste of 310 HYDRAULICS. energy. A velocity of 3 to 6 feet is found to give the best results. Pumps work under different conditions from turbines, and hence there are corresponding differences necessary in their design. They work best for the particular lift for which they are designed, and any variation from this lift causes a rapid reduction in the efficiency. 30. Theory of Centrifugal Pump. Denote the velocities at the inlet- and out- let-surfaces of the pump-fan by the same symbols as in turbines. Let Q be the delivery of the pump. Let H s be the gross lift, including the actual lift (ff a ), the head due to the velocity FIG. 195. O f delivery, the heads due to the frictional resistances in the ascending main, in the suction-pipe and in the wheel-passages, and the head corresponding to the losses " in shock " at entrance and exit. Let H a be the actual lift. The total work done on the wheel The useful work done by the pump Hence the efficiency (rf) = g g At the inlet-surface the flow is usually radial, so that y = 90, and the velocity of whirl vj is nil. Thus, the efficiency = fr * = 77, and the equation HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 311 is the fundamental equation governing the design of centrifu- gal pumps. Again, H : the efficiency n = jh- = i " For a given speeed ( 3 ) this is a maximum when j (VJJ+(U,-V "y y v y -77 = a minimum -- v w w w Hence, differentiating, - u* tan 2 ft - -<^- + sec- /> = Q, and therefore vj f = u^ sin ft, and is the velocity of whirl at exit which, for a given speed (w a ), will give a maximum efficiency. Note.li u^ = v w ", then and the water leaves the fan with a velocity equal to that due to at least one half of the gross lift. The efficiency must therefore be necessarily less than .5. Again, since v r " cot ft = u^ v w ", ft must be 90 if u^ = vj'-, but ft is generally much less than 90, and therefore v w " is generally less than u y Let v w " = ku^> k being an empirical coefficient less than unity. Then kit? gH e and the efficiency = -~> KU^ Consider two cases. CASE I. Pump without a vortex-chamber. When the water is discharged into the volute, the velocity of flow (v r ") is wasted and the velocity of whirl (v w ' f ) is sud- denly changed to the velocity v s of the mass of water in the 312 HYDRAULICS. volute assumed to be moving in a direction tangential to the pump-disc. Thus, (yjy - far (vj f - vy the gam of pressure-head = - i fe/') 9 ^ " which is a maximum and equal to when v s = -^ . 4 g This gain of head is always very small and may be dis- regarded as being almost inappreciable. Neglecting also the losses due to frictional resistances, etc., then, precisely as in the case of turbines, v_ , TT __ f variation of pressure-head between 2g ( outlet and inlet surfaces. *.*-. FV-F? But V? = u? + T^ 2 , since y = 90, and therefore _ u * ~ ~_JL. _ u * ~ ( u i ~ v '}* sec2 ft and u 2 (u v // ) 2 sec 2 ft the efficiency - w .. 2^ w " which is a maximum for a given speed & a and equal to ; j- : -5 when v w ff = u^ sin /?. Thus the efficiency increases as ft diminishes. When ft = 90, or ^ w r/ & 2 , the maximum efficiency is , and therefore one half of the work done in driving the pump is wasted. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 313 Note. Loss of head = loss due to hydraulic friction -f- loss due to abrupt change from v w " to v, -\- loss due to dissipation of v r " -f- loss due to v s carried away = loss due to friction (hydraulic) + *,(*" - ^ (*l_ _ (VT 2g 2g 2g = loss due to friction (hydraulic) , (*.") , ~~ when v s = \vj'. CASE II. Pump ivith a vortex-chamber (Fig. 199). The diameter ( 2r 3 ) of the outer surface of this chamber should be at least twice that of the outlet-surface of the pump- disc. Assuming that the motion in the chamber is a free vortex, then the gain of ) _ v^_ I r?