CENTRIFUGAL PUMPING MACHINERY Published by the raw -Hill Book- Company New Voirk. %Succe.s.sons to the Book Departments of tKe McGraw Publishing Company Hill Publishing- Company Publishers of Books for Electrical World The Engineering and Mining Journal Engineering Record American Machinist Electric Railway Journal Coal Age Metallurgical and Chemical Engineering Power CENTRIFUGAL PUMPING MACHINERY THE THEORY AND PRACTICE OF CENTRIFUGAL AND TURBINE PUMPS BY CARL GEORGE de LAVAL Member American Society Mechanical Engineers, Naval Architects and Marine Engineers, and American Society of Naval Engineers McGRAW-HILL BOOK COMPANY 239 WEST 39TH STREET, NEW YORK 6 BOUVEKIE STREET, LONDON, KG. 1912 COPYRIGHT, 1912, BY THE McGRAW-HILL BOOK COMPANY Stanbopc jprcss F. H.GILSON COMPANY BOSTON, U.S.A. PREFACE A WRITER upon centrifugal pumps has said that they defy the mathema- tician and possess more tricks than a circus mule. This book has been prepared with the idea of supplying accurate and definite information, which can be used in actual design. The data are based upon experience in design, construction and installation of this type of pumping machinery. It attempts to set forth the underlying principles, which, if properly applied and used, will give the designer enough information to enable him to calculate results with reasonable certainty. Almost all books on the subject in the English language are silent upon the principles which govern the practical designer, and give only empirical formulas which would prove very costly in actual practice. This book confines itself to material which has been used successfully in practice. The author has himself had charge of almost all of the installations described and they become of interest, therefore, as records of fact. Necessarily, the installations given are from the practice of Henry R. Worthington, and it has not been deemed wise or necessary to introduce other practice to illustrate the principles which the book seeks primarily to set forth. This is not merely because of the large number of centrif- ugal and turbine pumps which this company has put out, but because it seems best to confine the work to actual experience and to such installa- tions as have been in service sufficiently long to be beyond the experi- mental stage. No attempt has been made to go into the history of the subject, or to treat it from an elementary standpoint, as it is assumed that the reader is familiar with the laws of hydraulics. HARRISON, NEW JERSEY April, 1912. V 241322 CONTENTS. PAGE PREFACE >. v PART I. CHAPTER I. GENERAL REMARKS 1 Low- and High-lift Pumps Low-lift Pumps High-lift Pumps. CHAPTER II. DIFFUSORS 5 CHAPTER III. PALANCING THRUST 7 CHAPTER IV. PRIMING AND FOOT VALVES 8 Priming Foot Valves. CHAPTER V. EFFICIENCY 14 CHAPTER VI. CHARACTERISTICS 20 CHAPTER VII. OPERATING 27 Lubrication of Bearings Priming Starting Stuffing Boxes Suction Foot Valve and Strainer. PART II. CHAPTER VIII. GENERAL REMARKS 31 Discussion of the Centrifugal Pump. CHAPTER IX. FIRST THEORY OR ANALYSIS 32 Theory of Impellers Theory of Diffusion Guides and Vanes Application of Theory Blades Capacity of Centrifugal Pumps. CHAPTER X. SECOND ANALYSIS OR THEORY 48 Theory of Impellers Application of Analysis to Problem Short Method of Finding Characteristics. CHAPTER XI. GRAPHICAL ILLUSTRATION FOR DETERMINING THE IMPORTANT ANGLES 60 Method of Correcting Impeller-vane Angle. vii viii CONTENTS CHAPTER XII. THIRD ANALYSIS OR THEORY 66 Theory of Impellers Application of Analysis to Problem. CHAPTER XIII. SCREW OR PROPELLER PUMPS 73 PART III. CHAPTER XIV. GENERAL REMARKS 81 Waterworks Installation Waterworks Tests of Centrifugal Pumping Engines at Montreal. CHAPTER XV. IRRIGATION DRAINAGE AND SEWAGE 89 CHAPTER XVI. HYDRAULIC MINING AND DREDGING 100 CHAPTER XVII. MINING WORK 104 CHAPTER XVIII. POWER-STATION WORK 107 Boiler Feeding Circulation of Water Hot-well Pumps Centrifugal Jet Condensers. CHAPTER XIX. DOCKS 116 New Dry-dock at Norfolk Navy Yard. CHAPTER XX. CENTRAL FIRE-STATION SERVICE 129 CHAPTER XXI. FlREBOATS AND SHIPBOARD SERVICE 142 Fireboats On Shipboard. CHAPTER XXII. SPECIAL HIGH-SPEED INSTALLATIONS 147 CHAPTER XXIII. COMMERCIAL PUMPS FOR GENERAL INDUSTRIAL USES 153 PART IV. CHAPTER XXIV. ELECTRIC MOTORS 161 CHAPTER XXV. STEAM ENGINES AND MISCELLANEOUS 164 CHAPTER XXVI. STEAM TURBINES 166 APPENDIX . . 169 CENTRIFUGAL PUMPING MACHINERY. PAKT I. CHAPTER I. GENERAL REMARKS. LOW- AND HIGH-LIFT PUMPS; THE methods used for figuring and designing centrifugal pumps are usually regarded as more or less mysterious. This is due to the insufficiency of the data and information available in the English technical literature at the present time. Foreign writers have treated the theoretical side very carefully and thoroughly. Private investigations have been developing and enlarging upon the original theories and showing how they work out in actual practice, but only a few of these results have been placed in the hands of the public. The centrifugal pump presents many interesting phases which do not appear in any other style of pumping machinery, and these must be under- stood and their importance appreciated for the intelligent design, operation, or application of the pump. Many of the peculiar features have been demonstrated graphically, and should be carefully studied in Chapter V on Efficiency in this part. Some of these peculiarities afford a convenient classification of centrifugal pumps by characteristics. The present dis- cussion, however, will classify the pumps as low-lift and high-lift, accord- ing to the head pumped against. LOW-LIFT PUMPS. Although an increase in speed conditions may change a low-head pump to high-head, pumps intended for the latter service are differently con- structed, as will be seen. Low-lift pumps have, until recently, been little understood. In design and construction they have been crude and uneconomical, but pumps of this type are now built which secure high economy at both high and low heads. This has been the means of putting them into general use for a great variety of purposes. This style of pump is generally of the volute type. Under proper con- ditions it is probably the lightest and cheapest pump that can be used for moderate and large quantities of water. The best conditions are total 1 2 : CjENT&lfrtWAL PUMPING MACHINERY heads of from to 150 feet, short and direct suction and delivery pipes, moderate and large quantities of water never small quantities. The chief applications for low-lift pumps are : drainage, reclamation, and irrigation work; waterworks where the lift is not high and where low first cost is desirable; pumping into filter beds; sewerage work, dock work, sluicing, leakage in tunnels, circulating water in condensers for power stations, and general water-service pumps for buildings, mills, and heating plants. The method of applying the centrifugal pump to these services, and the conditions demanded for each service, are explained in detail in Part 4, and should be carefully studied by the designer and engineer when considering pumps for such installations. Centrifugal pumps will deal with very large volumes of water. Several have been installed which handle as much as 130,000 gallons per minute. These pumps can be built with a discharge pipe as large as 72 inches, and with pipe velocities of 8, 10, 12, 14, and 16 feet per second, corresponding to the volume of water. In fact, the amount of water that can be moved is almost unlimited, as there is no difficulty in constructing pumps for these large amounts of water for heads from to 40 feet and with smaller quantities up to 150 feet. The ordinary form of low-lift volute centrifugal pump is not adapted to high speeds, and for these the multirotor pump or some other type must be adopted. These types are described in the chapter, "Special High-speed Turbine Installations," at the end of Part 3. For very low heads or suction heads only and pumps working on sealed pipes, a combination screw and centrifugal pump or a double-screw pump is employed. This is treated separately in Part 4. HIGH-LIFT PUMPS. For heads of more than 150 feet, instead of the volute casing a round casing is used, which resembles that of a water wheel. Because of this resemblance, high-lift pumps are generally known as turbine pumps. It has only recently become known that the simpler form of centrifugal pumps, with few parts and only one moving piece, could be used for fairly high heads. In the Transactions of the Institution of Civil Engineers, London, Volume XXXII, it is stated that an 18-inch pump will work well on a 20-foot lift, and a 36-inch pump on a 30-foot lift. Other writers have stated that the ordinary centrifugal pump has a low efficiency at high heads. The relation existing between the actual pressure in the pump discharge and the theoretical pressure was not known, and but little attempt had been made to ascertain it. In the design of centrifugal pumps this ratio, called by some authorities the manometric coefficient based on experiments, must be the basis from which to work, together with the percentage of the useful work to that expended in operating the pump. GENERAL REMARKS s s 2 o - JOAV oj asaoH ai 1 13 s ^^01* . Tx ?^ "" > "" 7 \ 1 \ \ L \ 4 | 1 A \ i , \ \ \ S L \ \ t UyLpwag o g\o o\ ?||2|gSKSS "jaaj ui pBaH jaAv 0< i j JOH85I sug \ ^ / \ / / \ 2 \ / 7 ^ \ / / / PH V w /\ CO \ > V \ l(_ \\ =r p II \ \ \\ ^u 4ouaplBa\ 1 S V 5 ^V * a a?{T5.ig \ aapgsi| g i^T ! H 1 '!- > .2 = 11 3 p *o V 4 CENTRIFUGAL PUMPING MACHINERY Practical results have shown that single impellers without diffusion or guide vanes can be made for heads as high as 350 feet, and it is not unusual to find commercial pumps for heads of 150 feet in which the rapidly moving water leaves the impeller with very little loss from shock or eddies. It is of course necessary to have high peripheral velocities in order to obtain such heads. Figs. 1 and 2 give results for single impellers without diffusion vanes for heads from 40 to 135 feet, in the same pump, under different speeds. Very high heads can be obtained by multistage pumps, which consist of a succession of rotors connected in series through the casing, so that each stage draws its suction from the discharge of the preceding one and so raises the pressure progressively. Pumps of the multistage type are made for heads as high as 200 feet per stage when proper conditions are met as to capacity and speed. In good practice the head per stage varies between 100 to 150 feet, in order to keep down the velocity of the water so that it will not cause trouble from pitting the vanes or producing excessive wear on the impeller and diffusion tips. The velocity of flow in the suction inlets of the pump must also be kept low, or the water may separate from the entrained air and the pump become noisy, due to the high speed of the impellers and imperfect filling. CHAPTER II. DIFFUSORS. IN the design and operation of centrifugal pumps it is important that no head should be lost by shocks or abrupt changes of velocity. The ordinary method of taking care of abrupt changes in velocity consists in introducing a whirlpool chamber and volute, stationary or movable diffusors, or stream lines. In low-lift pumps the casing is usually in the form of a volute, with a gradually increasing cross section. In high-lift multistage pumps, each section or stage is fitted with a whirlpool chamber or a device with divided stream passages known as diffusion vanes which divert the water on leav- ing the wheel from a tangential direction into the proper one for discharge. This whirlpool chamber is usually an Archimedian spiral, to allow the water to move freely and to convert its velocity into pressure with as little friction as possible. It has been shown by experiment that a pump with a proper whirlpool chamber will work with greater efficiency and against a greater head than one without it. This is easily understood when we consider that the mass of water revolving outside of the wheel has some centrifugal force, which can be added to that produced by the wheel to increase the pumping head. Or, instead of this chamber, a stationary guide-vane chamber, called a diffusor, can be made. Both of these types are intended to utilize the energy of the rotating water as it leaves the wheel, and increase the pump- ing power. The guide vanes or diffusors must be so designed that the stream lines will not produce too much skin friction due to additional surfaces. Efforts have been made to have these guide vanes adjustable like those of the water turbine, but this has not proved practical. For the best effi- ciency, therefore, they should have the shape suited to the water path corresponding to the conditions for which they were designed. The vanes, in some cases, are made as a separate or removable casting, in others as spiral grooves cast in the main casing, or as passages in delivery compart- ments of the casings. The spaces in the vanes are supposed to have the shape which the water would assume in its passage from the wheel under certain and fixed conditions, but the exact conditions of flow are not known, and as the guides are always less in number than the vanes in the wheel, the water is liable to strike them at the entrances of the ring and thus cause injurious eddies. Diffusion rings can be made without vanes and good efficiencies obtained even for high heads. 5 6 CENTRIFUGAL PUMPING MACHINERY Investigation of movable diffusors for high heads, made to rotate freely on the impeller, has also been made, the results of which show that higher efficiency and heads can be obtained. This diffusor was originated by Mr. Barber, March 31, 1896, and by Professor Novak of Austria in 1908. Its purpose is to reduce the friction on the sides of the runners, and to pro- vide ample whirlpool for the diffusor chamber in the pump, doing away with the vanes usually employed. In the usual construction, the water between the runner or impeller and the stationary chamber has a tendency to churn, causing a waste of energy which increases with the clearance. A properly designed movable diffusor will greatly reduce this loss and thereby increase the efficiency. CHAPTER III. BALANCING THRUST. MORE or less axial thrust is present in all centrifugal pumps. It is caused by the unbalanced pressure between the impeller and casing, be- tween the impeller and channel or filling-in rings, and also where pressure acts upon unequal surfaces. The water in passing through the wheel alters its velocity and pressure, and in changing from an axial to a radial direction produces a centrifugal force causing a thrust. The vane angles have some influence upon the amount of thrust, dependent upon the veloci- ties at the inner and outer angles of the vane with respect to the impeller. The shape or form of the vanes may also cause a slight thrust under certain conditions. Very little is known definitely of this matter, but experiments have shown that if we assume an impeller with a cross section S, and a y 2 velocity Vz at the hub, the axial thrust will be approximately 3500 S = T, where T is expressed in pounds of thrust for each impeller. Various arrangements of impellers opposed to each other have been tried in order to eliminate this axial thrust. Impellers having two faces exposed, the smaller side to the higher pressure, the larger one to the lower pressure, have been used, and by so doing the thrust has been materially reduced; but, owing to complications introduced, this method is open to more or less criticism. Arrangements with bushing rings, and impellers with balancing holes, will materially aid in securing equal pressure upon the areas within the bushing rings. For the remaining unbalanced thrust a marine-type ball or roller bearing can be used on the outer end of shaft. Hydraulic thrusts of different types have been designed with a view of automatically adjusting themselves to the conditions, and of taking care of the additional thrust that may be produced by future wear at various points where leakage occurs. The most successful of this type, with a revolving steel disk having very close-running surfaces, is known as hydrau- lic step bearing, and can be operated by pressure from the discharge with very small loss. Another type consists of an internal disk step located either on the discharge side of last impeller or suction side of the first, and is designed to control the thrust automatically. An adjustable ring or collar within the pump, securable so as to control the opposing pressures in the chambers or casings and produce an opposite thrust, has also been designed. CHAPTER IV. PRIMING AND FOOT VALVES. PRIMING. BEFORE a pump is started all air must be expelled, and it must be filled completely with water. A centrifugal pump running in air cannot create the vacuum necessary to raise the water up to the impeller. Some pumps, particularly those handling hot water or any liquid giving off a vapor, Steam Ejector Steam "Valve Steam Fig. 2A. Methods of Priming. should always be so placed that the water will flow to the pump; otherwise the vapor will collect in the pump and it will cease pumping. Pumps so located that water will not flow to them must be primed. Methods of priming are illustrated in Fig. 2A. There are two principal methods of priming, i.e., by producing a vacuum with an air pump, or by pumping water into the casing; and all the devices 8 PRIMING AND FOOT VALVES 9 used are different forms of one of these methods. The conditions will determine whether steam, compressed air, or water should be used. A simple method is to have a foot valve on the end of the suction pipe, and to fill the casing through a small hole at the top with city water or other water under pressure. A second method uses the siphon principle. A foot valve is placed at the end of the suction pipe, a gate valve in the dis- charge close to the outlet, where it extends horizontally from bottom of pump, an air cock at the top of the casing, and water is supplied from a connection on the side. When water appears at the top pet cock the pump is primed. After closing the water supply, open the discharge gradually. This initial priming is sufficient for all future starting, pro- Fig. 2B. Priming Device for Small Centrifugal Pumps. vided that the foot valve is tight, and that the discharge gate valve is gradually closed as the pump is shut down. The water will then remain in the pump. For heads of over 30 feet, it is usual to place a check valve in the dis- charge in order to prevent shock on the pump, and it is customary to place a foot valve on the suction and a by-pass for priming between the dis- charge, above the check valve, and the suction pipe above the foot valve. Another method for use with steam or compressed air is to pump water into the casing by an injector. This should be so placed that it is within easy suction lift of the water. Priming without a foot valve, such as is necessary with pumps operating on wells where it is impossible to place one, requires a check valve in the suction pipe close to the pump opening. 10 CENTRIFUGAL PUMPING MACHINERY An ejector is connected at the top of the casing and its suction pipe tapped into the main suction under the check valve. A method of priming with a foot valve, and a check valve in the discharge, using an ejector as an exhauster, is also employed. For this method both steam and compressed air are used. On large engine-driven centrifugal pumps, running condensing, the top of the casing can be connected with the condenser and a sufficient vacuum created for priming. In such installations it is best to fit a glass water gauge on the top of the pump so that the operator may know when the casing is full and prevent the water from going into air pump, particularly if this is of the dry, rotative type. For small centrifugal pumps, up to and includ- ing 12 inches, a specially designed priming device may be used. (See Fig. 2s.) It consists of a fitting or casting placed directly against the suc- tion opening and takes the place of the usual elbow. It is fitted with a small hand pump, and a clapper or foot valve in the main passageway. Water is drawn through the main opening into the hand pump, and forced out through a small check valve into the main pump, the water being retained there by the main check valve in the primer. In order to prevent losses from friction in this type of primer, the openings are made very large. A similar type of primer providing the same features consists of an ordi- nary hand pump attached to the suction opening and connected with the water supply. A foot valve is required on main suction. Still another method is to attach a hand air pump onto the pump at any place below the discharge gate valve for exhausting the air between foot valve and discharge valve. In larger installations, separately driven electric or steam-driven vacuum pumps, operated automatically, are used, particularly where the suction pipes are large and quite long. In mining installations it is necessary with sinking pumps to use an automatic repriming arrangement, illustrated in Figs. 3 and 4, consisting of a foot valve in the suction and an automatically operated check valve Fig. 3. Automatic Priming Arrange- ment for Sinking Pumps. PRIMING AND FOOT VALVES 11 Section in Balanced 1 By-Pass Valve B Fig. 4. Automatic Priming Arrangement for Sinking Pumps. 