\ pressure-head ) 2g \ r 3 2 / and hence the efficiency = T,, . . FIG. 106. This, again, is a maximum for a given speed, when vj = u^ sin fi y its value being I +'(l - Sj) sin ft I + sin ft 3 1 4 HYDRA ULICS. This expression increases as ft diminishes, but the value of ft is not of so much importance as in Case I, and it is very common to make ft equal to 30 or 40. When ft = 90 the maximum efficiency = - ( 2 - -M = if r a = 2r,. 31. Practical Values. The following values are often adopted : 3 = d^ when faces of pump-disc are parallel ; ^ = \d^ when pump-disk is coned. HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. 31$ EXAMPLES. 1. An accumulator ram is 9 inches diameter and 21 feet stroke. Find the store of energy in foot-pounds when the ram is at the top of its stroke, and is loaded till the pressure is 750 Ibs. per square inch. Ans. 958,000 ft.-lbs. 2. In a differential accumulator the diameters of the spindle are 7 inches and 5 inches ; the stroke is 10 feet. Find the store of energy when full and loaded to 2000 Ibs. per square inch. Ans. 377,000 ft.-lbs. 3. A direct-acting lift has a ram 9 inches diameter, and works under a constant head of 73 feet, of which 13 per cent, is required by ram-fric- tion and friction of mechanism. The supply-pipe is 100 feet long and 4 inches diameter. Find the speed of steady motion when raising a load of 1350 Ibs., and also the load it would raise at double that speed. If a valve in the supply-pipe is partially closed so as to increase the coefficient of resistance by 5!, what would the speed be ? Ans. Speed = 2 ft. per second ; load = 150 Ibs. 4. Eight cwt. of ore is to be raised from a mine at the rate of 900 feet per minute by a water-pressure engine, which has four single-acting cylinders, 6 inches diameter, 18 inches stroke, making 60 revolutions per minute. Find the diameter of a supply-pipe 230 feet long for a head of 230 feet, not including friction of mechanism. Ans. Diameter = 4 inches. 5. If A. be the length equivalent to the inertia of a water-pressure engine, F the coefficient of hydraulic resistance, both reduced to the ram, -z/o the speed of steady motion, find the velocity of ram after moving from rest through a space x against a constant useful resistance. Also find the time occupied. Ans. v* = V 6. An hydraulic motor is driven from an accumulator, the pressure in which is 750 Ibs. per square inch, by means of a supply-pipe 900 feet long, 4 inches diameter; what would be the maximum power theoreti- cally attainable, and what would be the velocity in the pipe correspond- ing to that power? Find approximately the efficiency of transmission at half power. Ans. H.P. = 240 ; v = 22 ft. ; efficiency = .96 nearly. 7. A gun recoils with a maximum velocity of 10 feet per second. The area of the orifices in the compressor, after allowing for contraction, may be taken as one twentieth the area of the piston. Find the initial pressure in the compressor in feet of liquid. HYDRAULICS. Assuming the weight of the gun to be 12 tons, friction of sUde 3 tons, diameter of compressor 6 inches, fluid in compressor, water, find the recoil. Find the mean resistance to recoil. Compare the maximum and mean resistances, each exclusive of friction of slide. Ans. 621; 4ft. 2^ in. ; total mean resistance = 4.4 tons; ratio = 2.5. 8. A reaction wheel is inverted and worked as a pump. Find the speed of maximum efficiency and the maximum efficiency, the coeffi- cient of hydraulic resistance referred to the orifices being .125. Ans. Speed = twice that due to lift ; .758. 9. A reaction wheel with orifices 2 in. in diameter makes 80 revolu- tions per minute under a head of 5 ft. The distance between the centre of an orifice and the axis of rotation is 12 inches. Find the H.P. and the efficiency. Ans. .146; .596. 