12 CENTRIFUGAL PUMPING MACHINERY in the discharge, with necessary air-relief valves at the highest point of the entrance pipe to the pump. Its purpose is to discharge the air and to refill the suction pipe. The illustration shows the automatic check valve, which is wide open when pumping, allowing a free discharge, at the same time closing off the connection between the delivery pipe and the entrance to the first impeller. Should the pump stop for any reason, the check valve closes the connection from the last impeller to the column and allows the water in the column pipe to flow through the small pipe and submerge the impellers, priming the entire pump down to the foot valve. On starting up any accumulation of air in the suction pipe will be discharged through the automatic air-relief valve. When all the air has been dis- charged the air-relief valve closes. It is advisable in this design of pump to introduce the water on top rather than on the bottom. The location of the automatic relief valve allows the pump to free itself from air before the water enters the first impeller. The automatic operation of the discharge check valve can be accomplished in various ways, as shown in the illustrations. FOOT VALVES. So much trouble has arisen in practice because engineers have not appreciated the importance of a properly designed foot valve for a centrif- ugal pump that it has been deemed necessary to devote a separate section to this subject. The construction of a centrifugal or volute pump is weak in itself, as the pumping head is formed at the. periphery of impeller, and when the pump is working, the pressure on the side plates is much lower. These cannot be stayed in this form of pump, and are not supposed to stand the internal strain or pressure due to the total pump head. A sudden stoppage of the column of water traveling through the pipes causes a heavy pressure on the sides, tending to open and rupture the casing. The column of water is supported by the rotation of the impeller, and if from any cause this rota- tion suddenly ceases, the intensity of the reaction or shock is dependent on the weight of water in the pipes, and on the velocity acquired by the re- turning column before it is finally arrested. The sudden closing of the foot valve is frequently sufficient to split the pump casing, pipe and heads. This danger can be reduced by furnishing the foot valves with a relief or safety valve. This is particularly necessary when priming water is used under a pressure heavier than that for which the pump was built. A relief valve of this kind can be made a part of the foot valve, or it can be attached to suction pipe and discharge back into the well. It is not, however, advisable to employ foot valves on large pumps, and other means for priming the pumps should be used. Foot valves should have at least 150 per cent of the area of the suction PRIMING AND FOOT VALVES 13 pipe, which should be the next size larger than discharge. Thus a pump with 10-inch discharge should have 12-inch suction pipe, and if a foot valve be employed its area should be one and one-half that of the 12-inch pipe to reduce the frictional resistance through the system; otherwise these losses Fig. 5. Centrifugal Foot Valve. Dsorb a large percentage of the total work in low-lift pumps, which means poor economy, preventable with proper sizes of piping. The valves should be of the flap design, so made that when open they rest on the sides of the body, thus allowing a clear passage for water through le center of the valves, as illustrated in Fig. 5. CHAPTER Y. EFFICIENCY. THE word efficiency in connection with centrifugal pumps has become very ambiguous, and has led to many disputes in connection with guaran- tees and contracts. This ambiguity has been brought about by the fact that the word has been used without modification to designate the efficiency of the pump only, of the pump and prime mover, and of the entire plant measured back to the boiler. This uncertainty can and should be elimi- nated. The efficiency of a centrifugal pump, when not otherwise modified, can mean but one thing, the ratio of the water horse-power output at the pump to the brake horse-power input at the coupling or pulley. The water horse-power output is determined by the total head against which the water is pumped and the quantity of water delivered. The usual method of finding the total head is to place a gauge on the suction line close to the pump, another on the discharge, and to note the vertical distance between the gauges. The algebraic difference between the gauge readings plus the vertical distance between the gauges, all in feet, is con- sidered the total head. Another method is to add to the head thus ob- tained the velocity head in the discharge pipe. Still another method is to add to the difference in gauge readings and the vertical distance between gauges, the difference between the velocity heads in the suction and dis- charge pipes. It is therefore important in considering efficiencies that the total head be clearly interpreted. /\^ The three methods of obtaining the total head may be represented by the following formulae : H = Total head in feet. HI = Discharge head in feet. Hi = Suction head in feet. A = Vertical distance between gauges in- feet. HS = Velocity head in discharge pipe in feet. H 4 = Velocity head in suction pipe in feet. First, H = H 1 -H 2 + A. Second, H = HI H 2 + A + H%. Third, H = H, - H 2 + A + (# 3 - H*). HZ and H^ are made up of the flow in feet per second through the pipes thus : TT (velocity in feet per second) 2 3= ~W in which g is the acceleration due to gravity, or 32.2. 14 EFFICIENCY 15 Third formula represents the actual head pumped against and should always be used since the gauge reading shows the difference between the total head above the gauge and the velocity head. Where the suction and discharge pipes are of the same diameter, the velocity heads are equal, and equation (3) becomes the same as equation (1). High efficiency in pumps is obtained by changing the kinetic energy of the water, as it issues fro.m the wheel, into pressure, by reducing water friction, such as churning in the chambers, and the skin friction of rotating disks, and by reducing the friction of the bearings. The elements of total pump efficiency are therefore the absolute hydraulic efficiency, the mechani- cal efficiency, and the volumetric efficiency. The absolute hydraulic efficiency r, h + expresses the ratio between the use- ful head and the total pumping head. If h represents the former -and H the latter, then the total head H is made up of h and all frictional and shock losses of water in the pump, and is the most important factor to be considered. Information relating to these losses is very meager and incomplete. The most serious loss is due to skin friction between the impeller and the water. The head pumped against varies as the square of the velocity; hence the wasted power varies as the cube of the head. As the head increases, the loss from skin friction increases at a more rapid rate and thereby imposes a limit on the head against which the pump can be economically operated. The work wasted in disk friction varies as the square of the radius, hence a smaller impeller at a higher number of revolutions absorbs less power in friction than a larger one at a less number of revolutions but with the same peripheral velocity. The disk friction of the water near the axis is lower than that at the outer surface of the impeller. Experiments have been made abroad on the power lost by skin friction and the following formula has been obtained: W = 8132 X - 2 X h 2 - 5 foot pounds. W = Power due to resistance in foot pounds, n = Revolutions per minute, h = Head in feet. The largest losses by surface friction are along the walls of the casings, and were the surfaces of the stationary walls and impellers alike the water ^ would receive a rotating motion one-half that of the impeller. These losses are reduced by having the internal walls smooth and impellers polished, and by proper clearance between impellers and casings. In addition to surface friction there are losses due to molecular friction in the water itself. The mechanical efficiency depends upon the friction between the shaft, \/ impeller, etc., and their bearings, and is a function of the workmanship, fit, and lubrication. 16 CENTRIFUGAL PUMPING MACHINERY The volumetric efficiency is the ratio of the amount of water discharged to the amount entering the pump. The difference is the loss due to leakage in running fits. The greatest loss in a turbine pump is between the impeller and the diffusion ring. These losses vary from 2 to 10 per cent. Efficiencies on large pumps have been obtained as high as 90 to 92 per cent, and in multistage turbines as high as 85 to 87 per cent. This has been accomplished by careful analysis, as shown above, by reducing the internal losses and skin friction, by dissipating shocks and disturbances and turning more of the velocity into effective head. The advance which this involves is shown by the fact that until recently the best efficiency obtained from the ordinary volute was 40 per cent. The conditions and approximate efficiencies which may be expected with correctly designed impellers are as follows: "C Capacities from 75 gallons to 250 gallons will give about 55 to 65 per cent efficiencies ; 250 gallons to 900 gallons, 70 per cent; 900 gallons to 3000 gallons, 70 to 73 per cent; 3000 gallons to 6000 gallons, 73 to 75 per cent; and 6000 gallons to 10,000 gallons, 75 to 78 per cent. Above this we obtain from 75 to 85 per cent efficiencies. A side entrance or single suction pump will give slightly less in each case. Sizes of the discharge pipes of pumps for the above vary from 4 to 60 inches diameter. Speeds for the small pumps vary from as high as 3500 revolutions per minute for the smallest down to 600 revolutions for 8- to 10-inch pumps. The larger may run as slow as 150 revolutions with good efficiency under favorable conditions. The losses at the impeller outlets are due to the discharge velocity and can be expressed as (aU 2 2 g & in which W the weight in pounds of the water pumped per second; H = the vertical distance in feet between the levels of the water in the suction and discharge reservoirs; v and V = the tangential velocities of the wheel in feet per second at its inner and outer periphery respectively; u and U = the tangential velocities in feet per second of the water at the inner and outer periphery of the wheel respectively; g = the acceleration due to gravity ; E = the hydraulic efficiency of the wheel. The left-hand member of the above equation represents the work re- quired to overcome the hydraulic resistance of the wheel, but does not include any mechanical resistances such as friction of the bearings, etc. The right- hand member represents the work done. This equation contains the fundamental principle used in calculating or- dinary centrifugal pumps as given by most of the writers. As stated, it is assumed in ordinary calculations for centrifugal pumps that the water enters the eye of the wheel radially, and consequently the term vu becomes zero, and we may write V = Jriy From the triangle of velocities in Fig. 28 it will be easy to deduce the following formulae : /gH(sm(A-~B)\ JgH ( tangZA /gH / tang~B\ " V E \smAcosBj V E \ tangA/ V E \ r tang D/ IgH/ 1 + cotT\ V E \ ^ cot CJ tangg = tangg = cot A = cot T EV 2 tang A tang D cot B ~ cot C ~ " SH ' 32 FIRST THEORY OR ANALYSIS 33 and j = cot B - cot A = tang T + tang C, cot B = The above embrace all the formulae given hi various forms. The greater the number of blades the more nearly will the assumption agree with the facts, but too many blades will cause undue friction. Brix and other writers give complicated formulae for the number of blades, but this is usually < Tang.Vel. U ofWi Cot.D E = V (Y- Fig. 28. Triangle of Velocities. assumed. The diameter, speed, and width of the wheel are dependent on the motor and on the allowable velocity of the wheel and the water through it, features which each designer will have to settle. The angle of the blade as well as its radius is determined by the following geometrical and analytical considerations. To construct the curve of the blade, it is necessary to know the inner and outer radii of the wheel, and the angles < and D which the blade makes with the inner and outer periphery of the wheel. The angle D is determined by calculation from formulae given, and usually the radial velocity at the inner periphery of the wheel is assumed to be the same as that at the outer; the angle , unit of blade, may be graphically determined as shown in Fig. 29. 34 CENTRIFUGAL PUMPING MACHINERY Construction of blade Given R,r, D,and$ Referring to Fig. 29 we proceed as follows: Let r and R be the radii of the wheel. From a draw a line, making the angle D as shown, and a line from the center 0, making the angle D + <. From a draw the line abc] then c will be the inner extremity of the blade. Draw the line de perpendicular to and bisecting the line ac ; then the point d will be the center of the curve of the blade as shown. This assumes that the angles < and D are known. To find the angle 4>, the veloc- ity v at the inner circumference of the wheel varies inversely with ratio of the radii and directly with the velocity of the wheel at the outer circumference, that is, Fig. 29. Construction of Curve of Blade. Since it has been assumed that the water enters the eye of the wheel radially, it has no velocity of rotation, and the triangle of velocities is as shown in Fig. 30. If Q be the volume of water in cubic feet pumped per second, w the width of the wheel at the outer circumference, then, disregarding the thickness of the blades, the radial velocity of the water at the outer circumfer- ence of the wheel will be Q Determinate n of Angle d> 2irrw' all dimensions being in feet. The relative velocity of the water at entrance and exit of the wheel should have the same direction as that of the blade at the corresponding points, and this is what governs the angles D and <. If this is not followed injurious shocks will result. This discussion does not consider the influence of a vortex chamber or the form of the discharge passage around the wheel. ab Fig. 30. Triangle of Velocity with Water Entering Wheel Radially. FIRST THEORY OR ANALYSIS 35 The calculation of an 8-inch three-stage pump, with details of impeller and diffusion vanes, according to the above analysis, is shown hi Fig. 31 and described below. Impellers and Diffusion Rings Fig. 31. Impeller and Diffusion Vanes of 8-inch 3-stage. CALCULATION OF 8-INCH 3-STAGE Conditions: 1400 gallons per minute. 1170 revolutions per minute. 210 pounds pressure or 70 pounds, each stage. Outer diameter of impeller 18 Jf" Inner diameter of impeller 8|" Hub diameter of impeller 4f " Width of opening at outer circumference of impeller 1" Width of opening at inner circumference of impeller 2f" Number of blades 12 Thickness of blades at outer circumference f " Outer diameter of diffusion ring 27" Inner diameter of diffusion ring 18%" Number of guides in diffusion ring 10 36 CENTRIFUGAL PUMPING MACHINERY Formula Used. JS . tan*, + tanft. DttfoBion Bing vf J sfficiency. tan 02 / 1 g H cot z VzU^ tan 2 vtfiz H = net or useful, here 70 pounds head. All dimensions in feet; velocity in feet per second. log r x = log r 2 + cot fa X ~ X log e. 180 For equal steps of 4> there is a ( Trcotfo log e I s ^ _ ^ constant difference, for log r x = ( 180 j These dimensions in inches. 18.46875 X 1170X7T 12X60 d 2 - 18.46875 log 1 . 2664374 TT log 0. 4971499 1170 log 3. 0681859 4. 8317732 60 X 12 = 720 log 2.8573325 v z = 94. 284 ft. log 1. 9744407 18. 46875 log 1.2664374 TT 0. 4971499 1.7635873 Outer circumference of impeller = 58. 02 in. Blades 12 X I = 4 - 50 m - Net outer circumference 53. 52 in. As opening is 1 inch wide this is also the area of the opening in square inches. 1400 X 231 60 5390 log 3. 7315888 53.52 log 1.7285161 12.00 log 1.0791812 2.8076973 8. 3925 log 0. 9238915 r 2 = cos 5 2 = 8 log 554 591 + 9H = 554 591 2.7435098 2.7715875 or log cos of 20-22'-53. 55" = 9. 9719223 cos 5 2 = sin 2 tan 2 = cot 5 2 . 5 2 = 20-20'-53. 55" log cot 0. 4307806 tan of 2 = tan69-39'-6.45" v, = 94.284 log 1.9744407 u z 8. 3925 log 0. 9238915 ^ 11.234 log 1.0505492 tan nat. 2 tan nat. < 2 2.692 8. 542 FIRST THEORY OR ANALYSIS 37 = 830-19'-22. 28" tan 2 = 2.692 log 0. 4307806 tan 02 = 8. 542 log 0. 9315596 tan 2 -J- tan 02 = 0. 31566 log 9. 4992210 = 32.16 log 1.5073160 H = 70 Ibs. = 161 . 63 ft. log 2. 2084414 3. 7157574 t> 2 2 log 3.9488814 9. 7668760-10 log 0.1 191007-10 tan0 2 Efficiency 17 1.31566 0. 76908 log 9. 8859767 This is the hydraulic efficiency as computed from the formula. Diffusion Ring. Calculation of one radius vector of the curve of guide. X log T X = log r 2 + cot 02 X 180 Xloge. re0 = 36; then log r x = log 9.25 + cot 02 X X log e o cot 02 log 9. 0684404-10 TT log 0.4971499 log e log 9. 6377843-10 9. 2033696-10 5 log 0. 6989700 0. 0319447 log 8. 5043996-10 log 9. 25 0.9661417 ^ cot 02 log e 0. 0319447 0. 9980864 r x = 9.956" r 2 = 9.25" diff. =0.706" 2 Gain in pressure in free vortex = ~- j 1 ( t* u z u 2 = 8. 3925 log 0. 9238915 tan 5 2 tan 5 2 9. 5692194^-10 22.629 9.3546721-10 v 2 = 94.284 -f- tanS 2 = 22.629 u, tan 6 2 71.655 log 71. 655 or ^ = 1. 8552464 log o> 2 2 = 3. 7104928 V 1547 j 2916 1547 log 2916 log 3. 1894903 3. 4647875 9. 7247028-10 2. 9 = 64. 32 log 1.8083460 0. 0082481 log 7. 9163568-10 -6T 3. 7104928 7. 9163568-10 1.6268496 42. 349 ft. gain in head, or 18. 34 pounds. 38 CENTRIFUGAL PUMPING MACHINERY To set this before the reader in simple form, the following diagrams, with calculation, have been developed, which can be applied to impellers and diffusion vanes in turbine pumps (Fig. 32). STUDY OF CENTRIFUGAL PUMP. ' k q \ 1 Graphical Solution. Ha = useful lift in feet. 77 < 1 = efficiency. All dimensions in feet. All velocities in feet per second. Radial velocity = cubic feet per second divided by the net circumferential area in square feet; that is, allowance must be made for thickness of blades. Fig. 32. Diagram for Calculating Impellers and Diffusion Vanes. = y 2 2 c 2 2 (l - sin 2 2 ) = t' 2 2 c 2 2 cos 2 2 sin 2 4 cos sin 2 (a t - . ' hence, tan 2 _ 1 tan 2 tan a.-! ~ tan 2 u a Z-L = COS 02 COS 02 V 2 = c 2 sin 2 + z 2 sin 2 _ sin 2 sin 2 2 cos 2 2 cos 2 = Uz tan 2 -}- u X log e, where a is the radius of the outside of the wheel, is the angle in circular measure which the radius through the point in the curve makes with the radius of the point at the outer circumference of the wheel, and e is the Napierian base. If is the angle in degrees, then = X 3.1416 -r- 180. Example. a = 9.25 inches; B = 80; = n X 10 X 3.1416 -4- 180, where n is a multiple of an angle of 10. We then have log r = log 9.25 -h (cot 80 X n X 10 X 3.1416 X log e -4- 180). cot 80 log 9.2463188 - 10 cot 10 log 1.0000000 3.1416 log 0.4971499 log e log 9.6377843 - 10 0.3812530 180 log 2.2552725 0.0133654 log 8. 1259805 - 10 9.25 log 0.9661417 0= (log r - log a) X 180 3.1416 X cot B X loge log r = 0.9661417 + 0.0133654 X n. If = 30, then n = 3 and logr = 0.9661417 + 0.0400962 = 1.0062379 and r = 10.14467. Suppose r is given and we wish to find the corresponding angle , then we have r = 10.14467 log 1.0062379 a = 9.25 log 0.9661417 0.0400961 40 CENTRIFUGAL PUMPING MACHINERY 0.0400961 log 8.6031020 180 log 2.2552725 10 0.8583745 9.3812530 - 10 3.1416 cot 80 loge log 0.4971499 log 9.2463188 - 10 log 9.6377843 - 10 9.3812530 - 10 = 30.00001 log 1.4771215 or 30, which checks = 30 = 10 X n = 10 X 3. Fig. 34 will illustrate graphically this method applied to example. Fig. 34. Graphical Method for Diffusion Vanes, log r = log a + cot B X X log e, given * = ^w L = (log r - log a) X 180 unknown TT X cot B X log e given r As frequently stated, the object is to utilize the velocity of the water as it issues from the wheel by changing the energy of motion into energy of pressure. To obtain the full benefit of a vortex chamber, the direction of flow of the water must always be the same; in other words, the angle which the water path at any point makes with the corresponding radius must be a constant. Furthermore, the line of direction must be that of the absolute velocity of the water as it leaves the wheel. The equation to this curve, which is a logarithmic, spiral is r = ae m when = 0. e = Napierian base, r = a. m is the point determined by calculation from the problem. In the ordinary type of turbine pump the water cannot take this path, except FIRST THEORY OR ANALYSIS 41 the last wheel, but must flow around the circumference of the wheel. An easy turn is made from the correct curve. The area of the passageways at the circumference is equal to the area of discharge pipe, or slightly larger, that there may be no sudden velocity changes. In some recent pumps the flow of water throughout the pump from suction to discharge is in a con- tinuous spiral path in order to obviate all shocks and sudden changes. The curves depend, however, upon the fact that the radial flow in the dif- fusion ring varies inversely as the distance from the center, and the curve of the path is the resultant of the absolute and radial flow. A method of constructing diffusion rings by aid of circles, not governed by any laws, is shown in Fig. 35. The logarithmic spiral curve will give the best efficiency. A simple way to construct curves in the diffusion ring is shown in Fig. 36. Let r x be any radius of the curve calculated by r x = R e e * *, where in 0X 3.1416 circular measure is = 180 C Bisect angle by line o and prolong. The bisector cb is the mean proportional of the radii R and r x . Erect perpendicular od. Describe arc cdb on cb as a diameter; then od, cut off by the arc, is equal to the radius of. a By bisecting - we obtain another point in a similar manner, and thus as ft many points as are desired may be found. NXArea=Discharge Pipe N=No. of Vanes Fig. 35. Fig. 36. Method of Constructing Diffusion Vanes by Circles. APPLICATION OF THEORY. In order to cover the details involved in the mathematical consideration of the problem we will calculate a concrete case, and will consider a pump having 12 stages with approximately 50 feet to the stage. It is now possible 42 CENTRIFUGAL PUMPING MACHINERY to work single stages as high as 150 feet and even more under proper con- ditions, but we make the above assumption in order to illustrate Pump, 4-inch, 12-stage. Capacity, 150 gallons per minute. Speed, 1460 revolutions. Total or useful head, 610 feet. Diameter of inlet and outlet, 4 inches. Outside diameter of diffusion ring, 13 inches. Inside diameter of impeller, 3J inches. Diameter of hub of impeller, 2 inches. Outside opening of diffusion ring, J-inch wide (assumed). Outside diameter of impeller, 8jf^ inches, to suit speed. Width of opening of impeller at outside, f inch. Inside diameter of diffusion ring, 9 inches. Angle which tangent to blade makes with radius of outside, or 2 = 72. Number of blades, 12. Calculations. Tangential velocity of outer circumference of impeller _ 1460 X 8 J> inches X IT _ 146 X 287 X TT 60 X12 2304 146 log 2.1643529 287 log 2.4578819 TT log 0.4971499 5.1193847 2304 log 3.3624825 v 2 = 57.13500 1.7569022 feet per second. Radial velocity of water at outside, 150 X 231 1 60 i X f X TT - 12 X t X f The thickness t of the blades at the outer circumference measures f , and 12X| = 4.5. The circumference of 8| = 28.676; of 28.676 - 4.5 = 23.676, leaving a net length, then 150 X 231 60 385 . 12 7L028 ' ?* = 5 -^- = 10.56 and - = 10.56 = tang 2 + tan 2 Uz O.4Z U-2. tang 6 2 = tang 72 = 3.0777 tang = 10.56 - 3.0777 = 7.4823, FIRST THEORY OR ANALYSIS 43 we have gH gH 32.16 X610 u tang fa tang fa = n 77 ?/ 2 t; 2 tang4> 2 12 X u 2 X v z X tang< 2 32.16 log 1.5073160 610 log 2.7853298 4.2926458 12 log 1.0791812 5.42 log 0.7339993 57.135 log 1.7569022 7.4823 log 0.8740351 er cent effi( 4.4441178 0.70555 1.8485280= 70 p Path of the water or curve in diffusion ring, r x = radial distance to any point = r x = r e cot fa X , or log r x = log r + < X cot fa X log e , where r inside radius of impeller = 4J inches. = circular measure of angle between r and r x (see Fig. 37) , in de- , grees to e = Napierian base. r x is very simply calculated : (1) r xl = log r + cot fa X X log e . (2) r x = log r + cot faX'x log e . (3) r x " = log r + cot < 2 X <" X log e . The difference between (2) and (1) and (3) and (2) is ('-<) (Cot fa loge), and by making r x progress by uniform steps this is a constant. Make the difference in angles 5; then _5XxXir_Xir HLSOT ~36' where x = the multiple of 5 in the angle 0. Then XTT r x = logr + cot< 2 X Xloge XTT log 4.5 + cot fa X X log e 44 CENTRIFUGAL PUMPING MACHINERY Note where the diffusion ring is used 2 is not so important as without. tang 02 = 7.4823 log 0.8740351 cot 02 = log 1.1259649 . ^ tang 02 log cot 02 = .1259649 TT 0.4971499 log of log e 1.6377843 log 36 1.2608991 1.5563025 = 0.0050652 log 3.7045966 log 4.5 0.6532125 = 5 10 15 20 25 30 35 40 45 50 55 Constant 0.0050652 r x r x = r log 0.6532125 4.5 r x = log 0.6582777 4.552 0.6633429 4.606 0.6684081 4.660 0.6734733 4.715 0.6785385 4.763 0.6836037 4.826 0.6886689 4.883 0.6937341 4.940 0.6987993 4.999 0.7038645 5.057 0.7089297 5.116 We have simply to add constant above to log of r. To obtain r x = for = 50 we have added the constant ten times: 10 X 0.0050652 = 0.0506520 log r = 0.6532125 log r x for = 50 0.7038645 This checks the work. The next step is to find the arc of a circle which shall pass through the greatest number of points as given by r x (see Fig. 38). The center of this curve should be on a line perpendicular to the tan at r = 4.5, or very nearly so. Regarding the diffusion ring we may make the following observations: The width, or difference between the inner and outer radii, varies with different builders, as no one can tell the exact jrj gs 37^ 35 an( j 39 Method of Designing Diffusion Vane. FIRST THEORY OR ANALYSIS 45 gain in passing through the ring. It will not do to have too large an angle 02 or too many guides, as the space becomes so small that the velocity is very great, and it should not be more than the absolute velocity of the water leaving the wheel. Where the water discharges around the ring the curves at the end should be radial as shown in Fig. 39. The angle ^ is easily calculated, thus I X w when w = width is found; then d* * TT = ratio of total opening, Dwir with reference to discharge pipe. Here D = 13; w = \ inch; d = 4 inches; 16 4 then 13 16 2 8^ 4 13 13' X 360 = lo = 221.50 We find 6 guides give about the right passage for the veloc- ity and space at the outer cir- cumference ; we have, therefore, = 36.9. We will call it, Figs. 40 and 41. Method of Designing Impeller Blade. Fig. 42. Section of Impeller. for practical purposes, 35, and then round the corners. The water will j then enter the passage around the wheel at about the same velocity as the [water in the passage. When this is not the case there should be a free ! passage on an Archimedean spiral. 