10. In a reaction wheel the speed of maximum efficiency is that due to the head. In what ratio must the resistance be diminished to work at | this speed, and what will then be the efficiency? Obtain similar results when the speed is diminished to three fourths its original amount. Ans. .949; .8896; 1.071; .753. 11. In a reaction wheel, determine the per cent of available effect lost, (i) if i? = 2gH\ (2) if tt* = ^gH; (3) if u 1 = ZgH. What conclusion may be drawn from the results? Efficiencies are respectively .828, .9, .945. 12. An undershot water-wheel with straight floats works in a straight rectangular channel of the same width as the wheel, viz., 4 ft.; the stream delivers 28 cub. ft. of water per second, and the efficiency is . Find the relation between the up-stream and down-stream velocities. If the velocity of the inflowing water is 2 ft. per second, find the velocity on the down-stream side and determine the mechanical effect of the wheel, its diameter being 20 ft., the diameter of the gudgeons being 4 in., and the coefficient of friction .008. 13. A vane rotates about an axis with an angular velocity A, and and water moves freely along the vane. Show that the work per unit of weight of water, due to centrifugal force, in moving from a point distant A1(i = 17 ft. ; u = n ft. per second; elbow- angle = 70; division-angle = 5; water enters the first bucket at 12 from summit of wheel. Find (a) the relative velocity Fso that water may enter unimpeded; () the direction of the entering water; (c) the diameter of the wheel, which makes 5 revolutions per minute ; (d) the position and direction of the sluice, which is 2 ft., measured hori- zontally from the point of entrance. 40. In an overshot wheel the deviation of the impinging water from the direction of motion of the wheel is 10 ; the velocity (vi) of the im- pinging stream = 15 ft. per second; of the circumference of the wheel () = 15 cos 10. What proportion of the head is sacrificed? 41. A 3o-ft. water-wheel with 72 buckets and a 12-in. shrouding makes 5 revolutions and receives 240 cub. ft. of water per minute. Find the width and sectional area of a bucket. The fall is 30 ft. ; at what point does the water enter the wheel, the inflowing velocity being i| times that of the wheel's periphery? Also find the deviation of the water- surface from the horizontal at the point at which discharging com- mences, i.e., 140 from the summit. 42. What number of buckets should be given to an overshot wheel of 3 2O H YDRA UL ICS. 40 ft. diameter and 12 in. width in wheel, pitch-angle = 4, thickness of bucket lip = i in., water area = 24^ sq. in. ? 43. A wheel makes 5 revolutions per minute, the radius is 16 ft., and the discharging angle 50. Find deviation of water-surface from the horizon. Ans. 4 .29. 44. A wheel makes 20 revolutions per minute; radius = 5 ft., angle of discharge = o. Find deviation of water-surface from horizon. Also find deviation at 44 35' above centre. Ans. 4 33' ; 44 34'. 45. The water in a head-race stands 4.66 ft. above the sole and leaves the race under a gate which is raised 6 in. above the sole, the coefficient of velocity (v*) being .95. The water enters a breast wheel in a direction making an angle of 30 with the tangent to the wheel's periphery at the point of entrance. The speed (u) of the periphery is 10 ft. per second, the breadth of the wheel is 5 ft., the depth of the water beneath the axle is 8 in., and the length of the flume is 8.2 ft. Find the loss of head (a) due to the destruction of the relative velocity (V) at entrance; (b) due to the velocity of flow in the tail-race ; (c) in the circular flume. Ans. (a) i.u ft.; () 1.57 ft. ; (c] .44 ft. 46. In the preceding example, find how the losses of head would be modified if the flume were lowered 1.03 ft., and if the point of entrance were raised so as to make u = v\ cos 30. 