46 CENTRIFUGAL PUMPING MACHINERY BLADES. A few remarks only are necessary. To be sure of the inner and outer angles the center 0, Fig. 40, should be such as to give a short straight line at both ends, a and C. Strictly speaking, the water should enter the im- peller in the direction of the blade at the inner end, and as it is usual to make the radial velocity the same at the inner and outer circumference this angle may be determined as per Fig. 41 ; but owing to the doubt about the real direction of entrance, this angle is usually made somewhat larger. This is wholly a matter of judgment, as no test can absolutely de- termine it. Much has been' written on it, and, strictly speaking, the blade, having the section A, Fig. 42, should be calculated for each radius, and formula have been given for this. But as the calculation is complicated and founded wholly upon assumptions, it is questionable whether it pays to consider this. In fact, most authors assume the tangential velocity of the water at the entrance as zero. CAPACITY OF CENTRIFUGAL PUMPS. The velocity of flow through the impeller governs the capacity. When this velocity and the circumferential area, deducting the thickness of vanes, are known, the total capacity can easily be calculated. The speed for a given discharge can be calculated and depends upon the head against which the pump is to work. The water is supposed to rotate in the pump as a solid mass, and delivery commences when the centrifugal force is greater than the total lift including all losses and friction. Let HI = absolute velocity inlet, Uz = absolute velocity outlet. u r = - 0.84913453. s = 0.010273562, t = 0.00001747915. For 580 r.p.m.: * H.P. = - 0.89545 Q 2 + 6.2837 Q + 6.2. L = sec. ft. pound. H.P. = horse power. 800 1000 1200 1400 1600 Gallons per Minute Constant Speed. Fig. 47. increase this capacity the speed must be increased, as the capacity varies directly with the speed at constant head. With a constant head, and a variable speed and capacity, the efficiency and horse-power curves become Capacity Fig. 48. Fig. 49. Power Curves Showing the Effect of Different Vane Angles. different, and a point is reached when no increase in capacity can be obtained. The power curve is P = rnQ 2 + sn*Q + tn*. 52 CENTRIFUGAL PUMPING MACHINERY The horse power is directly proportional to the square of capacity if the speed is constant, and directly proportional to the cube of revolutions if the capacity is constant. When the discharge valve is closed and there is no delivery, the horse power is directly proportional to the cube of the revo- lutions. With the speed constant, the equation becomes a parabola, with the capacities as abscissae and the power as ordinates. Figs. 48 and 49 show that the variation in the power depends upon the angles. There are three different parabolas, 1, 2, or 3, depending upon the angles. We have case 1 if we select (az + a 3 ) < 90 degrees; case 2, nearly a straight line, if the angles (az + s) = 90 degrees; and case 3 if the angles (az + 3 ) > 90 degrees. The equation will solve the question of maximum power at constant speed for the maximum capacity, as Q ma *H = 0. APPLICATION OF ANALYSIS TO PROBLEM. The following will illustrate the analysis of any problem. 77 = total efficiency of pump; 7) m = mechanical efficiency; r) v = volumetric efficiency; rjh = total hydraulic efficiency; 77/11= absolute hydraulic efficiency; rj e = efficiency motor; f] = over-all efficiency or wire to water. Total efficiency is the relation between the brake horse power of motor and the actual water horse power. rr. 4. i /c f water horse power Q X H X 8.33 X 60 Total efficiency of pump, r = = - = - brake horse power 33,000 V A , . , r j- voltage X amperes Electric horse power for direct current = _._ 74o VXA Brake horse power, fj e X _. n rr. . water horse power r/o, over-all efficiency, -. . r electric horse power Break-horse-power output for alternating-current motors _ volts X amperes X Vn X cos X 77 746 where n = number of phases; cos = power factor of motor; 77 = motor efficiency. The speed of motors at full load will vary from 2 to 5 per cent on a motor of 200 to 10 horse power, and from 1 to 2 per cent from 500 to 200 horse SECOND ANALYSIS OR THEORY 53 power. Small motors of 1 to 10 horse power may vary from 10 to 15 per cent. The ratio between the number of poles and the speeds and cycles is as follows: When/ = frequency, p = number of poles. r = number of revolutions; r = 120-- The slip is expressed in a percentage of the actual revolutions. Actual speed = revolutions (1 per cent of slip) . This slip is due to the resistance opposed to the rotor current. The elements of total pump efficiency are, therefore, made up of Mechanical efficiency = rj m . Volumetric efficiency = t\ v . Absolute hydraulic efficiency = T//^. Total hydraulic efficiency = rj h = T?A< X Tj r . Total pump efficiency = 77 X ij v X r] h t- The elements of mechanical efficiency are made up of friction of shaft and impeller, etc., in their bearings and is a function of the workmanship and fits. The clearances vary from 0.002 to 0.005 inch, and should never be larger than 0.006 inch. The volumetric efficiency, rj v , expresses the ratio of the amount of water entering the pump to that which is discharged, the loss being due to leakage in the running fits. The greatest losses in a turbine pump occur between the impeller and the diffusion ring. These vary between 2 to 10 per cent. The absolute hydraulic efficiency ri ht expresses the ratio between the useful head and the total pumping head. .If h represents the former and H the latter, then the total head H is made up of h and all frictional and shock losses of water in the pump, and is the most important factor to consider in designing pumps. Information re- lating to these losses is meager and incomplete. Another serious loss is due to skin friction between the impeller and the water. The skin friction increases with head pumped against more rapidly than the head increases with respect to the velocity. The head against which the pump operates varies as the square of the velocity and the wasted power as the cube of the head, or H 3 . The work lost in disk or skin friction varies as the square of the radius, therefore a small impeller at a high number of revolutions will waste less power than a large one running slower, both having the same peripheral velocity. The largest losses by surface friction occur along the walls of the casings, and if the surfaces of the stationary walls and im- pellers were alike the water would have a rotating motion at a speed one- half that of the impeller. This loss may be reduced by having the walls smooth and the impellers polished and by having proper clearance between impellers and casings. Experiments have been made abroad on the power loss by skin friction, and the following formula has been obtained : n* 54 CENTRIFUGAL PUMPING MACHINERY W = power due to resistance of rotating disk in foot Ibs. ; F = constant 8132; n = revolutions per minute; h = head in feet. In addition to the surface friction of the water, there are losses due to molecular friction. The head is shown to be directly proportional to the square of the revolutions and the power lost on account of friction to some power of revolutions. The results show that these losses are greater in] high-head pumps than in low-head pumps. In designing we may expect a total efficiency of 90 per cent in the large pumps, under favorable conditions, with a lower efficiency in the smaller ones. It is absolutely necessary to select the right efficiency when den signing this class of pumps, and this can be obtained from tests. The curve of relationship between the capacity and efficiency is a parabola commencing at zero, its vertex showing the maximum efficiency, and coming down again to zero as the head approaches zero. The maxi- mum point or vertex should occur under the conditions for which the pump is designed. The general equation for the impeller is Av 2 - Bu 2 v - Cu 2 2 = -2gh, where u 2 = circumferential speed of impeller; v = radial speed of water on outer circumference of impeller; g = acceleration; h = head. A, B, C are constants. This equation is analogous to the capacity-head curve, v being directly proportional to the capacity FXv = Q. Q = capacity; F = sectional area. u 2 is directly proportional to the speed 2riru U= ^' The general equation for capacity head is a hyperbolic parabola, aQ 2 -(3uQ -yu 2 = -2gh, where Q, h, and u are variables. If n is constant it becomes a parabola as illustrated in Fig. 46 on page 50. The figure shows the hyperbolic paraboloids for speeds at 600, 800, 1000, and 1200 revolutions. Fig. 50 illustrates the curve with a constant head and the capacity and speed as variables. It represents an actual perform- ance under test. The equation aQ 2 (3uQ yu 2 2 gh becomes a hyper- bola. The figure shows the relations between these hyperbola and the SECOND ANALYSIS OR THEORY 55 capacity and speed at constant heads of 20, 25, 35, and 45 feet. Fig. 51 shows the same condition of a larger pump operated under constant head and a variable speed and capacity. . 650 Callous per Minute Fig. 50. Curves of Relation between Capacity and Efficiency. As an example, Fig. 52 will illustrate the method used in designing im- pellers. General equation, Av 2 - Buv - Cu 2 =-2gH, 3 = ^ ' at M M M m a _^ - ^ >* L == - I i 833813832 s 3 3 J 2 3 2 Curve is Figured Gallons per Minute Fig. 51. Curves of Relation between Capacity and Efficiency. A, B, C being constants or coefficients. H = head. 2g = 64.4. For v and u see Fig. 52. r , capacity in cubic feet per second . v = feet per second = 77^- - f r^ 7 - , (D 2 X TT s X z)b in square feet D 2 in ft X TT X revolutions per minute u = feet per second = The equation 60 Av 2 -Buv - Cv? = -2 gH 56 CENTRIFUGAL PUMPING MACHINERY represents a hyperbolic paraboloid, and by taking u, v, or H as constants the three characteristic curves of the pump can be obtained, namely: Revo- lutions constant, capacity and head varying; H constant, capacity and revolutions varying; capacity constant, head and revolutions varying. Fig. 52. Double Suction Impeller. A, B, C depend upon the design of pump and impeller, and can be found from previous available tests, or approximately from the design of the impeller. The impeller shown in Fig. 52, for a 66-inch pump, gives the| following values for A = 4.15; B = 0.17; C 0.98, making the equa- tion read 4.15 v* + 0.17 uv - 0.98 u 2 = -64.4#. Capacity H Constant /'Rev. per Minute Capacity Constant Fig. 53. Equation Curves for Impellers. In order to determine these constants from the fundamental equation, which contains the variables Q, the capacity; u, the revolutions; and H, the head, one can be taken as constant and we can obtain three curves as per Fig. 53. These are the characteristic curves for constant head, con- SECOND ANALYSIS OR THEORY 57 slant speed, and constant capacity. The values of A, B, and C are usual- ly, however, determined by tests, for which the following readings are required as per Fig. 54. Points 1, 2, 3 are test readings with respective >Capacity heads Hi, H 2) H 3 , and capacities Qi, $2, and Q 3 . Introducing these Fig. 54. Readings to Determine Equation values into the equation given, we Curves. have : . $ i H, 1 H 2 H 3 -Q Q, > r- ^3 AQ l -BQu -Cu z = AQ 2 - BQ 2 u - Cv? = -2 gH 2 , AQ 3 - BQ 3 u - Cu* = -2gH,. The unknown constants A, B, C can then be obtained. The values can be determined without the aid of tests, but it is necessary to substitute in place of capacity Q the relative velocity w z ^ Capacity ^ Fig. 57. Method for Finding Characteristics. of water leaving impeller, and in place of u, the revolutions, the circumferential speed V 6 or u 2 , making the equation as follows : A - wz 2 - Bw 2 u 2 - Cu 2 2 = -2gH. and u z are in direct proportion to Q and therefore, referring to Figs. 68 and 69, Fig. 55. Various Forms of Im- pellers for Different Commercial 68 CENTRIFUGAL PUMPING MACHINERY ^^ =- Capacity Fig. 56. Various Forms of Impellers for Different Commercial Problems. f = coefficient of friction along vanes, approximately 0.1; = coefficient of losses due to shocks, approximately 1.2; ' ,. . discharge Capacity velocity; Fv = discharge area; 77 = hydraulic efficiency. SHORTER METHOD OF FINDING CHARACTERISTICS. A shorter method of finding the char- acteristics is as follows : Take H f as the head with the discharge valve closed, which by previous equation is H r = ju ; fj, = varying between 0.9 to 1.1, accord- ing to the proportions of impeller. The ordinary equation for a centrifugal im- peller is I = V - r 1 The direction of the lines changes ac- cording to angle ai (see Figs. 55 and 56) . Assuming that we intend to figure an impeller for condition Z given in Fig. 57, and knowing the head H\, we can plot capacity the curve from the three elements, two Power Characteristics for Dif- points and a tangent. Referring to Figs. 55, 56, and 58, the various forms of im- Fig. 58. ferent Shaped Vanes of Impellers. peller vanes show the different characteristics found in the usual type of commercial pump. In No. 1 the power increases considerably, over the SECOND ANALYSIS OR THEORY 59 limit, and this form is limited to conditions where the load is variable, due to f fictional resistances other than the regular pumping head. No. 2 will produce a constant head for best efficiency, but has the disad- vantage that the power increases rapidly. No. 3, with a constant speed, will maintain a uniform efficiency under a varying head. The extent to which a designer can go in the development of the shape of the vanes depends upon the conditions of service, and should be closely investigated. Fig. 59 shows the curves described. Curves A. Connecting Points of Highest Efficiencies Speed or Revolution Constant. Curves B. Connecting Points by Highest Efficiencies Capacity Constant. Curves C. Connecting Points of Equal Efficiencies. Curves D. Showing Head for Constant Revolutions. Capacity Fig. 59. Characteristics for Curves. CHAPTER XI. GRAPHICAL ILLUSTRATION FOR DETERMINING THE IMPORTANT ANGLES. IT is advisable to recall the relations existing in plane trigonometry, and the measuring of angles in degrees and minutes in circular measure. Lines having the names sines, cosines, tangents, and cotangents bear a fixed relation to each other for any given angle and radius. Fig. 60 gives these ratios. Relation of Angles. a . r, b hence a sine A = - sine B = - c c b a cos A = - cos B = - c c b a tang A = 7 tang B cot A = - cotg B = ^ c sin A b = c sin 5 a = c cos 5 a = 6 tang A 6 = acotgA 6 = atangJ3 a = 6 cotg B Therefore sin A = cos B sin B = cos A tang A = cot B tang B = cot A a = is the sine 6 = is the cosine / = is the tangent m = is the cotangent The general equation as obtained in the first part of the analysis gives y, = V/^-(l- f\h or or tang ao/ 1 tang cotg COtg (XQ cotg /3 cotg = tang, Fig. 61, = 1 _ Z5^ No> L This gives an equation for finding angle CK O . It can also be found by taking the angle between the vertical line and the absolute velocity of water (see Fig. 62) and i, the angle between the vertical line and the rela- tive velocity of water. Then tang tangjS (l - and tang /3 = =F tang a.\ 60 ILLUSTRATION FOR DETERMINING ANGLES 61 Fig. 61. Diagram for Angles. Fig. 62. Diagram for Angles. 62 CENTRIFUGAL PUMPING MACHINERY cti = angle between radial and relative velocity. az = 90- ai. ao = ai + ft + flo = angle between tangential and relative velocity. j8 = angle between radial and absolute velocity. 0o = angle between tangential and absolute velocity. V = velocity of water in discharge. Vi = radial velocity water at outlet in wheel. F 2 = radial velocity water at inlet in wheel or Si. Fs = tangential velocity also used as Uz at outlet. V 6 = tangential velocity also used as HI at inlet. F 7 = velocity of whirl at inlet. Vs = velocity of whirl at outlet. F 9 = relative inlet speed also Wi. Vio = tangential velocity of water. S = absolute outlet speed. Si = absolute inlet speed considered also as F 2 to prevent complication, assuming water enters radially at entrance without shock. The negative sign is used for a\ and when they are on different sides of center line, and the plus sign when on same side (see Fig. 64) . Referring again to Fig. 62, having found the angle a , it remains to find angle if de = absolute speed of water leaving impeller; ef = relative speed of water de leaving impeller. ef V, sin a sin sin (a 180) (see Fig. 65). Vi = radial speed = de sin /3o, and by this we obtain T , Tr sin a sin (an X sin /3 , \ or cotg cotg j3 cotgao' cotg o = TT No. 2. Fig. 63. Diagram for Angles. This equation, together with the one for a , will solve the angle /3 - For simplicity in solving Fig. 63, it is assumed that the water enters the impellers without loss and radially, hence F 2 is equal to s\. The water entering the impeller has a radial velocity 7 2 and inner part of the impeller a circumferential velocity of V G . Let the outer part of the ILLUSTRATION FOR DETERMINING ANGLES 63 impeller have a circumferential velocity F 5 , and let the velocity of whirl at entrance be Vi and at outer circumference F 8 . Plac- ing Vz radially and 7 tangenti- ally a parallelogram is obtained, ab denoting the velocity at entrance, and by making ac equal to the tangential velocity, we obtain be, the velocity of the w r ater relative to the impeller. This relative velocity determines the entrance angle, as the vane must be tangent to it. For the outer diameter, making the velo- city of whirl Fs, the velocity of wheel F 5 , and the radial velocity Fi, the outer relative velocity ef can be ascertained, which deter- mines the direction of the vane at outer circumference. The work done by the water through the impeller is - (F 8 X F 5 - V 7 X F.) y foot pounds per pound of water, Minus Figs. 64 and 65. Velocity Diagram. or assuming a radial entrance . , Vs X F 5 the expression becomes Assuming water on entering is turned from a radial direction into one of rotation, that its velocity is F 2 and the direction of vanes is Vg, the water would enter without shock. After entering the actual velocity be- comes ab. The change of velo- city relative to impeller is from ok to ag, and the loss of head is (ag X aA-) 2 (F 2 X ok - V z ag) Fig. 66. Graphical Diagram for Angles. 2g = Fe Vz cotg at entrance. 2g = loss of head Calling this L = - (F 6 - F 2 cotg c^) 2 (see Fig. 63). 64 CENTRIFUGAL PUMPING MACHINERY GRAPHICAL CHART FOR QUICK REFERENCE. This can be laid out graphically (see Fig. 66), by using the abscissas for tang i and tang /3 as ordinates. Equation No. 1 will give straight lines, beginning at 0, for different values of the tangential velocity F 5 , the efficiency 77^, and the total head H. Equation No. 2 will give parallel lines when less than an angle of 45 degrees with the coordinate axis. Tang. OL =Diff vision Vane Angle 5.5 5 4<5 4 3>5 2.5 1.5 1.4 7.5 2.5 1.5 This Line divided into Equal Parts advancing by -jo Fig. 67. Graphical Chart for Angles. The method to be adopted is to choose a tangential velocity according to the speed and to select a convenient diameter for the impeller, thus obtain- Vi ing a value for 1 gh which will give one of the lines radiating from 0. Furthermore, we have the radial velocity V\ = _ !r ; TT 2 n, W i Wi being width of impeller at the outside, Q being capacity in cubic feet or gallons per second, R = radius of the outside of impeller. When we have T7 obtained values for -^ we will have one of the parallel lines. The intersec- tion between the radiating line and 45-degree line gives the two angles a\ and |8 as shown in Fig. 66. A chart can be made for ready reference in order to determine graphi- cally the diffusion vane and impeller angles as shown in Fig. 67. ILLUSTRATION FOR DETERMINING ANGLES 65 Vi V 2 Parallel lines with a value of ^ and intersecting lines with a value -~ K5 0/1 will give tang a by projecting parallel to A-B, for a by projecting parallel to C-D. >j EXAMPLE. See equation No. 1. Vfa 3504 X 0.80 0ff- 32.2X64 where head 64 feet =H\ F 5 = 58 feet; ij h = 0.80; FI = 65 feet. F 5 58 If the diagram were laid out the line 1.36 would intersect the ^r une 0.112, and the tangent of a would be 6.65 and of i 2.4. Hence a = 81.5 and ai 67.5. Angle 3 = 90 - 67.5 = 22.5 (see Fig. 68), and for the diffusion ring the angle would be 90 - 81.5, or 8.5. METHOD OF CORRECTING IMPELLER-VANE ANGLE. Assuming a pump designed for a head of 170 feet which on being tested gave an actual head obtained on test of 157 feet, how much must the vane angle be increased in order to get the required head? (See Fig. 68.) Fig. 68. Correcting Angle at Tip of Vane. tang aix being the new angle = {H (tang a + tang i) tang a. Assuming a = 74 and i = 63, HJ (3.48 + 1.96) - 3.48 = 5.02 - 3.48 = 1.54. tang aix = 1.54 or 57, or it must be raised the difference between 63 and 57, or 6. CHAPTER XII. THIRD ANALYSIS OR THEORY. THEORY OF IMPELLERS. H = head in feet. Q = capacity in gallons or cubic feet per second. n = number of revolutions per minute. 2 R = outside diameter in feet. 2 r = inside diameter in feet. F 5 = tangential velocity at outer circumference in feet. V 6 = tangential velocity at inner circumference in feet. 77^ = absolute hydraulic efficiency or the ratio between the useful head h and the total pumping head H. g = acceleration, or 32.2. a = diffusion- vane angle at outer circumference of impellers. ai = impeller- vane angle at outer circumference of impellers. 2 = impeller- vane angle at inner circumference of impellers. W = width of impeller at hub. Wi = width of impeller at circumference. A = area of discharge pipe in square feet. AI = area of impeller at outer circumference in square feet. AI = 2 # X TT X TFi. A 2 = area of impeller at inner circumference in square feet. A 2 = 2r X TV X W. A 3 = area at hub of impeller, the latter being C in diameter. A 3 = 2r 2 ^ - C 2 ^in square feet. A 4 = area of suction pipe in square feet. Z = number of diffusion vanes. Zi = number of vanes in impeller at outer circumference. Zi number of vanes in impeller at inner circumference. It is customary to make Zi = 2 Z 2 . / = thickness of diffusion vanes. ji = thickness of impeller vanes at outer circumference. / 2 = thickness of impeller vanes at inner circumference. y = velocity of water in discharge pipe. Vi = radial velocity in section AI of impeller. V 2 = radial velocity in section A 2 of impeller. V 3 = velocity in section A s . V^ = velocity in suction pipe. 66 THIRD ANALYSIS OR THEORY 67 To prevent shocks and losses, an impeller must be so designed that the velocities will increase gradually in going from section A 3 to A\. The most important values required are the inner and outer vane angles, which may be determined by the following equations : F 5 = V v (1 -f- cotg a x or 5 = V C 1 r *?fti V r- -- r, g (1-f cotgatangai) V^ h cotgatangai = - - 1. and H The angle a is usually between 70 and 83 degrees, depending somewhat upon the capacity. The absolute hydraulic efficiency to be selected is between 69 to 80 per cent, depending upon whether the pump is of the volute or diffusion type, and also upon the head, as it governs the diameter of the impeller. The angle a is such that the velocity F 5 is larger than V 2) but care must be taken that the space W\ does not, with small angles, become too small. When the value of a has been calculated we have tang a tang ai = -- --* - - tang a, V, r Example No. 1. Capacity, 25,000,000 gallons per 24 hours; head, 200 feet. Dividing this head into two stages of 100 feet each, we find as follows: Size of suction and discharge: Selecting a water speed in feet per second from 6.6 to 8.2 feet, we will find 2325 2325 * = t0 Square feet > = 4.72 to 5.87 square feet, = 29.4 to 32.8 inches diameter. We therefore select a 30-inch discharge, and as it is desirable to have the suction pipe a size larger a 36-inch suction is chosen in order to cut down the friction and other losses. To find the inside diameter of impeller, or 2 r : The velocity of the water entering the impeller at A 3 is F 3 can be determined by the empirical formula, 7 3 = 0.09 to 0.12 \/2 gH, in which g = 32.2 and H = 100 feet. F 3 becomes 7.23 feet to 9.64 feet per second. Selecting 9.2 feet per second and remembering that V 3 must be larger than 7 4 , we find that cotga X tang! = -* 68 CENTRIFUGAL PUMPING MACHINERY Taking tang a 4.2, we obtain 79.1 2 X 0.83 X 4.2 tangai= 32.2 X 100 - 4 - 2 > or2 - 8 > c^ = tang ' This angle gives the radial velocity V\ of 11.5 feet per second of the water at the circumference of the impeller, which is satisfactory, as it is larger than F 2 and F 3 . Width of impeller at circumference, Wi, and at hub = W. In order to determine this the impeller should be laid out on the drawing board and from the information already obtained the internal passages can be properly determined, and with the selection of the number of vanes the space can be calculated. In laying the impeller out it is well to allow 10 per cent for losses in the passage through it. The angle o^ of the im- peller at inner circumference can be obtained as follows: VK 39 6 tang a: 2 = -7- = ypr^ , or 3.66, where A% = 10.8 feet per second, and Fe = 39.6 feet per second. 2325 The area at A 3 will be ' 9 = 4.2 square feet; adding 0.5 square foot OvJ /N t/.^J for the hub, it becomes 4.7 square feet. The diameter of impeller at the inside is / - , or 2 feet 6 inches diameter. This would give a very slow 4 speed and would require a large wheel, hence it would be best to increase the velocity at the entrance of impeller to 10.8 feet per second, which would 2325 give A 3 = iyr^ = 3.6 square feet, plus 0.5 square foot, or 4.1 square uU X -LU.o feet in all, or a diameter of 2 feet 3J inches. It is recommended that the outside diameter of the impeller 2 R = 4 r, which in this case would give 2R = 4 feet 7 inches diameter, or R = 2 feet 3J inches. The circumferential velocity of the impeller: / 1.40. # c/1.4X 32.2X100 _, This being V 5 = / - 775 ^ = V i 77T^\2 = ' " ^ ee ^ P er second. / -| f* r \ ~ (v-b) 4 / 1 The number of revolutions would then become 60 X V, 60 . 77.6 n= jr-fr = - ^-^~ 325 revolutions per minute. 7T Z K 7T 4. DO Calling this 330 revolutions, the speed F 5 becomes 79.1 feet per second, -17 and Vz = - = 39.6 feet per second. THIRD ANALYSIS OR THEORY 69 The angles of vanes may be calculated as described on pages 63 and 64, [Chapter XI. a = diffusion- vane angle; i = impeller- vane angle; rjh = absolute hydraulic efficiency, assumed at 83 per cent. The method thus described, with the velocity diagram, Chapter XI, will give the principal details governing the calculation of a turbine pump. I The shape of the vanes must be a continuous curve. In order to start the pump against a full head the velocity of impeller F 5 must be larger than 1.35 or 1.4 X 32.2 H i-^Y APPLICATION OF ANALYSIS TO PROBLEM. Example No. 2. 8-inch 3-stage turbine pump. Conditions. 1800 gallons or 241 cubic feet per minute; 250 feet head, or 83.3 feet per stage; 800 revolutions per minute. Volumetric efficiency, 95 per cent; total, 65 per cent. Hydraulic efficiency, f f = 68 per cent. Diameters of impellers: hub, 2^ inches, assumed; inside diameter, 2 inches, assumed; outside diameter, 2 R = 18f inches, assumed. The outside diameter gives a speed of 65 feet per second. uU Speed required to start pump against full head, /lAxgXH /1.4X32.2X83.3 F 5 > -- = - = 64.3 feet per second. Velocity of the water in 8-inch discharge, 241 V = - . \ / o \2 = 11-5 feet per second. 60Xm(r7r " Assuming a suction pipe the same diameter as the discharge, the speed with 95 per cent volumetric efficiency would be F 4 = ^-^ = 12.1 feet per u.yo second. This is high, hence it would be well to select a 10-inch suction. A pump designed upon the basis of the given piping would probably not give over 65 per cent efficiency. Calculating the angles a and i would further- more show that an 8-inch diameter would not be as suitable for these con- 70 CENTRIFUGAL PUMPING MACHINERY ditions, as a 10-inch pump, and that the speed should be increased to about 1000 revolutions in order to get good results. Analyzing the angles, a and ai have to be taken so that they will make Vi > F 3 . We have, therefore, 241. 12 2 F 3 = - - - = 20.5 feet per second. 60^(6.5 2 -2.5 2 ) Taking tang a = 3, we have tang ai = tang a\ ^~ 1 L gu. j _ 65* X 0.68 3X 32.2X83.3~M and Fi = - - = = 20.3 feet per second. tang a X tang i 3 + 0.21 These angles, a = 72 and i = 12, are too large for good results. Example No. 3. 4-inch 12-stage. Conditions. 150 gallons or 20 cubic feet per minute; 610 feet head, or 50.9 feet per stage; 1460 revolutions per minute. Efficiency, volumetric = 90 per cent. Total efficiency = 56 per cent. Hydraulic efficiency = 62 per cent. Diameter of impellers: hub, 2 inches; 2 r = 3| inches inside; 2 R = 9 inches outside. Speed for 9-inch impeller = F 5 = - ~^o -- =57 feet per second. Speed to start under full head, T , . / 1.4X32.2X50.9 K0 , , F 5 > / - /0 Q>7K > 2 = 52 feet per second. /3.875\ 2 ( 9 / Speed of water in suction and discharge, assuming the same diameters, 20 1 F = - - = 3.83 feet per second, 3 83 F 4 = 7 ;r = 4.26 feet per second. u.y It would be advisable to make suction pipe 5 inches diameter in order to reduce velocity. Velocity F 3 = - - = 5.11 feet per second. 60 (3.S75 2 - 2 2 ) THIRD ANALYSIS OR THEORY 71 F 3 , as it should be. Angle 2 , speed 7 2 = 5 feet. 2x11 = 5 * 3 875 speed F 6 = 57 X - = 24.6 feet per second. y 24.6 tang 2 = F~ = 4.92. o Assuming 8 diffusion vanes and 12 impeller vanes at outlet, with thick- ness of diffusion vanes H m ch an d impeller vanes /? inch, we have 20 ( /0.6875inch\ /0.28 12 ^^ V 19 - -- o.o J2.36 - 8 X 0.057 - 12 X 0.023J, TFx = 0.69 -*- J2.36 - 0.456 - 0.276], Wi = J> = 0.42 inch. l.oo lowing 10 per cent for losses, the width W\ becomes 0.46 inch, or about T B- inch. Inside width W, therefore, may be similarly calculated, assuming 6 vanes at entrance J mch thick. W = 5p r y -5- f 3.14 X 0.32 - 6 X 0.02}, ft 7QQ W = 0.799 -HI- 0.12} = 7 ^ = 0.896 inch. U.oo Allowing for 10 per cent, W = 0.985, or about 1 mch. It should be noted that hydraulic efficiencies of 62 to 79 per cent should be used for volute type of pumps, about 62 to 70 per cent for small turbine pumps, and 70 to 85 per cent for large ones, the value varying with the type and design of the particular pumps considered. A method which may be applied in a somewhat different manner is given below. Radial velocity at outer circumference, V\\ Radial velocity at inner circumference, F 2 ; Wi = width at outer circumference; W = width at inner circumference; /i = thickness of vane at outer circumference; /2 = thickness of vane at inner circumference; k = coefficient of contraction = 0.90. 72 CENTRIFUGAL PUMPING MACHINERY Q / = ej sin 0:3 = -7^. ^^ (27TJKTF ! - (2irrW-nWcoseca 4 f 2 )k See Fig. 62. The angle a 3 varies in practice from 15 to 30 degrees. The width W\ and W can be calculated from the following formulae: Q cos a cos ai Q Z = number of diffusion vanes. Zi = number of impeller vane's at outlet. Z 2 = number of impeller vanes at inlet. These values of W should be increased about 10 per cent for interna leakage and losses. Volumes should be calculated in cubic feet per second and dimensions in feet. CHAPTER XIII. SCREW OR PROPELLER PUMPS. THE latest development of the centrifugal pump is its use combined as a unit with the steam turbine for circulating water at very low heads, and in moving large bodies of water under low velocities and at comparatively low heads. In this field, high speed causes the designer special difficulties in determining the proper inlet and outlet diameters of the impellers, and "the length of the vanes; and frequently the path of the water in the impeller becomes too short. For such work the moving of the water in an axial direction is necessary, and a screw or propeller is particularly adapted to it. Screw pumps belong to the same class of velocity pumps as centrifugal. They consist of guide vanes at the inlet, with screw or propeller and guide vanes at the outlet. All are mounted in a cylinder or chamber. The velocity of the water, being parallel or axial, lends itself readily to operation at high speeds, with high efficiency for water at low heads. By arranging the screws or propellers in opposition to one another, all end or lateral thrust on the shaft is eliminated, and at the same time the capacity can be doubled without increase of speed. The angles of the stationary inlet and outlet vanes and blades of the screws are so arranged that the water enters axially and is given a radial motion, finally being discharged parallel to the axis in the vortex or volute chamber, thereby utilizing both the impulse and reactional forces. In a centrifugal pump the liquid flows in a radial direction through the impeller and obtains its energy through the difference in circumferential speed at the inlet and outlet of the impeller. In a screw-propeller pump the liquid flows axially, there is no increase in circumferential speed, hence the liquid must obtain its pressure through some other cause. Fig. 69. Path of Water. Suppose, first, that a screw propeller revolves in the water so that no shock occurs either at the inlet or outlet. The water flows as shown in diagrams at A and B (see Fig. 69). 73 74 CENTRIFUGAL PUMPING MACHINERY u-i, circumferential speed at A, is equal to u z , circumferential speed at B, v roJ absolute inlet speed, equal to tv a , absolute outlet speed. w\ } relative inlet speed, equal to Wz, relative outlet speed. The diagrams at A and at B are similar. The liquid, therefore, has the same energy at A and B. During the flow of the liquid from A to B no energy has been put into the liquid and no pumping head can be produced. ' A smooth, shockless flow through a screw propeller does not produce any pump- ing head. Slip. A certain shock must, therefore, be produced in order to obtain pressure. The full theoretical capacity cannot pass through the propeller. This reduction in capacity is called slip. Instead of letting the water enter as per diagram uv ro w, we let it enter as per diagram ucw f (see Fig. 70) . 1C ' Diagram with Slip Fig. 70. Entrance Angles for Water. Definition of Slip. Qo = capacity without slip, or the theoretical capacity. Q = capacity with slip, or the actual capacity. z = slip. ^~ = 1 z = ratio between obtained capacity and maximum capacity. Section through Propeller Development of Vanes at Diameter D Fig. 71. Guide Vanes. Assuming radial guide vanes we find as in Fig. 71, Qo = A v ro = A - u tang 5. A = radial area = [D 2 2 5 = pitch angle. z = number of vanes. u = circumferential speed = _ D 7r r.p.m. 60 SCREW OR PROPELLER PUMPS 75 Assuming guide vanes entering at an angle e in direction of rotation (see Fig. 72), Qo'= theoretical maximum capacity; tang e/ Fig. 72. Guide Vanes. Assuming guide vanes entering at an angle e, pointing opposite to direc- tion of rotation, . . Qo" = theoretical maximum capacity; For Qo' the only difference in the equations is in the values of tang e, which are positive and negative respectively for $ Qo'. If the inclination of the guide vanes is in the direction of rotation, the maximum capacity for slipless flow is smaller than for radial guide vanes. For guide vanes pointing against the direction of rotation, the maximum capacity is larger than for radial guide vanes: Qo' < Qo < Qo". 2 = 1-^- fore = 90, z = l- 90, < QQ'J Qo" being figured by the above equations. Law of Proportionality and Slip. For water turbines and centrifugal pumps a law exists which can be expressed by the formula 76 CENTRIFUGAL PUMPING MACHINERY where N = revolutions per minute; Q = capacity; H = total head; H.P. = horse power. This law implies that the capacity changes in direct proportion to the speed, the head changes in proportion to the square of the speed, and the horse power changes approximately as the cube of the speed. This law has been found to be correct within practical limits. Propeller-pump tests show the interesting fact that this law of propor- tionality can also be applied to ship propellers. For low speeds the tests show conformity in regard to speed, capacity, and head; for the higher speeds, however, and for the horse powers, some irregularities occur. The law of proportionality can be shown by the curves Q Q -^ = constant and ^ = constant. See Fig. 73. \ SlipC Q LU Constant Slip Curv 'for Maximum Efficiency These | Curves are made from Test Headings, Two Propell rs working Profiler Dimen'sior 1000 2000 3000 4000 Gallons per Minute Fig. 73. Performance Curves. It can be proved that if the law of proportionality is correct then the curves -^ = constant and -4=- = constant represent curves of constant slip. J\ \ H If ^ = constant, jj = constant = law of proportionality. 9. N constant, then z = constant. SCREW OR PROPELLER PUMPS 77 This result is of great importance, as it shows that points of same slip must be points of similar efficiency. For changes of speed the best efficiency will occur at a constant slip. For differently designed propellers the best effi- ciency will occur at dif- ferent slips. For a pro- peller of dimensions given in Fig. 73, the best ?4 er ^ without Q efficiency, 66 per cent occurred at 1800 r.p.m. with 37 per cent slip. Fig. 74 shows the position of vanes, with free space or clearance and no overlap. ^^-- Characteristic Curve. Since slip is necessary in order to produce pres- Total Number of Vanes, Three Fig. 75. Step Diagram. sure, the relation between slip and pressure must be found. This theory is based upon the assumption that all the shock produced [>y the slip is transformed into pressure, and that the pressure can be igured from this shock, provided a proper co- efficient is applied. On this basis we have the formula 120 110 100 90 80 a S 50 40 30 20 10 i \ \ **. \ 7 ~ P- \ \ / 1 \ / 1 \ / \ x \ "**/ N^ \ ^^^s A=i B=c 68t CU irveflj -ve for uredf 1800 B om Eq P.M oatioii \ \ ^ S N \ B H = * 1? -o- siuo) X \ ,-i urve fi T Coeft^ cient \ s^ \ ^ \ D,- 6*" 2= 11M" 5= 13 ^ ^^ L. > 1000 20UU 3000 4000 5000 Q 601 Gallons per Minute x o z = slip; u = average circum- ferential speed; g = acceleration of so gravity; a = absolute angle formed by the direction of the inflowing liquid with the moving propeller vane; y = coefficient. ^ T^- rjr Fig. 76. Application of Equation Described. While this equation does not give reliable results, it can, if properly applied, be used to give an approximation. Fig. 76 shows how this equation can be used. * Hollander's formula. 78 CENTRIFUGAL PUMPING MACHINERY The coefficient y stays nearly constant for a large portion of the curve; for the lower heads, however, it changes suddenly. It changes for various propellers and working conditions, and its value must therefore be deter- mined by tests in each particular case. Fig. 77. General Arrangement of a Screw Pump. Fig. 78. Section of a Screw Pump. A propeller pump is subject to the same general law of proportionality as water turbines and centrifugal pumps. A smooth water flow does not produce pressure : shock is a necessity. A high-speed propeller pump can, therefore, never have the same efficiency as a properly designed high-speed centrifugal pump. SCREW OR PROPELLER PUMPS 79 The general arrangement of a screw pump is shown in Fig. 77, which gives 'approximate outside dimensions, and in Fig. 78, showing the pump in section. Fig. 79 is a section showing the combination of screw and volute centrifugal pump, adapted for a very low head. The inlet is trumpet- shaped to receive the current of water with the least loss. The impeller is Fig. 79. Section Showing the Combination of Screw and Volute Centrifugal Pump. j formed by helicoidal vanes generated from the lowest part of the impeller ] and pitched backwards like a screw. From this form they gradually : change into the regular shape of volute impeller vanes. The discharge * passages can form a regular volute casing when water is to be delivered , into a pipe or conduit. For canal work the open discharge casing is suit- ! able in connection with the necessary sluice valves in the canal. PART III. APPLICATIONS AND USES. CHAPTER XIV. GENERAL REMARKS. WATERWORKS INSTALLATION. OWING to its characteristics the centrifugal pump is better adapted to some engineering problems than to others. Its simplicity of construction, Fig. 80. Vertical, Self-contained, Two-million Gallon Turbine Pump. wide passageways, absence of valves, low first cost, comparatively light weight for its capacity and its adaptability for motordrives have given this type of pump a constantly widening field of application during the last fifteen years. 81 82 CENTRIFUGAL PUMPING MACHINERY GENERAL REMARKS 83 In this and following sections are described the more general and impor- tant :r^~l: cut ions of the centrifugal pump up to the present time. WATERWORKS. Turbine pumps are now supplying water to municipalities like Buffalo, Lockport, Toronto, Montreal, Louisville, Minneapolis, and others, and furnish a constant and uninterrupted supply. The sizes range from small plants of a million gallons per day to those of twenty-five to thirty mil- lion gallons per day. In applying centrifugal pumps to service of this sort, the best practice calls for delivery to the top of standpipes or reservoirs. In this way both the capacity and pressure can be kept constant. Fig. 82. Twelve-inch Three-stage Pump. Fig. 80 illustrates a vertical, self-contained, two-million-gallon turbine pump installed at Athens, Ga., a type which takes up little floor space, and which has proved very satisfactory for small waterworks and manufactur- ing plants. Figs. 81 and 82 illustrate the three 12-inch three-stage pumps of the waterworks plant of the city of Lockport, N. Y. Fig. 83 is the test curves of the motor. The pumps are directly connected to 500-horse-power motors and are of the horizontal-shaft three-stage turbine type, with a single suction. The suction openings are 14 inches in diameter and fitted with special vapor openings. The main casing is cast iron of a tensile strength of 30,000 pounds, annular in form, and fitted with suitable supports to attach to base. The suction-head casting is of special design to facilitate the re- moval of internal parts without dismantling the pumps. The impeller is special bronze, of the inclosed type, and is arranged with balancing cham- 84 CENTRIFUGAL PUMPING MACHINERY bers, which reduce the end thrust to a minimum. All impellers are per- fectly balanced individually and when mounted together on the shaft. The discharge from each impeller is conducted through a set of guides or diffusion vanes, designed to transform the velocity into pressure with the least possible loss. These diffusion vanes are removable. The shafts are of nickel steel, ground and polished, and run in ring oil babbitted bearings. Stuffing boxes are fitted with water seals, consisting of lan- tern glands connected up with suitable piping, obviating all air leaks. All nuts are case- hardened, and the backs of the flanges spot-faced. The installation is a good sample of a modern turbine waterworks installation. Each pump has a capacity of 5,000,000 gallons of water in 24 hours, when operating at speeds given in the test. The motors were intended to run at 720 r.p.m., but actually run at 745 r.p.m. The pumps deliver, therefore, considerably more than the contract requirement, and maintain their efficiency. They were guaranteed to give 68 per cent efficiency, but ac- tually give 82.5 per cent. Water consumption in the city of Lockport is between 400,000 and 500,000 gallons per day and the water is delivered through 68,500 feet of 30-inch steel pipes from the Niagara River at Tonawanda, into a standpipe of limited capacity near Lockport. The test was made with a constant level in the standpipe at Lockport, by throttling the discharge valves on the pumps until the Venturi meters showed that the pump was delivering at a five-million-gallon rate. The speed was 745 r.p.m. The gate valves were then opened and records taken for ratings of five and a half million, six million, seven million, and seven and three-quarter million gallons. The following log of these tests is of interest because it shows he extraor- dinarily high efficiency and capacity of comparatively small pumps for waterworks service. A record was also kept of the pressures in the force main, the rate of pumping resulting, and the horse power of the motor as 500 1000 1500 2000 2000 3000 3500 4000 4000 5000 5500 Torque Fig. 83. Performance Curves of Motor Driving Pump, shown in Figure 83. GENERAL REMARKS 85 indicated by the wattmeter on the gauge board. A com- plete record of the test is shown in the table. Repeated tests on turbine pumps after years of serv- ice have shown that, when rightly constructed, there is little or no falling off in efficiency, a condition hardly met with in recipro- cating pumps, where wear of valves is always present. In considering a waterworks or mill installation, original cost and maintenance should both be taken into account, as the turbine pump has many advantages over the plunger pump in lubrication, repairs, or replacement of operating parts, together with first cost, foundation, and building. They should be designed for a fairly wide range of discharge, and usually for a constant head. Over-all efficiencies of from 68 to 72 per cent can be obtained as against 75 to 80 per cent for the reciprocating pumping engine. The latter still has an advantage in the cost of power, which, however, is offset by the lower cost of maintenance of turbine pumps, of their buildings, foundations, interest on in- vestment, depreciation, and repairs. SI ^y* r . 