47. A water-wheel has an internal diameter of 4 ft. and an external diameter of 8 ft.; the direction of the entering water makes an angle of 15 with the tangent to the circumference. Find the angle subtended at the centre of the wheel by the bucket, which is in the form of a cir- cular arc, and also find the radius of the bucket. 48. An overshot wheel 5 ft. wide, 30 ft. in diameter, having a 12-in. crown and 72 buckets, receives 10 cub. ft. of water per second and makes 5 revolutions per minute. Determine the deviation from the horizontal at which the water begins to spill, and also the corresponding depression of the water-surface. 49. An overshot wheel makes revolutions per minute ; its mean iTt diameter is 32 ft. ; the water enters the buckets with a velocity of 8 ft. per second at a point 12 30' from the summit of the wheel. At the point of entrance the path of the inflowing water makes an angle of 30 with the horizontal. Show that the path is horizontal vertically above the centre. The sluice-board is placed at a point whose horizontal distance from the centre is one half that of the point of entrance. Find its position relatively to the centre and its inclination to the hori- zon. (Sin 12 30' = .2165). 50. The water enters the buckets of the wheel in the preceding example without shock. Find the elbow-angle. Also, if the buckets HYDRAULIC MOTORS AND CENTRIFUGAL PUMPS. $21 begin to spill at 150 from the summit, find where the bucket is empty and the number of buckets. (Depth of crown = 12 in.; thickness of bucket = \\ in.) 51. Given 7/1 = 15 ft. per second, and S 2oJ. Find the position of the centre of the sluice, which is 4 in. above the point of entrance. Ans. .097 ft. vertically below and 1.114 ft- horizontally from the summit. The axis of the sluice is inclined at 9 58' to the horizontal. 52. In an overshot water-wheel 2/1 = 15 ft.; u = 10 ft. ; elbow-angle = 70^ ; division-angle = 4^ ; deviation from summit of point of en- trance = 12. Find the deviation of the layer from that of the arm, so that the water might enter unimpeded; also find the inclination of the layer to the horizon, and the value of V . If the centre of the sluice- aperture is to be 4 in. above point of entrance, find its vertical and horizontal distance Trom the vertex of the stream's parabolic path which is vertically above the centre of the wheel, and also find inclination oi sluice-board to horizon. Ans. 15! ; 2oJ ; 5.3 ft. per sec. ; .42 ft. ; 1.04 ft. ; 9 34', 53. In an overshot wheel Q= 18 cub. ft.; r\ =6 ft. ; d i ft. ; b = 4 ft. ; N 24. At the moment spilling commences the area afd = 1.025 sq. ft.; between this point and the point where the spilling is com- pleted three buckets are interposed, the sectional areas of the water being .591, .409, and .195 sq. ft., respectively. Find (a) the sectional area of bucket, (b) the point where the spilling commences, (c] the point where the spilling is completed, (d) the height of the arc of discharge, ( 255 head in turbine, 298 pumps, 307 " theory of, 309 " vortex-chamber in, 309, 313 Chamber, whirlpool, 50 Channel, bottom velocity of flow in a, 154 flow in an open, 131 form of, 135, 136 " maximum velocity of flow in a, 150, 153 mean velocity of flow in a, 151, 154 mid-depth velocity of flow in a, 1ST " steady flow in a, 132 " surface velocity of flow in a, 150, 15-1 value of yin a, 144 variation of velocity in a sec- tion of a, 148 Channels, differential equation of flow in, 159 examples of, 162 " of constant section, steady flow in, 132 " of varying section, flow in, 156 surface slope in, 160, 161 Chezy's formula, 88 Cock in cylindrical pipe, 93 Cocks, loss of head due to, 93 Coefficient of contraction, 22, 89 " discharge, 24 " friction, 73, 144 " " resistance, 21 333 334 INDEX. Coefficient of velocity, 20 Combined-flow turbines, 284 Compressibility, 2 Continuity, 2, 5 Contraction, imperfect, 22 incomplete, 23 loss of head due to ab- rupt, 89 Coulomb, 72 Critical velocity, 97 Cunningham, 148 Current-meters, 180 Darcy, 72, 74, 75, 97, 148, 154 gauge, 176, 178 D'Aubuisson, 74 Density, 2 Downward-flow turbine, 282 Draught-tube, theory of, 301 Dubuat, 154 Elasticity of volume, 3, 4 Elbows, loss of head due to, 91 Ellis, 106 Energy lost in shock, 32 " of fall of water, 4 " jet of water, 27 " transmission of, 84 Enlargement of section, loss of head due to, 32, 91 Equations, general, 30 Equivalent uniform main, TOO Erosion caused by watercourses, 136 Examples, 60, 122, 170, 209, 315 Exner, 183 Eytelwein, 74, 134 Floats, sub-surface, 175 " surface, 175 " twin, 175 Flow from vessel in motion, 16 in a frictionless pipe, 18 " in aqueducts, 142 " influence of pipe's inclination and position upon the, 83 " in pipes, 78 " in pipe of uniform section, 86 " " " of varying diameter, 98. Fluid friction, 70 " motion, I Fourneyron's turbine, 281 Francis, 176 Friction, coefficients of, 70, 73, 74, 75 in pipes, surface, 73, 97 laws of fluid, 72 Froude, u, 13, 70, 76, 97 Froude's table of frictional resistances, 70 Ganguillet, 147 Gauge, Darcy, 176, 178 Gauges, experiments on, 148 Gauging, method of, 173 Gaugings on the Ganges, 148 " " Mississippi, 146 General equations, 30 Gerstner's formula, 229 Graphical representation of losses of head, 94 Grassi, 3 Head, 2, 3 Herschel, 184 Hook gauge, Boy den's, 173 Humphreys, 148, 151 Hurdy-gurdy, 279 Hydraulic gradient, 10 mean depth, 133 " radius, 80 " resistances, 20 Hydraulics, definition of, i Hydrodynamometer, Perrodil's, 183 Hydrometric pendulum, 183 Impact, 1 86 on a curved vane, 199 on a surface of revolution, 192 on a vane with borders, 195 Inclination, influence of pipe's, 83 Injector, 12 Inward-flow turbine, 282 Jackson, 148 Jet, energy of, 27 inversion ot, 27 momentum of, 27 propeller, 191 reaction wheel, 272 efficiency of, 274 " useful effect of, 274 Kutter, 147 Laminar motion, 2 Lesbros, 27 Limit turbine, 283 Loss of energy in shock, 32 Loss of head due to abrupt change of section, 89 " " " " bends, 92 " *' " " cocks, 93 " " " " contraction of sec- tion, 89 " " elbows, 91 " " " enlargement of section, 91 IAD EX. 335 Loss of head due to orifice in dia- phragm, 90 " " " " " sluices, 93 " " *" valves, 93 Losses of head, graphical representa- tion of, 93 Magnus, 27 Main of uniform diameter, branch, 101 " with several branches, 118 Meters, 180 " inferential, 184 " piston, 184 " rotary, 184 Meyer, 156 Miner's inch, 26 Mississippi, experiments on, 148 Mixed-flow turbines, 284 Motion, fluid, I " in plane layers, 2 " in stream-lines, 2 " laminar, 2 " permanent, I " steady, I Motor driven by water flowing along a pipe, 107 Mouthpiece, Borda's, 34 convergent, 44 cylindrical, 39 divergent, 42 ring-nozzle, 37 Navier, 149 Notch, 54 " circular, 55 " rectangular, 54 triangular, 56 Nozzles, 104 " Ellis's experiments on, 106 Open channels, 131 Orifice fed by two reservoirs, 115 " flow through an, 16 :-;*' in a diaphragm, loss of head due to, 98 in a thin plate, 13 " in vertical plane surfaces, 50 with a sharp edge, 14 Orifices, circular, 53 large, 50 " rectangular, 50 Outward-flow turbine, 281 Overshot wheel, 225, 254 " arc of discharge in, 256 bucket angle of, 262 " " division angle in, 262 Overshot wheel, effect of centrifugal force in, 255 " effect of impact on, 270 " " weight on, 268 " number of buckets in, 262, 264 " pitch-angle in, 262 " " speed of. 254 " useful effect of, 268, 271 weight of water on, 256 Parabolic path of jet, 16 Pelton wheel, 280 Permanent regime, i Piezometer, 9 Pipe connecting three reservoirs, branched, in two reservoirs, 86 " of rectangular section, sluice in, 93 " uniform section, flow in, 78 " " varying section, 18, 98 Pitch-back wheel, 272 Pilot tube, 176 Plane layers, motion in, 2 Poiseuille, 96, 97 Poncelet, 27, 227 Poncelet's wheel, 232 Position, influence of pipe's, 83 Pressure-head, 4 Prony, 74, 134 Pumps, centrifugal, 307 " " theory of, 309 '* vortex - chamber in, 309, 313 Radiating current, 46 Rayleigh, Lord, 