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Tt< to 10 CO t^ t>- ;00 96 CENTRIFUGAL PUMPING MACHINERY Fig. 87 shows the Dry Irrigating pumps, which were installed by the Madras Government. They comprise eight 39-inch volute pumps de- j signed to run at 180 revolutions per minute, direct-connected to vertical Diesel engines. Each pump will pump 26,000 gallons per minute against a head of 12 feet. Fig. 88 shows the type of irrigating pumps in use in Colorado and places where water has to be elevated 100 feet or more. There are a number of these installations in operation for the Redlands Irrigating Company, the Orchard Construction Company, and several other companies which are i rendering productive land which was formerly considered of little value. The United States Government has introduced the policy of reclaiming land in the arid and semiarid regions that are located between the Missouri River and Rocky Mountains, and about 8000 canals and 3500 miles of ditches had been built up to 1909. The most important work has been Fig. 87. Dry Irrigating Pump at Madras. done in the North Platte district, which includes a part of Wyoming and Nebraska; the Shoshone district in Wyoming, with an elevation of 4000 to 5000 feet; the Salt River district in Arizona; the Huntley district of Montana, with an elevation of 3000 feet; the Wittestone district of North Dakota, where water is pumped from the Missouri River, and where the banks of the Missouri are so exposed to changes that floating pumping stations were made necessary; the Buford district, North Dakota; the Garden City district of Kansas, which uses an underground supply for irrigation. In all these stations, the pumping equipments are either of the motor-driven centrifugal or of the gas-engine-driven centrifugal type. When a pumping station is required for occasional use only, the equip- ment should be simple, and economy is of secondary importance. For draining or irrigating large areas of land, where pumping is to be carried on uninterruptedly for months during the wet or dry season, there should be an economical installation, with carefully designed collecting canals, piping and outlets, so that water is not pumped even a few inches higher than necessary. IRRIGATION, DRAINAGE, AND SEWAGE 97 To secure an economical design, the pipes should be arranged to form a siphon, with the pump on the top of the siphon. The ends of the suction and delivery pipe form the ends of the siphon, which are determined by the lowest inner and outer levels reached under any working conditions. Suc- tion pipes should be short, and plenty of room should be allowed around them so as to prevent currents or vortexes. Water should enter the suc- tion pipe quietly and equally around the entire rim. The minimum diam- Fig. 88. Type of Pump used in Colorado and Places where Water is to be elevated 100 feet or more. eter of the suction well should be four times that of the pipe, and when properly designed the water may be pumped to within a foot of the rim of the pipe. The bottom or rim of the suction pipe should be of the bell form, known as vena contracta, which gives good results for low lifts. It is also desirable to have the suction pipe gradually enlarge from the pump, so that the high velocities will be reached gradually. No screen should be fitted on the suction pipe itself, but one should be placed across the canal or well to prevent foreign matter from entering the pump. 98 CENTRIFUGAL PUMPING MACHINERY The delivery pipes should be of gradually expanding section, so that the velocities at the point of discharge are reduced to from 4 to 2 feet per second, and should discharge into a still pool. This change of diameter in the pipe should take place in not less than 8 to 12 diameter lengths, other- wise the velocities will be changed too abruptly. The remarks about irrigating and drainage pumps and the description of types apply in a great measure to pumps for sewage, of which there are many notable installations in such cities as Boston, Chicago, New Orleans, and Milwaukee, which give excellent results and good economy. Fig. 89. Arrangement of Horizontal Sewage Pump installed by the Belfast Cor- poration, England, at the Greencastle Pumping Station. Fig. 89 shows the arrangement of horizontal sewage pumps installed by the Belfast Corporation at the Greencastle Pumping Station, Belfast, Eng- land. In this case most of the pumping head is on the suction side, the pumps being automatically controlled by the suction level. Where vertical pumps are necessary, the type shown in Fig. 90 is used, which allows natural priming. Care should be taken in the design of pumps for unscreened sewage, as special impellers and large casings are necessary, and storm relief pumps must be so designed that they can handle considerable sand and grit. IRRIGATION, DRAINAGE, AND SEWAGE 99 Fig. 90. Arrangement of Vertical Pumps at the Greencastle Pumping Station, England. CHAPTER XVI. HYDRAULIC MINING AND DREDGING. THE removal of earth in large quantities can be accomplished more quickly and cheaply by means of centrifugal pumps than with steam shovels. This is being done on the Pacific Coast in various places, both for mining and grading, and also at the phosphate mines in Florida. To remove the loosened material special dredging pumps are usedj which are fitted with manganese impellers. The casings may be lined or- unlined, depending upon the size of the pump and the nature of the work to be done. Such pumps are capable of handling a large amount of sus- pended solid matter, even up to 25 per cent of the water, and are designed so as to take care of stones 4 to 12 inches in diameter. Fig. 91. 10-inch Ten-stage Hydraulic Mining Pump. Fig. 91 shows one of the four 10-inch ten-stage pumps installed in Seattle for grading Jackson Street, used in removing two million cubic yards of earth. The Panama Canal Commission have installed centrifugal pumping engines to remove 7,500,000 cubic yards of material, consisting of soft silt, earth, clay, and hard rock. Various attempts were made with dredges, but the final conclusion was to handle it by hydraulic methods. The material to be loosened and moved by the dredging pumps is dark loam, containing 15 per cent sand and gravel, and weighs 75 pounds per cubic foot. Centrifugal pumps are frequently used for dredging and pumping mud from rivers and harbors. The cost of operating centrifugal dredging pumps compares favorably with ladder or bucket dredges, and in some cases is more economical. A modification of this pumping of suspended 100 HYDRAULIC MINING AND DREDGING 101 material is found in wrecking work, where hulls are cleared of wheat and [other substances which have no cohesion. Centrifugal dredgers have been employed extensively by the United States Government for deepening rivers and harbors, and filling in and reclaiming land by pumping the dredged material through piping. Such work requires velocities in the pipes of about 10 feet for about 3 to 5 per cent of solid material, which* consists of rock, clay, sand, shells, and mud. Discharge piping is made of either wood or metal, wood giving the better results in regard to wear and tear. Fig. 92. Horizontal 20-inch Wilredging Pump. Horizontal Steel Lincel Type. Fig. 92 shows the large dredging pumps installed at Panama by the Canal Commission. There are four of the horizontal-shaft type, each with a sin- gle 20-inch side suction and a 20-inch delivery, and are exceedingly heavy in construction. The runners or impellers are of manganese steel and of the inclosed type. The entire internal surface of the casing is covered with soft boiler plate, and the casings are of extra thickness to stand the heavy work. The suction openings are protected by removable cast-steel throats or rings. This construction makes one of the best arrangements for easy repair. The shafts are nickel steel. Each pump is designed to deliver 10,000 gallons of water per minute, against a height of 60 feet, exclusive of pipe losses and suction heads of 10 feet through 1200 feet of pipe line. A suitable thrust bearing of marine type is provided to take care of the unbalanced pressure. 102 CENTRIFUGAL PUMPING MACHINERY Each is capable of excavating and disposing of 300 cubic yards of solid matter per hour, using 10,000 gallons of water per minute, or about 10 cubic feet of water per cubic foot of material, which is loam with 15 per cent sand. With 3000 horse power available, the efficiency will reach approximately 60 per cent. On the basis of 10 cubic feet of water per cubic foot of material, and estimating the 85 per cent of dirt at 80 pounds per cubic foot, and the 15 per cent of sand at 110 pounds per cubic foot, we obtain 68 pounds and 16J pounds respectively, or a total of 84 J pounds per cubic foot. Assuming a mixture of 40 per cent clay at 120 pounds per cubic foot, giving 40 pounds, 15 per cent sand at 16J pounds, and 45 per cent loam at 45 pounds, we have an average weight of 101| pounds for the material. Taking the weight at 110 pounds per cubic foot, and the volume at 300 cubic yards per hour, the material equals 891,000 pounds per hour, or 14,850 pounds per minute, plus 10,000 gallons or 84,042 pounds of water, making a total of 98,892 pounds of material to be pumped per minute by three pumps; the fourth being a " booster " pump. The total of 98,892 pounds divided by 10,000 gallons will give 9 T 9 o pounds of weight per gallon, or 74J pounds per cubic foot of mixture. This gives 10 per cent of solid matter by volume and 15 per cent by weight. Some of the mixtures to be carried weigh 90 pounds per cubic foot, which equals a weight of 12 pounds per gallon, or a total of 120,000 pounds per 10,000 gallons. Deducting the 84,042 pounds of water leaves 25,958 pounds of solid material, or about 21 per cent by weight. These figures give an idea of the hard service under which these pumps operate. Each is driven by a 655-horse-power motor of the two-bearing type, running at 480 revolutions and fitted with autostarters and autotrans- formers. There are also four motors operating the vacuum pumps of the priming equipment. These vacuum pumps are of the single horizontal type operated by a 5-horse-power motor. There is also a complete set of transformers and marine switchboard for starting automatically, fitted with the necessary wattmeters, oil switches, and overload trip arrange- ment. There are all the accessories for starting the motor on the compen- sator and cutting it out when the switches are in running position. The current for operating the oil switches is 25 cycles, stepped down from 2080 volts to 110 volts. The motor connections are brought to a box near the base at the side of the motor. All of the leads are insulated and bushed to insure safety in case of accidental flooding. The six hydraulic giants or monitors are of the latest type used in mining operations in California and other parts of the West. Each weighs 1500 pounds and consists of a base for attachment to a 16-inch gate valve at the terminus of the pipe line, a horizontal and vertical joint, and a long, conical reducing piece. The frictional resistance is decreased by a ball bearing, and a weighted lever is attached to control the direction of the HYDRAULIC MINING AND DREDGING 103 jet. A deflecting nozzle is fitted to the discharge end of the giant, which permits deflections through a small angle without changing the position of the main body. The tapering piece of the giant is fitted on the inner side with two sets of guide vanes which prevent a scattering or rotary motion of the water after it has issued from the nozzle. The nozzles used vary from 4 to 6 inches in diameter, according to the character of the material in which they are working, and at full head the water coming through them exerts a pressure of 130 pounds to the square inch, the equivalent of a ton and one-half of pressure against a bank 100 feet away within range of the deflectors. As it is expected that the positions of the monitors will be shifted frequently, their bases are of temporary construction. The areas excavated and filled by the giants are 8 feet above mean tide and the average depth to be excavated is 45 feet. This is accomplished by wash- ing down the material in sluices which carry the water containing earth in suspension to the sump where the barge or dredging pumps are at work. When it becomes necessary to move a barge, the giants cut a new sump with a channel leading into it through which one or more units of the fleet is floated. The banks are excavated as nearly perpendicularly as possible, in order that benches may be cut in them and the banks undermined so as to cause the material to fall by gravity. CHAPTER XVII. MINING WORK. THE turbine pump has been extensively used for removing water from mines, for station or drainage service, and for sinking. Fig. 93 shows an 8-inch eight-stage pump for a lift of 1250 feet. Sev- eral of these pumps are in successful operation handling large quantities of Fig. 93. 8-inch Eight-stage Pump. Deep Mining Pumps. water, each impeller operating against as high as 157 feet, the efficiency being 75 per cent. In a turbine type of pump it is practicable to have single lifts as high as 1600 to 1800 feet. In reciprocating pumps the lifts rarely exceed 1000 to 1200 feet. Several large pumps handling 5000 to 6000 gallons per minute, against 500 feet lift, have been in successful operation. For acid water the working parts are made of special acid-resisting bronze, containing copper, tin, and lead. The use of electric sinking pumps is becoming general, particularly for emptying mines that are flooded. In sinking work a simple suction pump is provided, so that the suction pipe cannot be uncovered. This allows the pump to operate continuously with an ample supply of water. Figs. 94 and 95 illustrate sinking pumps for deep working. The use of this type eliminates all steam and exhaust pipes. It is to be noted that a sinking pump should be proportional for the final head against which it has to pump, and special precautions must, therefore, be exercised in its design. The pumps are either self-supporting or hung in frames. This freedom from shocks, particularly in freely suspended pumps, allows high velocities of water, thereby reducing the size of pipe column. 104 MINING WORK 105 By means of specially designed gate valves in the discharge, the quantity of water delivered can be regulated to the flow into the pump, and by a special design of suction chambers the accumulation of air can be prevented. IMC Figs. 94 and 95. Sinking Pumps for Deep Working. The advantage of a centrifugal sinking pump lies in its compactness and it weight, with the consequent facility in handling and transporting in shafts. 106 CENTRIFUGAL PUMPING MACHINERY In introducing the turbine pump for mining purposes, considerable prej- udice will have to be overcome, on account of the unsatisfactory results which have been due to wrong selection of pumps, and to the fact that manufacturers have not studied the requirements properly. The voltage may vary from outside causes, and a motor-driven turbine pump should be so installed that it will be reliable under all conditions. If the speed of the motor and pump under the fluctuations of the line voltage and variation of capacity, and also the power of the motor against a constant head, be carefully considered, and the pump designed accordingly, there is no reason why it should not be used satisfactorily. The efficiencies obtainable are dependent upon the capacities for the heads under which the pump is to be installed, and attention should be paid to that point if good results are to be expected. With the proper relation between capacity and head, efficiencies equal to reciprocating power pumps may be expected, and with provisions for taking care of the fluctuations in the line, centrifugal pumps should meet with no difficulties. CHAPTER XVIII. POWER-STATION WORK. Qjfi^ BOILER FEEDING. IT is only lately that this promising field for the multistage turbine pump has been invaded. The turbine pump is well adapted to medium and large power-station work, where it will furnish an uninterrupted and( continuous flow! of water, free from shocks and water hammers, obviates the otherwise necessary (air chambers, relief valves,) and has a distinct advantage over the reciprocating pump in point of 'lubrication and attendance.' The (life of both 1 piping and pumps Vill be prolonged, -and the plant will be increased in efficiency." It is advisable in large power stations, however, to have a stand-by pump of the reciprocating type to be used in connection with the turbine pumps. There is no possibility of building up undue pressure by closing the discharge or feed valves. No trouble will come from steam- pressure regulators and safety valves if properly installed, but judgment must be used in relation to the drop in steam pressure, to low water, and peak loads; and a reciprocating pump must be kept for emergency and to take care of variations in capacity and irregularities. f Centrifugal pumps,) if motor-driven under constant speed, (should be designed for a considerable Jrange of delivery that increase of pressure will be taken care of, with corresponding capacity. In a steam-turbine-driven pump this can be accomplished by increased speed. A motor-driven pump can also be arranged for variation in speed)if desired. There is no fear of accident with a turbine pump, whether motor- or steam-turbine-driven. Quite a number of these pumps are installed in large power stations like the Commonwealth station at Chicago, the New York Central power station at Xew York, the Interboro power station, and others, and abroad in the Edison Milano Univeraria Valdarno, in Turin, Milan, Florence, Birming- ham, Amsterdam, and other places. Fig. 96 illustrates a steam-turbine boiler-feed pump for 250 pounds pressure. Fig. 97 shows a steam-turbine-driven pump for 350 pounds pressure. Fig. 98 illustrates a motor-driven boiler-feed pump for 250 pounds pressure. In boiler-feed pumps it is particularly/desirable to have perfectly balanced conditions and free, uninterrupted flow to impellers, with velocities as low 107 108 CENTRIFUGAL PUMPING MACHINERY Fig. 96. Steam-turbine Boiler-feed Pump. Fig. 97. Steam-turbine-driven Boiler-feed Pump. Fig. 98. Motor-driven Boiler-feed Pump. POWER-STATION WORK 109 as possible. J An excellent design is shown in Fig. 99, where this is accom- plished by admitting the water simultaneously to two impellers turned back to back, which discharge it into a central double-entrance impeller. Good engineering requires the use of the simplest, most durable, and most economical apparatus. Since this type of pump has been introduced there is a tendency to do away with cumbersome and costly power pumps and steam reciprocating pumps, as the turbine pumps now installed show ! 10*-- 1 Fig. 99. Plan of Boiler-feed Pump with Opposed Balanced Impellers. remarkable results for high speed, low cost of maintenance, reliability and smoothness of operation, economy in space, low cost of installation, and easy attendance. CIRCULATION OF WATER. In order to obtain circulating water for condenser plants, it is usual to employ a centrifugal pump on account of its high efficiency under the con- ditions which call for large volumes of water under low heads. Several designs are used, either engine-, motor-, or steam-turbine-driven, having a single or double suction, depending upon the conditions. Condenser overloads should be taken into account, and the capacity of the pumps should be great enough to take care of the extra work. The connections to the pump from the condenser should be as short as possible, with no air pockets. Means should be devised for priming the pump, or a small connection may be made to the vacuum pump. In engine-driven pumps the regular type of volute design of pump cham- ber and impeller can be easily applied, as the speed is normal. The siphon arrangement as used hi irrigating work is most desirable. By specially 110 CENTRIFUGAL PUMPING MACHINERY designed high-speed impellers this type can be used with motor or steam turbine drives within certain limits (see Fig. 100) . But to meet the higher speeds of steam turbines and to handle large quantities of water under small heads, usually only friction heads, a new birotor or trirotor pump has been placed on the market. These pumps, described in Chapter XXIX, can be made in sizes and speeds to run with the average steam turbine. In large condenser installations where all the auxiliaries are to be driven by steam turbines, a new type of centrifugal vacuum pump of the jet style Fig. 100. Steam-turbine-driven Pump with High-speed Impellers. has been developed. Several of these are being built, arranged on the same shaft with the circulating and hot-well pumps, making a very com- pact unit, the three auxiliaries being driven by one steam turbine. HOT-WELL PUMPS. Closely allied to the boiler-feed pump is the hot-well pump, designed for both steam-turbine and motor drive. It is rarely made more than two- stage, and has an additional vapor opening on the suction entrance. These pumps withdraw the water of condensation from the condenser. They may be either of the vertical or horizontal type, as most suitable to the particular installation. Fig. 101 shows a pump inclosed in the hot well, making a very compact arrangement. Fig. 102 shows the regular arrangement. The turbine hot-well pump has practically replaced the reciprocating type in all modern power stations, and has the advantage of being auto- matic in action, taking care of all the variations in the loads without any POWER-STATION WORK 111 attention, and is always in condition to operate without priming. These pumps are direct-connected to either a steam turbine or motor. It is usual Fig. 101. Hot Well Pump Inclosed in Special Well. to locate them three or four feet below the water line in condenser in order to obtain a good head on the pump. A type of the vertical pump installed Fig. 102. Two-stage Hot Well Pump. aboard ship is shown in Fig. 103, illustrating the compactness and lightness required in such installations. 112 CENTRIFUGAL PUMPING MACHINERY CENTRIFUGAL- JET CONDENSERS. The development of steam turbines requiring high vacuum has obliged the designers of centrifugal pumps to provide apparatus which will give a vacuum within one-half inch of the absolute vacuum. Barometric con- densers have been used more or less for this work, but, owing to the ex- tremely long piping with consequent air leaks, they have given way to the centrifugal jet. Fig. 104 illustrates an engine- driven unit which can also be driven by a steam turbine or motor. It contains 'some new features in the design of the working parts of the pump, and in the condenser, and may be built either on the two-stage or single-stage principle of the pump, depending upon the conditions. The principle involved is simple and offers the designer full range for his ability for getting good results with few working parts. The water is taken from the bottom of the con- denser and led to the center of the impeller, when the condenser is alongside of the pump. When the condenser is on top of the pump, the water is guided to each side of the impeller. The removal of the non- condensable vapors is accomplished either by a rotative dry-vacuum or by a centrifugal vacuum pump. From the illustration it can be seen that compactness and space have been considered of first impor- tance. Figs. 105 and 106 show the ar- rangement for vertical and horizontal jet condensers, with trimotor and birotor pumps, driven by steam tur- Fig. 103. Steam-turbine-driven Vertical Marine Hot Well Pump. bine, motor, or engine. Figs. 107 and 108 show a centrifugal air-jet conden- ser in which cold water for the condensation of the steam is forced through a Fig. 104. Engine-driven Pump Unit. .Air Suction n -Injection Fig. 105. Arrangement for Vertical Jet Condensers. (113) 114 CENTRIFUGAL PUMPING MACHINERY Air Suction Fig. 106. Arrangement for Horizontal Jet Condenser. Fig. 107. Centrifugal Water Jet Condenser with Centrifugal Air Exhausting Pump. POWER-STATION WORK 115 jet nozzle, absorbing all the vapor contained. This method of extracting the air dispenses with the usual dry-vacuum pump. The water and con- densed vapors are removed by a circulating pump in such a manner that Air Pump Discharge Fig. 108. Section of a Centrifugal Water Jet Condenser with Centrifugal Air Exhausting Pump. the air is drawn off by the outside impeller vanes of the circulating pump. This construction allows direct pumping, and obviates a separate air or vacuum pump, and so produces a compact unit of low cost, high efficiency, and minimum expense for operation and maintenance. CHAPTER XIX. DOCKS. ONE of the most extensive and important uses of the centrifugal pump is its application to pumping out docks, and it is here that the largest indi- vidual units, ranging from 36- to 72-mch discharge opening, are used. This service calls for the handling of large volumes of water in the least possible time under heads varying from to approximately 48 feet for the drainage pumps, and to 40 feet for the main pumps. Great difficulties are met, as alternating current is largely used, necessitating constant speed and constant horse power over the entire range of capacity and head. A great many pontoon and floating dry-docks have been equipped with this type of pump, and have shown that the time taken to pump out docks may be lowered considerably, and that the electric power required may be re- duced about 40 per cent below that which was used by the old-style pumps. It is to be noted that in a dry-dock the sectional area decreases as the lift increases, and that the most economical heads will be found from about one foot above keel block to the bottom of dock. The largest volumes are pumped at the smallest heads, and vice versa, therefore the average efficiency is very important, and in considering this average efficiency note should be made between what levels it is to be considered. The capacity is usually judged on the basis of an average between zero head at mean high-water elevation and an elevation 1 foot above the top of the keel blocks, which is usually from 4 to 6 feet above the bottom of the dock. The keel block having the greatest elevation is the one usually selected as the pumping point, correction being made for tidal variations. In this way the results are reduced to a condition of constant mean high- water level outside the caisson during the operation of pumping. Dock pumps are of both the vertical and horizontal type, the latter being the more common. The efficiencies in dry-dock main pumps should be judged on the average operating conditions given, and should be the ratio of the work done in pumping against the head to the power input of the motor measured at the motor terminal. The head for the main pump is usually taken from zero at mean high- water level, encountered when the level of the water in the dock is 1 foot above the keel blocks. The efficiency of a well-designed dock equipment will reach about 42 per cent rated on the above basis, and about 45 per cent from the bottom of 116 DOCKS 117 the dock, the pumping unit being charged with all f Fictional losses in piping, valves, etc., with a motor efficiency of 92 per cent and from 85 to 90 per cent efficiency for the pumps themselves. Since all such pumps are required to run only for short periods and at long intervals, efficiency is of small importance, and the first consideration is reliability. Pumps are generally furnished in duplicate. It can be assumed that a dock is emptied 2000 Fig. 109. Performance Curves of 54-inch Dry-dock Pump. times in ten years. With this as a basis, each per cent of increased efficiency would show the value of the pumps of high efficiency, as demonstrated towards the end of this chapter. The two types of docks are the graving or dry dock and the floating or pontoon dock. The former has a permanent pumping station located at a convenient place near the dock. The station is usually in a depressed chamber in the ground, so that no projections are encountered in handling the hawsers of the ships. The arrangement of piping is usually made as short as possible to avoid losses. The floating pontoon dock is of a somewhat different character, and may be either in sections or in one unit, arranged so that, when a vessel is to be clocked, the pontoon is submerged by admission of water into its compart- ments to pass it under the vessel. The pumps then empty the dock com- partments until the vessel and dock rise sufficiently to expose the vessel so that work can be done. The centrifugal pump has always been considered ideal for such pur- poses, and the low cost of installation, when compared with that of recipro- cating pumps, makes its use universal. 118 CENTRIFUGAL PUMPING MACHINERY The motive power may be either a steam engine or electric motor, depend- ing upon the installation. Engineers usually adopt an electrically driven c -a 1 Pump Based on Total Head (Bott 2 " " " " " (1 Foot Above Top >f Ke 1 Bio S Pump and Motor Based on Static Head (Bottom of Dock) ige 'Efflcieuc JU1 Depth of Water Dock Fig. 110. Performance Curves of 45-inch Dry Dock Pump. pump on account of the convenience with which current can be had for operating. The time taken for emptying large docks varies from 1 to 2 hours. Fig. 1. Pump based on Total Head (Bottom (jf Dock) ?L_? I" '1 " I " Average Efficiencies: -J !- I (1 Ift.above Toty of Keel Block) ip ijMoto based on Static Head |(Bottc|ni of Dock) | " " " " " " .(I Ft.above JTop of Kee 35 20 16 -Depth of Water in Dock Fig. 111. Performance Curve of 36-inch Dry Dock Pump. 109 shows the curves of a 54-inch pump, Fig. 110 of a 45-inch pump, and Fig. Ilia 36-inch pump. These show the changes in pumping conditions DOCKS 119 and the variation in capacities and heads under constant speed with the almost constant horse power necessary to prevent overload on the motors. It is evident from an examination of these figures that the problem is a / / \ Suction Head Impeller Casing Fig. 112. Velocities in Single-entrance Impellers for Dry-dock Work. difficult one, as the design of the impeller must take into account the varia- tion in area of the dock at its various levels, and the outside tide-water levels, and at the same time the varying power required from no head to full Suction Head Impeller Casing Fig. 113. Velocities in Double-entrance Impellers for Dry-dock Work. head must not overload the motor. The curves show how it can be accomplished. Vertical pumps having one suction entrance may not give as good effi- ciencies as horizontal ones with double-entrance pipes, as these allow a lower velocity for the same amount of water. Fig. 112 shows graphically 120 CENTRIFUGAL PUMPING MACHINERY the velocities in a single-entrance vertical pump, which can be compared with the ones in a same size horizontal or vertical pump having double- entrance pipes, shown in Fig. 113, the capacities and heads being the same. NEW DRY-DOCK AT NORFOLK NAVY YARD. The new dry-dock and pumping equipment at Norfolk Navy Yard were completed in 1909. The pump house is located at the side of the dock and below the surface of the ground, as shown in Fig. 114. The dock is known as No. 3, and the pump well is complete with roof, galleries, stairs, mezza- nine floor, and ladders to the suction pit, to give easy access to all parts. Owing to the fact that no anchor bolts could be secured to the walls or well bottom, it was necessary to carry the entire weight of the pumping machinery on cross beams and on trusses set into pockets in the side walls, which are clearly shown in Fig. 114. In running, no vibrations are set up, which proves the ability of the structure to easily support the entire weight. The plant comprises two 54-inch double-suction volute pumps, and two 12- inch drainage pumps, including all gate valves, operating gears, and motors. The main motors are 550-horse-power three-phase 60-cycle 220-volt in- duction motors, operating at a maximum speed of 200 r.p.m. when the secondary windings are short-circuited. The rotors are equipped with collector rings, and full controller and resistances for starting under full- load torque, with only full-load current, and operated continuously at any speed from 75 per cent to full-load speed. The motor frames have ventilating openings allowing free circulation of air around the windings. The windings are covered with impregnated insulation to prevent damage from dampness, and have been tested to stand an alternating electromotive force of 5500 volts for one minute. The shafts are open-hearth steel and all bearings are self-oiling and self-aligning, with ample surface to insure cool running. The rise in temperature in the motors does not exceed 40 C. with the surrounding air at temperature of 25 C. Fig. 82 shows the char- acteristic curves of the motors. The drainage-pump motors are of the constant-speed, 60-cycle three- phase, 220-volt induction type, with a normal output of 60 horse power at a speed of 514 revolutions, and are provided with starting devices so that the motors can be gradually brought up to full-load speed and full-load torque with no more than full-load current. These motors possess the same details of winding, ventilating devices, shafts, and bearings as the others, and were subjected to the same temperature tests. There is also provided a transformer bank, consisting of three 50-kilo- watt transformers for operating the capstan and valve motors, which are of the 2300-230-volt oil-insulated single-phase type. These were operated for twelve hours continuously with 2300 volts primary, at the rated out- put in amperes and a unity power factor, and heated to only 35 C. with DOCKS 121 122 CENTRIFUGAL PUMPING MACHINERY the surrounding air at 50 C. At the end of the run the load was taken off and the rise in potential did not exceed 40 volts. An alternating electro- motive force of 10,000 volts was applied for five minutes between primary and secondary, the latter to the core, together with an alternating electro- motive force of 4000 volts, momentarily applied between the low-tension winding and the core. Each transformer was also operated for two hours continuously with 2300 volts primary and an output of 50 per cent in ex- cess of normal. There is also provided a complete lighting transformer of the 5-kilo- watt 2300-volt, oil-insulated single-phase type, for operating motor-driven fan and lights. A full switchboard of four panels carried on suitable iron framework, two panels having three-pole single-throw 300-ampere automatic oil switches and 250-ampere ammeter for the 550-horse-power motors and similar ones for the 60-horse-power motors. The third panel controls the bank of three 50-kilowatt transformers and the fourth controls the capstan and lighting as well as the valve motors. The gate valves for the pumps consist of two 60-inch suction valves and two 54-inch discharge valves, besides smaller valves on the drainage pumps. They are of the double-gate type, with iror bodies flanged at both ends and bronze fittings throughout. Each valve is operated by three-phase 60- cycle constant-speed 230-volt induction motors, working through cone friction clutches, fitted with an adjustment, so that, in case of obstruction in the movement of the gate valves or failure of the limit stops, the friction clutches will slip without overloading the motors. The arrangement of the clutches is such that the motors can be started light and the loads applied gradually. All of these valves are arranged to be operated also by hand through extension stems and hand wheels, and are fitted with limit stops and indicators. The drainage pumps are fitted with 16-inch valves on the suction and 12-inch valves on the discharge, with equipment similar to that of the large ones. The main pumps are of the double-suction volute type, horizontally arranged on the shaft and directly connected to the 550-horse-power motors through special couplings. The motors and pumps are mounted on con- tinuous bedplates for proper alignment. The pump casings are in halves parted on the horizontal line in order to give easy access to the internal working parts. Suitable manholes are provided on the casings to facilitate internal inspection and cleaning. The suction elbows are made in halves to allow the removal of the impellers and shafts without disconnecting the couplings from the shafts. The bronze shaft bearings are in halves fitted to the hubs of the suction elbows, and are provided with lubricating de- vices. The main casings are of the volute type, with a diffusion throat designed to give the maximum of efficiency. The casings are ribbed to> DOCKS 123 rithstand any stress to which they may be subjected. The impellers are the inclosed type, with passages connected to the two suction elbows, turned and polished all over to minimize skin friction, and are balanced so as to run true. The shafts are of the best nickel steel, ground and highly finished, and fitted with couplings at one end to connect to the motor. The bearings and stuffing boxes are composition-bushed and provided dth lubricating devices. * The drainage pumps, smaller in size and of vertical form, possess the same details of construction. The main pumps were each required to handle an average of not less than 68,000 gallons per minute when starting against a static head 10 20 30 -10 50 60 70 80 90 100 Per Cent Eff. Fig. 115. Performance Curve for Total Plant Dry-dock Work. zero feet and ending with a static head of 36 feet through the system )f piping. The power delivered by the motor to the impeller shaft at 50-60 70 80 90 100 Capacity 1000 G.P.M. 120 Fig. 116. Pump Performance Curves for Dry-dock Work. no time could exceed 550 horse power. The average actual capacity of the pumps was 76,000 g.p.m. against a total head of 36 feet. The 124 CENTRIFUGAL PUMPING MACHINERY US' 24' 20' 16' 12' 8' Depth of AVater in Dock Fig. 117. Pump Performance Curves for Dry-dock Work. dock, with a capacity of 17,000,000 gallons, was emptied in two hours. Fig. 115 gives the performance of the entire plant as tested at the Navy Yard, Norfolk, Va. The individual performance of these pumps is shown in Fig. 116, giving the characteristic curves as usually laid out; Fig. 117 24 20 10 12 Depth of Water in Dock 500 100 Fig. 118. Pump Performance Curves for Dry-dock Work. shows the characteristic curves based on dock levels; Fig. 118 shows the average performance for the whole unit, including motors and piping. These curves show the exceedingly high efficiency of 92 per cent for the 90 86 82 78 74 Elevation of Water in Dock Fig. 119. Pump Performance Curves for Dry-dock Work. DOCKS 125 Fig. 120. Recent Type Dock Pump at Tees Dock, Middlesbrough, England. Fig. 121. Another Type of Dock Pump. 126 CENTRIFUGAL PUMPING MACHINERY pump alone, with allowance for frictional resistance through the piping and valves. The efficiency, including pipe and valve friction, reaches 85 to Fig. 122. Plan of Installation at Tees Dock, Middlesbrough, England. Fig. 123. Plan of Installation at Tees Dock, Middlesbrough, England. 88 per cent for a capacity of 68,000 g.p.m. at about 30-foot head, and re- mains practically constant up to total head of 36 feet, the capacity dropping to 62,000 gallons. DOCKS 127 This installation is the most comprehensive and up to date in dock pump- ing, and forms a valuable addition to the Navy, as it provides for the dock- ing of the new large battleships. Another dry-dock performance from the Portsmouth Navy Yard is shown in Fig. 119, showing an average efficiency of 43.2 per cent for motors, pumps, and piping, including all losses. Representative types of dock pumping, illustrating some later developments, are shown in Figs. 120, 121, 122, and 123, the first two showing 30-inch pumps, and the latter two an installation at Tees Dock, Middles- brough, England, consisting of 48-inch pumps with a mean capacity of 42,000 g.p.m. and 34 feet total head at 295 r.p.m. These pumps gave a maximum efficiency as high as 85 per cent for the pumps themselves, and the power of the motor at the lowest head did not exceed by more than 16 per cent that required at the point of best efficiency. Each pump is operated by a 400-horse-power three-phase motor. Fig. 124 shows the usual arrangement for pontoon or floating dock pumps, where the pumping conditions vary from those of graving docks, because the upper surface is uncovered and the weight of the ship is sup- ported by the dock. The side sections of a pontoon dock are small and therefore quickly emptied. The difference in the level of the water within the dock and without is to be taken into account, as the head curve shows an increasing lift from the start until the floor is uncovered, then a rather quick reduction of lift, which gradually increases until the dock is emptied. Fig. 125 shows part of the installation at the League Island or Philadelphia Navy Yard, consisting of four 54-inch units, mounted in pairs on one continuous bedplate. Fig. 126 shows a vertical installation for a Japanese Navy Yard, containing three 48-inch pumps. Careful construction of the integral parts of such pumps makes it possible to obtain efficiencies as high Fig. 124. Arrange- ment for Pontoon or Floating Dock Pump. as 90 per cent under the favorable conditions of dock work, handling a large amount of water, and demonstrates that this type is unquestionably the most economical one. It is in such installations that the centrifugal pump is at its best, and results can be obtained which cannot be surpassed by any other type. 128 CENTRIFUGAL PUMPING MACHINERY Fig. 125. Part of Installation at Philadelphia Navy Yard. Fig. 126. Vertical Pump at Japanese Navy Yard. CHAPTER XX. CENTRAL FIRE-STATION SERVICE. IN fire service, centrifugal turbine pumps have been a success from the beginning. This is largely due to the fact that they have been built to suit special requirements, which have been carefully studied out by the insurance companies and their board of inspection. Much of the trouble found in the installation of a turbine pump is due to insufficient or incorrect knowledge of the exact requirements and pumping conditions. Appreci- ating this fact, the Associated Factory Mutual Fire Insurance Companies developed a set of specifications covering the essential features under which many pumps have been built which are giving good results. The essential characteristics of Underwriter turbine fire pumps are ruggedness and strength, liberal water passages, noncorrosive material for all working parts, ease in dismantling, and certain special features for fire fighting. These are not secured at any sacrifice of simplicity and reliability, and with the help of such specifications the actual responsibility as to design and efficiency lies with the designer and manufacturer. Each manufacturer must have the drawings and details approved and a sample of the pump carefully tested out at the factory under supervision of the underwriters. After this the manufacturer must agree that all sub- sequent pumps shall be equal to the sample tested, and that no changes in design will be made without the sanction and approval of the board. These pumps may be run by a motor or a steam turbine; the former, however, is subject to the risk of loss due to the interruption of the electric current and is not considered as good a fire risk. In order to make this satisfactory, the source of electric power should be made as reliable as a steam supply. There are four standard sizes of these pumps of the following capacities : 500, 750, 1000, and 1500 gallons per minute. They are made for speeds suitable for the standard motors and steam turbines. The efficiencies required are between 50 and 70 per cent, depending upon the size, and are considered reasonable. Each pump is required to discharge a certain portion of its full capacity against a high pressure and considerably more than its capacity against a pressure of 75 pounds without overloading the motor more than 25 per cent. The higher pressures are needed for high buildings or for fires at distances which require long lines of hose. The pumps will give the neces- 129 130 CENTRIFUGAL PUMPING MACHINERY sary pressures within a certain range, with constant-speed motors, but it is better to use variable-speed motors. The pumps should have the supply under a head for priming, as they will not pick up their suction water. There is a possibility of making an automatic source of priming supply. In any case, for fire purposes a special priming tank of about four times the capacity of pump casing and pipes should be installed. On long suction pipes and large installations an independent motor-driven air or vacuum pump should be connected to the pump casing for exhausting the air. Where a reliable supply of steam or compressed air is available, an ejector or exhauster may be used, but care should be taken to have it properly proportioned for its service. Fig. 127. Standard Underwriter Type, a 1000-gallon Fire Pump with Waterproof Motors. When foot valves are used they should be of the multiflap type, with an area of at least 150 per cent of the pipe area, and the flaps should open toward the sides of the valve so as to give an unobstructed opening. Screens should have a clear area of 200 per cent of the pipe area. If check valves are used on the discharge, they should also have the clapper or valve folding back against the wall. The motors should not burn out if all the streams are shut off or when they are opened up to more than the capacity of the pump, and they should easily stand a load of 25 per cent over that for which the pump is built, and should be protected from possible leakages from the pump. The control affected by throttling, speed variation, or by both, should be CENTRAL FIRE-STATION SERVICE 131 considered. The throttling of a pump should be done in the discharge valve and not in the suction valve. Speed control is more satisfactory and can be obtained with both motors and steam turbines. There are several classes of electric motors used for fire pumps, alternating-current motors of constant-speed induction type, and direct-current motors. In the induction type of motor, with the pump discharge valve closed, the torque required is about oCTper cent of that for full load. From this it can be seen that the motors may be readily started on about 65 per cent of the regular voltage and brought up to speed with not more than double the full- load current. Full voltage can be obtained with a little increase in current and the valve opened for the desired delivery. 300 ^ __ ^ f f s ^ 260 \ ^ &> ^^ \ o 80^ ~0 . . \^ * s \ - 3 IllO a --> 1 ^o > ^ t. \ /O^H >J rt 110 10 i ^ / ^ & x ,\ o 1 00 10 ^ ^ ^~ ^ 10 > 80 /< /<\ v v GO'S 60 I s 100 200 300 400 500 GOO 700 800 9001000 1200 Gall. Per Min. 1400 1600 Fig. 128. Performance Curve Underwriter Pump. The method of regulating shunt-wound or direct-current motors of variable speed is simple and well understood, and the variable-speed in- duction motor is similar to the direct-current armature in control. Fig. 127 illustrates a 1000-gallon fire pump with waterproof motors of the standard Underwriter type. Fig. 128 shows the characteristics for a 1000-gallon four-stream pump, designed to meet all the requirements, and shows that on the Underwriter basis of 250 gallons per stream, 2j-inch nozzle, the pumps would give 4 streams at 240 feet head, or 103 pounds pressure; 5 streams at 185 feet head, or 80 pounds pressure; 6 streams at 105 feet head, or 45 pounds pressure; 3 streams at 290 feet head, or 126 pounds pressure; 2 streams at 310 feet head, or 135 pounds pressure. 