27 Reaction, 190 Reaction wheel, efficiency of, 274 " " jet, 272 Regime, permanent, I Reservoirs, Branched pipe connecting three, in orifice fed by two, 115 " pipe connecting two, 86 Resistance of ships, 76 " to flow, law of, 96 Revy's meter, 181 Reynolds, 97 Ring-nozzle, 37 River-bends, 143 Sagebien wheels, 254 336 INDEX. Schiele turbine, 284 Ships, resistance of, 76 Siphon, 108 " inverted, 109 Slotte, 156 Sluice in cylindrical pipe, 93 " in rectangular pipe, 93 loss of head due to a, 93 Sluices, 244 Standing wave, 165, 232 Steady flow in channels of constant section, 132 Steady motion, i, 132 " in pipe of uniform section, 78 Stream-line, 2 Suction-tube, theory of, 301 Surface-floats, 175 Surface-friction in pipes, 73 ' slope in channels, 160, 161 Table of bottom velocities, 155 " Castel's results, 45 " " coefficients of discharge, 24, 25, 45 " friction, 73, 75 " " velocity, 23 " " density of water, 3 " " discharge through nozzles, 1 06 " " elasticity of volume of water, 4 " " Ellis's experiments on nozzles, 106 " " frictional resistances, 70, 73, .74, 75 " maximum velocities, 155 " " values of f, 147 " " " " " Bazin's, 166 " '' viscosity of water and mer- cury, 155 Table of Weisbach's values of C v , 33 Theory of suction or draught tube, 301 " " turbines, 284 Thomson, James, 50, 143 Thomson's turbine, 282 Throttle valve, loss of head due to, 83 Time of emptying and filling a canal lock, 29 Torricelli's theorem, 14 Torricelli's theorem applied to the flow through a frictionless pipe of gradually changing section, 18 Transmission of energy by hydraulic pressure, 84 Turbine, axial-flow, 282 " centrifugal head in, 297 " combined, 284 Turbine, efficiency of, 288, 291, 292, 296, 297 Fontaine's, 282 impulse or Girard, 276 inward-flow, 282 Jouval, 282 limit, 283 losses of effect in, 303 mixed-flow, 284 outward-flow, 281 parallel-flow, 282 practical values of velocities in, 299 radial flow, 281 Schiele, 284 Scotch, 276 theory of, 284 Thomson, 282 useful work of, 292, 296, 297 ventilated, 278 vortex, 50 Undershot wheel, 225 " actual delivery in ft. - Ibs. of, 231 depth of crown of, 238 efficiency of, 227, 235, 239 form of course of, 236, in a straight race, 227 " losses of effect with, 228 " modifications to in- crease efficiency of, 231 " number of buckets in, 238 " Poncelet's, ^32 " Poncelet's efficiency of, 235, 239 useful work of, 228, 235 with flat vanes, 227 Uniform main, equivalent, 100 Unwin, 97 Valve, loss of head due to a, 93 Vane, best form of, 199 " cup, 195 Velocity, bottom, 151, 154, 155 " critical, 97 curve in a channel, 152, 154, " formulae, 150, 152, 154 Bazin's, 152 Boileau's, 153 " maximum, 151, 155 INDEX. 337 Velocity, mean, 151, 154 " mid-depth, 151 " of flow, 286 of whirl, 286 rod, 176 " surface, 150, 154 " variations of, 119, 131, 148 Velocities in turbines, practical values of, 299 Vena contracta, 14 Ventilated buckets, 272 Venturi water-meter, 13, 183 Virtual fall, 82 " slope, 10, 82, 84 Viscosity, 96, 97, 119, 149 Meyer's formula for, 156 Slotte's " "156 Vortex-chamber in centrifugal pump, 309, 313 Vortex, circular, 47 " compound, 50 free, 47 " free-spiral, 48 " forced, 49 Vortex, motion, 47 " turbine, 50, 282 Water-barometer, 5 Water-meter, 13 Weight of fresh water, 2 " " ice, 2 " " salt water, 2 Weir, 54 " broad-crested, 58 " rectangular, 54 Weisbach, 23, 26, 36, 76, 90, 91, 92, 93, 145 Wheel, breast, 242 hurdy-gurdy, 279 jet reaction, 272 overshot, 254 Pelton, 280 pitch-back, 272 Poncelet's, 232 Sagebien, 254 undershot, 225 Whirlpool-chamber, 50 Whirl, velocity of, 286 \ THIS BOOK IS DUE ON THE LAST DATE STAMPED BELOW AN INITIAL FINE OF 25 CENTS WILL BE ASSESSED FOR FAILURE TO RETURN THIS BOOK ON THE DATE DUE. THE PENALTY WILL INCREASE TO 5O CENTS ON THE FOURTH DAY AND TO $1.OO ON THE SEVENTH DAY OVERDUE. 1419! 6 1942 E LD 21-100m-8,'34 YC 13457 / C 1 iff} UNIVERSITY OF CALIFORNIA LIBRARY ' * * i