132 CENTRIFUGAL PUMPING MACHINERY A maximum head of 318 feet, or 137 pounds pressure, is reached with a capacity of from 300 to 500 gallons or one to two streams. With this range it can be seen that the centrifugal fire pump can compete with the reciprocating fire pump. The horse-power output of the motor never reaches 100, while the Under- writer requirements for this size of pump allows for 100 to 107 horse power. The theoretical horse power required to discharge one stream of 250 gal- lons under a pressure of 100 pounds is 14.61, to which must be added the lift of the suction supply and the losses in the pipes, or about 1.4 horse power, making a total of 16 theoretical horse power for each fire stream. The following is the record of a test of a 1000-gallon centrifugal fire pump made under the supervision of the inspectors of the Associated Factory Mutual Fire Insurance Company, Nov. 29, 1910. The pump was designed for a 100-horse-power induction motor, but the test was made with a direct- current shop motor, temporarily connected to the pump. The speed of the motor, with no load, was given on the plate as 1200 r.p.m. Assuming some falling off from this speed under load, the test was made at as nearly 1150 r.p.m. as was practicable. The suction was taken from a submerged tank through a strainer and foot valve ; the discharge was conducted to a large box with properly fitted baffle plates and a 36-inch weir. Two tests were made with three IJ-inch'and four IJ-inch Underwriter play pipes, discharging through 50 feet of cotton rubber-lined hose into the weir box, and they checked up reasonably close with the weir measurements. The revolutions were obtained by an English tachometer checked up occasion- ally by an ordinary speed indicator. The pump was run without stopping for about an hour and a half. No difficulty was experienced with any of the bearings. The behavior of the pump under all loads, from no load to overload, was very satisfactory. The use of three stages in the con- struction of this pump makes it possible to secure a much higher pressure when running at half capacity than would be possible were only two stages employed. The pump was tested up to 240 pounds pressure and showed no weak- ness. Subsequently the pump was opened and the first impeller and sec- tion of casing removed. The results of the test are given in the accompanying table. The capacity of 1000 gallons per minute was obtained under a total head of 236 feet at 1150 r.p.m. and with an efficiency of 66J per cent. The motor horse-power output drops off at the higher discharges, so that it is not likely that the motor could be overloaded. The electrical readings for computing the efficiency of pump were taken from Weston instruments used in the regular testing work of the shop. One of the first large cities to install this type of fire pump was Brooklyn, N. Y., followed by New York and San Francisco. Philadelphia, Winnipeg, CENTRAL FIRE-STATION SERVICE 133 g a s h I o 1 II O ss i i-i C Ms* Oi > oo ^ i-it^r^ioocoo Tt'COCOCOOOO' (O* i Gi i i *"^ C^ CO ^t 1 *O *O CO t^t^GOOOOOOOoaoOOOGOOOCXJCX) '^r'^iO CO CO CO '5.S 00000 il 'sa is- |l O 00 OO l 1 OO i-H r-l i-H Tj< T-H ,-H r-H = Kg 0. 3ci i iO O O O O O O *O O O iO O O "o~3 i OOiOOiOOOOtOiOTtiiOiO 134 CENTRIFUGAL PUMPING MACHINERY and the New York branch at Coney Island installed gas-engine-driven tri- plex reciprocating power pumps. A brief description of the first high- pressure fire-service turbine pumping plant installed at Brooklyn, N. Y. will be of interest. This consists of two separate stations, one situated at corner of Wil- loughby Avenue and St. Edwards Street, and comprises three units. The other or main station is at the northeast corner of Furman and Joralemon Streets and has eight units. Unlike the New York high-pressure fire- service stations, both these plants can secure water from either a salt-j water or a fresh-water supply, requiring a more careful design in order toj give the same results when working with a suction head of 50 pounds pressure and when lifting the suction water from a supply 20 feet below the pumps. The salt-water supply comes from the East River, and the fresh-water from a 20-inch main. This arrangement is of great importance, as it insures &j fire stream even though the city main that supplies the suction water should break. It is usually required that a pumping station relying upon a salt-water supply should be fitted up so that equally good results can be obtained with salt water or fresh. In installing the high-pressure fire-service stations in the Borough of Brooklyn, equal efficiency was demanded under both conditions. The main station at the foot of Joralemon Street consists of eight units, as shown in Fig. 129. The relief station at Willoughby Street is shown in] Fig. 130. The pumping units are all identical, and interchangeable. Fig. 131 shows one of the units complete. It consists of a six-stage turbine pump directly connected to the motor. The pumps are of the horizontal type, each capable of delivering 3000 gallons per minute, when operating at 735 revolutions and delivering against a pressure of 300 pounds per; square inch, with a suction lift of 20 feet, or under a supply pressure of 50 pounds per square inch or less when taking supply from the 20-inch main. The size of the suction openings for each pump is 12 inches and the dis-j charge 10 inches. These pumps so operate that the pressure system can be" regulated between 100 to 300 pounds, by increments of 50 pounds, the speed remaining constant, the regulation being effected by the design of impeller and by special regulating valves. The brake horse power of the motor when operating at any lower pressure than 300 pounds does not exceed that at 300 pounds, thereby obviating overloading the motor when the capacity increases and the pressure drops. Each pump is directly connected to a three-phase 25-cycle 6000-6300- volt 800-b.h.p. motor of the induction type, and fitted with all the necessary starting and controlling devices. The full-load efficiency of these motors is 95 per cent, the power factor 96, and the slip 2 per cent. At three- quarters load, the motor efficiency is 95 per cent and the power factor 95J per cent. CENTRAL FIRE-STATION SERVICE 135 136 CENTRIFUGAL PUMPING MACHINERY CENTRAL FIRE-STATION SERVICE 137 .s ex 138 CENTRIFUGAL PUMPING MACHINERY The results of the tests given in Figs. 132, 133, and 134 show an efficiency on salt water of 75 per cent and on fresh water of 76 per cent for the pumps, 425 400- 37& 350- . 325 jH 300- d- 275- "250- |225- 7200- '' 150- 125- 100- 75- 50- JWU ^-^^, ^ 800 700 600^ KAA W V 00 ^700 70, 60^ ? 3< ^ ^ S 4 s i \\" ^ #^ X ^ "Z 3 500 -&H -r-400 "^300 - 200 '- 100 y "^ \ \ j / \ \ 100 r| 300 fl 200 100 30 20 s \\ y N \\ 10' \ I 5 Gallons per Minute Fig. 132. Performance Curve for Brooklyn Fire Pump. and an over-all efficiency for motor and pump of 72.2 per cent. The water was measured by a Venturi meter in the mains. The capacity of the pumps at various pressures is shown in the curves, as well as the horse power re- quired for operation, and the characteristics controlling them. Gallons pei Minute Fig. 133. Performance Curve for Brooklyn Fire Pump. When operating with salt water the pumps are primed by a motor-driven vacuum pump operating automatically. It was found that with either salt or fresh water the pumps can be started and brought to full speed in less than 45 seconds, without having the starting current exceed 150 per cent of the full-load current of the motors. An additional time of 10 seconds was required to bring the pump up to a pressure of 100 pounds. Therefore, a unit can be started from complete rest and brought up to CENTRAL FIRE-STATION SERVICE 139 speed under a water pressure of 100 pounds in less than one minute. The pressure can be varied from 300 pounds down without overloading the motor, an extremely severe condition. It is interesting to note the results obtained by gas-engine-driven power pumps for similar purposes, notably those at Philadelphia and Winnipeg. The mechanical efficiency at Winnipeg is from 77 per cent to 82 per cent on pumps of 1800 gallons per minute capacity, and the pressure 300 pounds, giving a B.t.u. efficiency per indicated horse power of 8600, and on water horse power of 11,774. The Philadelphia plant gave 11,160 B.t.u. per indi- cated horse power and the Coney Island plant gave 12,682 B.t.u. per water horse power. Such gas-engine-driven fire pumping equipments operated on producer gas will give from 180 millions to 210 millions duty in foot .3^800 Gallons per Min. Fig. 134. Performance Curve for Brooklyn Fire Pump. pounds of work done per one million B.t.u. consumed, which is a very low cost for operating, but as they are not operated continuously this is not of great consideration. They have the advantage of low cost of installation and operation, but these are offset by the flexibility of turbine-pump instal- lations in handling larger capacities at any desired pressure. Electrical or turbine-driven centrifugal-pump fire stations are considered more reliable than gas installations, and the question of economy in opera- tion is secondary, because the total cost of pumping is only a small portion of the running expenses. The capacities of the various stations built are as follows: Brooklyn, 24,000 gallons per minute at 300 pounds pressure and 40,000 gallons at 150 pounds pressure; New York, 30,000 gallons; Coney Island, 4500 gallons; Philadelphia, 9000 gallons; San Francisco, 20,000 gallons; Winnipeg, 9100 gallons. From this it will be seen that the gas pumping engines are of considerably smaller capacities. The requirements in all such stations should be for a definite capacity and head, for although it might seem that an excess capacity is good, it usually overloads the motors, and extra power has to be paid for. 140 CENTRIFUGAL PUMPING MACHINERY High-pressure fire-service stations are becoming a necessity in this country owing to the high buildings and the larger volumes of water re- quired in fighting fires. A pressure of 300 pounds is usually needed to overcome frictional losses and supply 2J- to 3-inch nozzles at a pressure of 100 pounds. Portable fire engines to handle such conditions are entirely out of the question, and the only practical means for meeting the situation is, the permanent high-pressure station. Fire Engines. The application of turbine pumps to automobile fire engines is an interesting development. These pumps are made in stages of one, two or three, suitable for the speed of the automobile engine, and the capacity varies from 250 gallons to 1000 gallons for pressures of 220 pounds down to 120, when delivering water through 1000 feet or 500 feet of 2j-inch hose and IJ-inch nozzles. A usual requirement is that a pump should deliver! 700 gallons per minute against a pressure of 120 pounds and 300 gallons against a pressure of 200 pounds. The importance of obtaining the highest pressure is apparent on account of internal friction in the hose which reduces the effective pressure at the nozzle. The regularity of flow and absence of water hammer make this the best solution of a fire pump for a portable fire-fighting apparatus. The action of a turbine pump allows the shutting off of any particular hose, nozzle or all, without tendency to burst the hose as may be the case with reciprocating or rotary fire pumps. The pump is considerably lighter and takes less space, requiring no suction or discharge air cham- bers. The pumps are designed to fit on the back part of the chassis of an automobile to be driven by a shaft directly from the engine or through transmitting gears. The starting up when a heavy suction lift is encountered is accomplished by means of a small rotary vacuum pump operated by a magneto on the shaft of the engine, which exhausts the air in pump and pipes. When the tur- bine pump is primed, it is thrown into action by the clutch and develops immediately its head pressure and the small vacuum pump is then cut out. A special by-pass and valve are attached to the pump to allow the water to pass from discharge to suction in a small stream so as to prevent heating should all nozzles be shut off and the clutch left in while the engine is running. It is therefore apparent that rapid progress indeed has been made in adapting the turbine pump to an increased field, giving a better and more efficient fire engine. In large cities the automobile turbine fire engine is usually designed to carry nothing but the pump. In smaller towns a combination of fire engine, chemical and hose cart is desirable, as the most useful apparatus for a fire department at minimum upkeep cost. The economic advantages of such a combination seem beyond dispute and the success is already assured. The reliability is only a question of design CENTRAL FIRE-STATION SERVICE 141 with proper power of gasoline motor, which should be between 100 and 125 horse power for the capacities stated, and well designed to stand up under continuous service without overheating. The tendency in the large cities is to supplant the fire engine in centers of business, by high- pressure systems of independent mains and hydrants, such as have been described for New York and Brooklyn, but for the suburbs and smaller cities and towns these independent fire engines will be needed. The Underwriters have set a standard for such fire apparatus that can probably only be reached by the application of turbine pumps to the fire engines, just as the turbine has supplanted the reciprocating pump in stations for high-pressure fire mains and on fireboats. Such automobile fire engines will weigh complete from 9000 to 13,000 pounds, and will travel at the rate of 30 to 35 miles an hour over city streets. CHAPTER XXI. FIREBOATS AND SHIPBOARD SERVICE. FIREBOATS. TURBINE pumps are extensively used for fireboat equipments where reciprocating flywheel pumps were at one time considered indispensable. The first boats so equipped were the " James Duane " and " Thomas Wil- lett," belonging to the New York fire department; since then, however, Chicago, San Francisco, Seattle, and Duluth have followed the example set by New York. Fig. 135. Worthington Steam-turbine Pump used by New York Fire Department. The New York system of protecting its water front adds considerably to the shipping value of the harbor, and to the value of property along the water front. Fig. 135 illustrates one of the pumps installed on these boats. The boats are duplicates in all respects and are 123 feet long, 27 feet wide, and 14 feet deep. Each boat contains two units of two-stage pumps, piped to run in series or parallel, and each boat is capable of handling 9000 gallons per minute against 150 pounds pressure, or half that capacity at 300 pounds pressure for use in fighting land fires along the water front where the dis- tance is too great to be reached by direct streams. The pumps are driven 142 FIREBOATS AND SHIPBOARD SERVICE 143 by 600-horse-power General Electric steam turbines operating at a speed of 1800 revolutions. The following record of the actual tests of these boats gives the details of performance. On the " Thomas Willett," with both pumps running at 1800 revolutions per minute, the following readings were obtained: TABLE NO. 5. "THOMAS WILLETT." Pressure at nozzles. Pressure at pumps. G.p.m. R.p.ra. Size, 2 in. Size, 2J in. Size, 3 in. Size, Sin. Size, 3 in. Port. Starboard. Port. Starboard. 90 90 100 115 115 90 230 102 115 115 90 100 110 112 85 160 170 175 140 155 160 170 175 140 310 9,035 9,600 9,625 11,460 4,500 1800 1800 1900 1900 1780 1800 1800 1900 1900 1790 On the " James Duane " the following tests were made: TABLE NO. 6. "JAMES DAUNE " Pressure at nozzles. Pressure at pumps. G.p.m. R.p.m. Size, Size, 3 in. 3 in. Size, 3 in. Port. Starboard. Port. Starboard. 73 75 40 85 90 170 73 60 55 85 90 150 "i65 " 175 165 "l50" 125 165 175 330 5120 4920 6280 8280 8480 3900 1900 "i900" 1800 1800 1900 1800 53 85 87 .1800 1900 1800 On both boats the vacuum varied from 26 to 27 inches, and the steam pressure from 175 to 190 pounds. The pumps ran with no vibration, and standing on the deck one would scarcely know the pumps were running except for the water thrown. It was found that the speed for throwing full capacity could be obtained in from 25 to 40 seconds. The last test on the "Thomas Willett" consisted in putting the two pumps in series for 300 pounds pressure. Fig. 136 illustrates the internal arrangement of these pumps and the de- sign of impellers for high speeds and high pressures. Fig. 137 shows another size of similar type. These illustrations show turbine pumps especially designed for steam-turbine drives on fireboats, where the primary requirement is instant operation. The use of turbine pumps eliminates vibration in the boat when running full speed, and the ability to come up 144 CENTRIFUGAL PUMPING MACHINERY to speed in 20 to 30 seconds is something never before accomplished in fire pumps. The possibility of closing down all the hose streams suddenly A Section A-B-C Fig. 136. Internal Arrangement of Turbine Pump used on Fireboats. without harm is welcome to the operator, as in the reciprocating type of pumps care must be taken to slow down gradually or something will break. Fig. 137. Small Size Turbine Pump used on Fireboats. The boats have demonstrated their readiness for service and have rendered valuable services in several large water-front fires. There can be no doubt that for fire service the steam-turbine-driven turbine pump is far superior to any flywheel reciprocating pump, and that in time it will replace all others. ON SHIPBOARD. Because of its simplicity and compactness the centrifugal pump has come into very general use on shipboard. Some of these applications have been given in the previous chapter and under " Circulating Water," and so need FIREBOATS AND SHIPBOARD SERVICE 145 no further mention here. This type is also used very extensively for han- dling ballast water from the double bottoms, ballast tanks, and water-tight compartments, and also in bridge work and the sanitary service. Wrecking boats are always equipped with two centrifugal pumps with about 12- or 14-inch discharge permanently fixed on board. They are engine-driven and of either the vertical or horizontal type. The boilers are usually de- signed to work with salt water and to stand a lot of abuse. Fig. 138. One of the Latest Circulating Pumps of the Birotor Type used on the Battleship "Arkansas" of the Dreadnought Class. Fig. 138 is a photographic view of one of the latest circulating pumps of the birotor type used in marine practice and now installed on the battle- ship "Arkansas" of the dreadnought class. A steam-turbine-driven circu- lating unit can be arranged to operate under almost all conditions, such as high pressure exhausting to atmospheric or to the condenser, or with low- pressure steam or as a mixed-flow turbine, and will furnish circulating water with absolute reliability. It is to be noted that the impellers work in parallel, having one main inlet pipe and one outlet pipe, with a special opening on the side of the suction chamber for pumping out the bilges. The pump casing is split horizontally 146 CENTRIFUGAL PUMPING MACHINERY in order to facilitate repairs aboard ship, where space is of paramount importance. The compactness of the design lends itself readily to marine service, and in this instance the total length is only 1 1 feet for a unit having a capacity of 27,000 gallons with a maximum head of about 37 feet, including all losses through tubes, sea-cock piping, and valves. The particular feature in such a pump is the high water speed, which makes it possible to design a compact pump and at the same time to obtain the reasonably good efficiency of 66 per cent. "The minimum requirements are 15,000 gallons per minute against a head of 25 feet, of which 15 feet are suction lift, when working on the bilges. For this service there is provided a side suction opening. The main suction openings, through which the circulating water for the condensers is sup- plied, consists of two separate connections with pipes to the sea cocks." CHAPTER XXII. SPECIAL HIGH-SPEED INSTALLATIONS. THE high rotative speed of the steam turbine, and the convenience of operating it direct-connected to a centrifugal pump, have presented an interesting problem to the pump designer. Where the quantity of water is not large a small impeller may be safely used if constructed of a material Fig. 139. Novel Design of Turbine-driven Pump. which will not be cut by the water. Fig. 139 illustrates a rather novel de- sign of turbine-driven pump with a normal capacity of 2300 g.p.m. to 3300 g.p.m. and a minimum capacity of 900 g.p.m. for elevator service against a head of 150 pounds pressure. The shaft is carried in two spherical bearings, one near the steam turbine and the other on the outer end of the pump, to allow the shaft to take a true position. There is no thrust in either the turbine or the pump, but for safety a special bearing has been arranged at one end. The turbine and pump are secured to a common base plate, and the unit is entirely contained. Pumps of this design may be made with single impellers up to 300 feet 147 148 CENTRIFUGAL PUMPING MACHINERY head, without the need of diffusion vanes. Efficiencies of these pumps vary from 75 to 80 per cent. Fig. 140 illustrates a 14-inch two-stage high-speed steam-turbine- driven pump, arranged to run in series or in parallel. It was designed for 11,000 gallons per minute at 250 feet head, or 5500 gallons at 500 feet head, and is driven by a 1000-horse-power turbine running at 1600 r.p.m. The test of this unit is given in the following table: TABLE NO. 7. TEST OF 1000-H.P. TURBINE DIRECT-CONNECTED TO TWO 14-INCH WORTHINGTON TURBINE PUMPS. QUEMAHONING DAM, June 7, 1910. Series. Parallel. Number of test. 1 2 3 4 5 6 Steam Turbine. R.p.m. turbine 1,631 1,631 1,617 1,603 1,620 1,620 Steam pressure abs 121.6 122.3 112.5 113.6 124.0 126 Quality steam, per cent. 96.5 96.4 97.1 96.4 97.4 97.5 Inches vacuum 25 25 25 25 25 25 Barometer, inches 28.5 28.5 28.5 28.485 28.485 28.5 Vacuum ref. to 29.92.... 26.98 26.98 27.04 26.72 27.24 27.67 Temperature exhaust . . . 144 144 144 116.6 110 111 Hot well deg F 58 58 57 58 59 60 Air-pump suet., deg. F.. 50 60.6 61.5 67 68 71 Cooling HaOent., deg. F. 56 56 56 57.6 57.5 58.5 Cooling H 2 O, leaving, deg. F 66.5 69 64 58-64.6 60-76.8 67-85 R p m air pump 80.3 81.2 80.75 85.3 80.25 83.5 Pump. Lift in feet 15 14.65 15.6 16.15 13.2 13.0 First stage, Ibs. press. . . 104.8 123 55.5 100.3 135.5 136.3 First stage, ft. press.. . . 242.0 284 128 232.0 312.0 314 Second stage, Ibs. press. 219.8 257 149.2 101.0 137.2 138.0 Second stage, ft. press.. 507.0 294 342.0 233.3 317.0 319.0 Dist. between gauge, ft. 3.0 3.0 3.0 3.0 3.0 3.0 Total head, ft 525 611.5 360 251.8 330.7 332.5 R.p.m. pump 1,631 1,631 1,617 1,603 1,620 1,620 Weir reading, ft .594 .522 .652 .947 .388 .280 Capacity, gal. per min. . 5,430 4,480 6,180 11,020 2,880 1,790 Water horse power 722 691.5 563 654 241 151 Condensed steam 20,320 16,256 12,192 12,192 12,192 4064 Dry condensed steam . . . 19,600 15,670 11,840 11,720 11,750 3960 Dry steam per hour .... 20,400 19,620 16,430 19,450 13,240 8150 Dry steam per hour, water horse power . 28.3 28.4 29.3 29.8 55 54 Francis formulae used for weir calculations. Where large quantities of water are needed, an ordinary style of pump cannot be used, since the large diameter of the impeller gives a prohibitive Fig. 140. 14-inch Two-stage High-speed Steam-turbine-driven Pump. Sec Fig. 141. 24-inch Trirotor Pump Designed for Circulating Water for a Condenser. Fig. 142. Pump Designed for Circulating Water for a Condenser, Fitted with Diffusion Vanes. (149) 150 CENTRIFUGAL PUMPING MACHINERY peripheral velocity. For this service a multirotor pump has been designed which bears the same relation to the capacity of the pump at high speed that the multistage design does to the head. 16 Discharge, 23^ "Flange, 16- l" Bolts on lOVRadius. Fig. 143. Vertical Type of Birotor. Fig. 141 illustrates a 24-inch trirotor pump designed for circulating water for a condenser. It is practically three small impellers working in parallel so that the proper linear velocities may be obtained with the high Fig. 144. A 24-inch Trirotor High-speed Unit. SPECIAL HIGH-SPEED INSTALLATIONS 151 rotative speed of the steam turbine. Fig. 142 illustrates a similar pump fitted with diffusion vanes hi the discharge chamber and suitable for high Fig. 145. A Birotor Type of Steam-turbine-driven Pump for Circulating Water in Power Stations. circulating heads. Fig. 143 shows a vertical type of birotor suitable for certain classes of plants requiring this type. A 24-inch trirotor high-speed unit is shown in Fig. 144. It is designed Fig. 146. A Trirotor Type of Steam-turbine-driven Pump for Circulating Water in Power Stations. 152 CENTRIFUGAL PUMPING MACHINERY to handle 20,000,000 gallons of muddy river water per 24 hours at 1400 r.p.m. against a head of 130 to 160 feet, but at times as low as 95 feet. Figs. 145 and 146 illustrate a birotor and a trirotor type of steam-tur- bine-driven pump for circulating water in power stations. This application of steam-turbine drive to auxiliaries in power stations gives an economical installation and a maintenance cost less than for any other type. CHAPTER XXIH. COMMERCIAL PUMPS FOR GENERAL INDUSTRIAL USES. IT is the intention to here give a somewhat detailed description of the working parts of a standard type of pump. Many industries permit the use of such a type of pump, as, for instance, common contractors' work, irrigation, and similar work. Fig. 147 shows one of these commercial Fig. 147. Commercial Pump. pumps having an average efficiency and a low first cost. For paper-mill, dye-house, and sugar-house work a special type of pump is required, in which the internal parts are easily accessible for cleaning without dis- mantling the entire machine. This style is shown in Fig. 148, and has its casing split along the horizontal line, as shown in Fig. 149. Such pumps are especially adapted for handling material which leaves a deposit inside of the pump, which must be regularly removed. Another type, suitable for paper mills, sugar houses, and gas houses, is shown in Fig. 150, the casing and impeller being designed to handle heavy or thick liquors. They are suitable for pumping various kinds of liquors from chests or storage tanks to digesters and engines, for returning water 153 154 CENTRIFUGAL PUMPING MACHINERY from the screens back to the filters, for pumping clay water from mixers to the beating engines, and lime used in bleaching and chloride of lime ; in fact for every conceivable use in a mill. Fig. 148. Sugar-house Pump. The material out of which such pumps should be made depends upon the liquid to be handled. Most of the trouble in such pumps can be attributed Fig. 149. Sugar-house Pump Showing Method of Examination. to lack of knowledge of the liquid with which the pump is to be used, and if the conditions are known a pump can be made suitable to the particular service. Fig. 150. Paper-mill Pump. Fig. 151. Vertical Pump Showing Parts. (155) 156 CENTRIFUGAL PUMPING MACHINERY Where a vertical arrangement is necessary pumps like that shown in Fig. 151 may be used. A similar type for larger sizes and higher heads is shown in Fig. 152. In many modern buildings having deep basements, where water can accumulate, a type of sump pump, illustrated in Fig. 153, is used. This may be arranged to operate automatically, and as the pump is noiseless it is well adapted to buildings, apartment houses, and hotels. In small Fig. 152. Vertical Turbine Pump. units the turbine pump driven by electric motors with automatic float switches is being largely used in buildings for tank service. To give the reader a fair idea of how the details of a modern centrifugal pump look, the following illustrations have been inserted. Fig. 154 shows the pump itself and the equipment usually needed. The casing may be of cast iron, cast steel, or bronze, according to nature of material to be handled. The inlet and outlet are shown in their normal position, but they are some- times placed differently. Fig. 155 shows the impellers, which are also made of a material consistent with the properties of the liquid to be pumped. These impellers are balanced lor high rotative speeds and may have either COMMERCIAL PUMPS 157 r-v Fig. 153. Vertical House Sump Pump. a single side suction or a double suction inlet. They are designed accord- ing to mathematical formulas for each specific duty. The hubs or flanges shown are for the purpose of preventing leakage and for facilitating bal- Fig. 154. Horizontal Turbine Pump, Vertical Split Casing. 158 CENTRIFUGAL PUMPING MACHINERY anting. In the spaces between the sides of the impellers and the casing walls there is an area which would be unbalanced, and, in order to overcome Fig. 155. Arrangement of Impellers on a Multistage Pump. this, hubs are provided of equal diameters on each side, having a running fit of not over jo'W of an inch and having an end play of T V inch to allow Fig. 156. Thrust Bearing. the impeller to float to its proper position. Holes are drilled in the im- pellers at the middle, inside of these hubs, allowing water under the pressure to act upon the area within the rings, thus giving balance as near as is prac- COMMERCIAL PUMPS 159 ticable. This arrangement is followed for each impeller, so that there is a minimum thrust, but to compensate for any possible thrust, whether due to Fig. 157. Diffusion Vanes. suction lift or to the wear that will take place in time, a self-oiling thrust bearing is supplied at the outer end as shown in Fig. 156. The additional Fig. 158. Complete Diffusion. purpose of the thrust bearing is to provide a lateral alignment between impellers and diffusion vanes. The main bearing is integral with the 160 CENTRIFUGAL PUMPING MACHINERY stuffing box and secured to the casing by a recessed fitting. The stuffing box is sealed and of the lantern gland design, with soft packing each side of the gland. Fig. 157 illustrates the diffusion vanes, through which the water is guided in mathematically calculated passages, transforming the velocity head into pressure with the least possible losses. The material of these rings also depends upon the liquid. There are two rings, one with vanes and one without. Together they form the complete diffusion chamber shown in Fig. 158. PART IV. PRIME MOVERS FOR DRIVING CENTRIFUGAL PUMPS. CHAPTER XXIV. ELECTRIC MOTORS. THE great majority of the centrifugal pumps in service are motor-driven. In many cases the importance of the relationship between the pump and the motor has been too little appreciated. In order to attain the best re- sults, the engineer must fully understand the peculiarities of the pump, the variations in load during operations, the characteristics of electric power, and of the plant as a whole. The exact capacity and rating of the motor are of great importance, especially where power is bought on the motor rating. If the motor is too small, it will be constantly overloaded, and if it is too large, the customer pays for power not used. Furthermore, the power required at rated speed and head should not in itself determine the size of the motor. Maximum conditions should also be taken into account. Having determined the size of motor required, it is of utmost importance that the proper style be selected for the limits of current available. It is in the selection of the class or style of motor that an understanding of the characteristics of the pump is absolutely essential, as it is necessary to select a motor which will take care of the varying loads and meet the different starting conditions. Where the head must vary, this may be accomplished by changing the speed, and a motor must be selected which permits speed regulation. The designer of the pump must, therefore, carefully consider the nature of his motor when laying out the characteristics of his impeller. On the other hand, the electrical engineer should design his motor to suit the characteristics of the pump. A pump should be designed for an average head at maximum efficiency and maximum head at average efficiency, and motor and pump should be so proportioned that under the varying conditions the motor will not be overloaded except within a certain range, and will allow increased capacity for lower heads with nearly constant power. The pump should be designed for a wider range than is usual in order to give the motor a restricted output and to prevent excessive overloads under abnormally low heads. All pump characteristics should have a flat curve over a considerable range, making 161 162 CENTRIFUGAL PUMPING MACHINERY the horse power nearly constant between the limits of the working condi- tions. This is especially necessary with induction motors, as both the power factor and the efficiency are reduced at light loads. The problem of starting torque is easily solved with direct-current motors, provided their size is properly selected to correspond with the capacity and head of the pump under all conditions to be met. The majority, however, of pump installations are of the alternating-current type. An induction motor of proper capacity will start a pump successfully, pro- vided the initial rush of current is tolerated. A synchronous motor may fail to start the pump, as the torque of this type of motor is usually small, and increases only as synchronism is reached; while the torque of the pump increases with the square of the velocity, thus producing a critical point in starting. Several remedies have been suggested, chiefly to start the pump empty and prime after full speed has been reached, also to start with a smaller independent motor to build up to speed. The load on the motor may be relieved in a primed pump by closing the discharge valve and allow- ing the pump to deliver water through a by-pass back to the suction under a lower head, removing a portion of the load from the motor. As the start- ing torque in a synchronous motor is due to eddy currents and hysteresis, more to the former than to the latter, it follows that every means of increasing the eddy currents will help the starting of the pump. The synchronous motor will hardly equal the induction motor for starting turbine pumps, as it has a large air gap and leakage and an uneven distri- bution of the secondary winding, which produces a lesser torque than can be found in an induction motor. Another condition to be met with is the change of pressure by speed variation. In multipolar alternating-current motors this can be accom- plished by using a different number of poles for the various speeds. The motor builder should know how the load on the pump varies with varying head and constant speed. Variation in the speed of direct-current motors is usually accomplished by resistance in the field, but this has a limited application and may prove objectionable for commutation, as the fields distort at the high speeds. The interpole variable-speed motor with a rheostat is better, although more expensive. In any case, the motor should be large enough so that the field rheostat cannot cut in too much resistance and overload the motor. The speed of the pumps should be properly selected under such conditions to avoid endless trouble due to the different speeds at different heads. One must consider that in such combinations an increase of 10 per cent in speed may increase the power 50 per cent and cause a great increase in the armature losses. This is pointed out to show the great effect speed variation may have in an installation with a variable-speed motor. Speed is one of the most sensitive characteristics, and any unnecessary variation ELECTRIC MOTORS 163 is a great absorber of power. In a synchronous installation the speed is fixed by the prime mover and the motor is selected for maximum load. There are various methods for effecting the speed regulation. For example, a two-speed induction motor can operate at normal horse power on eight poles and at double the horse power on four poles. This is accomplished by primary windings connected to consecutive poles. As a four-pole motor the connections are in parallel, and for eight-pole in series. These motors can be designed and built for three speeds, operating on eight, ten, and sixteen poles, by having a primary with two sets of windings, one to eight- pole and sixteen-pole connections, and the other for ten-pole operations. This method of regulation is entirely practical and can be applied in con- nection with turbine pumps. In vertical installations the motor should carry the weight of all revolving parts such as armature, impellers, and shafting, in order to obviate any vibration which may be set up. The pumps should be placed on good foundations. If this is not done the vibrations set up will ruin the motor or the pump. Too much stress cannot be laid upon the fact that the pump and motor designers must work together and consider each other's problems more than has been done in the past; for the pump and motor must be considered together, just as a steam end in a direct-acting pump must be designed in connection with the pump end. CHAPTER XXV. STEAM ENGINES AND MISCELLANEOUS. SINCE the development of high-speed steam engines, more engine-driven centrifugal pumps have been installed. This combination has met with CN TtWt j6A " **" Fig. 159. Vertical Cross Compound Sewage Pump. 164 STEAM ENGINES AND MISCELLANEOUS 165 favor in small plants where the engineer in charge is better acquainted with the workings of a steam engine than with a motor or steam turbine. The speed of an engine-driven pump is necessarily limited. It rarely exceeds 800 r.p.m., and generally is in the neighborhood of 600 r.p.m. Such a unit cannot be used, therefore, for all kinds of service, and must be started slowly to give the engine sufficient tune to warm up. Gas and oil engines are also extensively used direct-connected to cen- trifugal pumps. One of the largest fields for this type is in contractors' work for emptying excavations, sumps, and ditches. Gasoline and kero- sene-oil engines are used almost exclusively for this work, as they are con- venient, portable, and need very little attention. Economy and efficiency are secondary considerations. Centrifugal pumps for almost any kind of service may be driven by belts or silent chains if shaft power is available, as we need only consider the speed of the shafting and the ratio of the driver and driven pulleys to obtain the necessary speed. Where there is an abundance of water under a few feet of head and a lesser quantity is desired at greater head, a combination centrifugal pump and water wheel has been found to be very economical. The water wheel or water turbine is direct-connected to a centrifugal pump which distributes the water under the desired head through a line of piping. This combina- tion may also be used where there is an available water power for the water wheel, and a separate supply of pure water or some other liquid to be pumped. Fig. 159 shows different types of engine-driven pumps, the former a compound engine with pump between, the latter the usual arrangement of single engine and pump. These types are commonly used for supplying circulating water for condenser, sewerage work, and similar duties. CHAPTER XXVI. STEAM TURBINES. CENTRIFUGAL pumps, driven by steam turbines, are being extensively used for hot-well, boiler-feed, and circulating pumps. These combinations are efficient, occupy small space, and require minimum attendance. In the design of such a unit the chief problem is to construct turbines of 20 to 500 horse power which can operate at a speed suited to the pump. A compromise must be made between the ideal speeds of the pump and the turbine. In large power plants, where steam can be used for heating feed water, the efficiency of an auxiliary prime mover, like a steam turbine, is of sec- ondary importance, as low cost of operation will offset increased steam consumption. The rating of steam turbines in connection with pumps should always be on the maximum load, as otherwise the expected efficiency will not be obtained. There are four principal types of steam turbines : First. De Laval; an impulse turbine in which the steam is completely expanded in a single set of nozzles and all the kinetic energy is given up to a single row of blades. Second. Parsons ; impulse-reaction, where the energy of reaction of an expansion in the moving blades is added to the impulse of the steam as received from the fixed nozzles. Third. Zoelly and Rateau; impulse turbine having a series of partial expansions, the energy of each expansion being absorbed in a single row of moving blades. Fourth. Curtis; where the velocity of the steam from the nozzles is absorbed in and passes through several rows of moving blades. All other types are modifications of these. There are certain points towards which the effort of designers should be directed in order to secure the highest efficiency with maximum durability, simplicity, and cost of construction, namely: First. Reduced steam consumption. Second. Increased peripheral speed. Third. Simplicity of design. Fourth. Accuracy of workmanship. Fifth. Provision for expansion and contraction under all conditions of load and steam pressure in a manner not to interfere with safe operation. 166 STEAM TURBINES 167 The small commercial turbines on the market to-day are the de Laval, Curtis, Terry, Kerr, Sturtevant, and Dake, all of the impulse type. The old-style de Laval has only one row of moving elements and one set of nozzles, necessitating high bucket speed; but in their recent designs there are several steam returns. No steam turbine should be thought of that cannot ultimately meet the economy of the reciprocating engine, and small turbines under 300 horse power should have as near as possible the economy of the best grade of reciprocating engines, if the useful field is to be extended. The following shows the effect of peripheral speeds on the economy of small turbines: 36-inch Curtis 19,000 feet per minute 30 pounds per horse power per hour; 36-inch Curtis 25,000 feet per minute 28 pounds per horse power per hour; 24-inch Terry 17,500 feet per minute 40 pounds per horse power per hour; 24-inch Terry 10,000 feet per minute 50 pounds per horse power per hour; 24-inch Kerr 17,500 feet per minute 40 pounds per horse power per hour; 24-inch Kerr 6,000 feet per minute 60 pounds per horse power per hour; 30-inch Bliss 20,000 feet per minute 40 pounds per horse power per hour. Single-stage turbines of the Electra type, of 45 horse power, running 3000 revolutions, with steam from 90 to 100 pounds, have given a steam rate of from 24 to 30 pounds per horse power condensing. This type is well adapted to small powers. It is known also by the name of Kolb, and is one of the best for use in connection with centrifugal pumps. It is entirely suitable for high-pressure steam. The only objection to it is the loss by friction in the guide blades, and the inability to use it for large powers, as there is not enough space on the periphery of the wheel for the necessary nozzles and guides. A mixed-flow turbine or one using both high- and low-pressure steam is one that will be valuable in a great many places, particularly in steel mills, where an abundance of exhaust steam from the various rolling-mill engines, hammers, etc., may be utilized for running the pumps. Turbines can be built direct-connected with pumps, and arranged to be operated either by steam at atmospheric pressure and exhaust into a vacuum, or by high- pressure steam, exhausting into a vacuum or into the atmosphere. Such a turbine could be designed with two nozzle chambers, one for high-pressure steam and the other for low-pressure steam, having independent governor control. The expansion of steam in a nozzle obtains a velocity of 7 = 224 Hi-lHzX +q(l-X)\, where HI = total head at PI; H 2 = total head at P 2 ; X = dryness fraction; q heat of liquid. 168 CENTRIFUGAL PUMPING MACHINERY The available energy in steam between boiling point and absolute vacuum is 890,000 foot pounds per pound, and the velocity V = V2 X 32.2 X 890,000 = 7550 feet per second. This is based upon the theory advanced as to molecular velocity, and be- comes important when the available energy in steam between different pressures is to be determined. The usual formula for determining this is : Foot pounds per pound of steam = 778 [Hi + C p ti (G 2 + ^2)]. HI = total heat at pressure pi\ C p = specific heat at superheated steam at pressure pi] ti = superheat in degrees Fahrenheit at pressure pi ; Gz = head of liquid at pressure p 2 ; X = quality of steam at pressure p 2 , or entropy; Vi = latent heat at pressure pi\ vz = latent heat at pressure p 2 . Entropy of superheated steam can be calculated of moist steam, ito , . -Ffi r 92, 1 2 where T and TI absolute temperature of saturated steam at pressures p\ and pz, which equals 461 (temperature in Fahrenheit). 9 = entropy of water at pressure pi; 9 2 = entropy of water at pressure p%. From this can be found the moisture in per cent and available energy in foot pounds of the steam and the amount of moisture entering into the condensing apparatus, although it is to be taken into account that some of the moisture is lost. After the total energy is found by taking the efficiency of the turbine, the energy available can then be found. The efficiency of a turbine is determined by the readings of pressure and temperatures and the amount of steam, exhaust pressure, and the electrical power of the generator, and is the theoretical water rate divided by the observed. The small steam turbines now coming into general use vary from 10 horse power to 500 horse power. They are nearly all of the impulse type and promise to become the preeminent driving power for centrifugal or turbine pumps. The field for this combination is considerable and covers centrifugal pumps, feed pumps, condenser equipments, and marine auxili- aries. Steam consumption in some cases is of importance, but not in others. Steam turbines of 200 to 500 horse power may be obtained having an economy equal to that of a reciprocating engine, and when the entire in- stallation is considered the steam turbine will show an advantage. APPENDIX. ELECTRICAL DATA. Full-load speeds for alternating-current motors based on 4 per cent slip. TABLE NO. 8. < Cycles. Number of poles. 25 27 30 33i 40 42 50 60 100 2 1440 1560 1730 1920 2300 2420 2880 4 720 780 865 960 1150 1210 1440 i725 2880 6 480 520 575 640 770 807 960 1150 1920 8 360 390 433 480 575 605 720 862 1440 10 290 310 345 385 460 485 575 690 1150 12 240 260 288 320 385 403 480 575 960 14 205 222 247 275 330 346 412 492 822 16 180 195 216 240 287 302 360 431 720 18 160 171 192 214 256 268 320 384 640 20 142 156 173 192 230 242 287 345 575 B.h.p. output for alternating-current motor = where volts X ampere X cos X ^n X .746 n = number of phases; = power factor of motor; m = motor efficiency. 169 170 CENTRIFUGAL PUMPING MACHINERY POINTS TO CONSIDER IN CENTRIFUGAL PUMP INSTALLATIONS. Floor Line to Discharge Level (Ft.): Should be given in open system. Pumping overboard or into a tank. Mine Service: Give information about discharge line in addi- tion to above. Gauge Reading at Discharge Nozzle (Lbs.): Should be given when pumping into a closed pressure system. (Heating System) or when pumping direct into city water main: Information about discharge line not needed. Gauge Reading at End of Discharge Line (Lbs.): Should be given when a certain pressure is required at a distance away from the pump ; for instance, in sprinkler systems and fire-hydrant systems. Infor- mation about discharge line is needed in addition to above. Head-on Suction: V///A Floor Line to Suction Level (Ft.): Should be given when suction water level is above floor line. =1 This is the case with all submerged pumps, also when pump = takes its water from an elevated tank, etc. Information about suction line is to be given in addition to above. Gauge Reading at Suction Nozzle (Lbs.): In closed system, when water enters under pressure; for instance, if connected to city main, in heating system, etc. Information in about suction line not needed. Suction Lift: Floor Line to Suction Level (Ft.): This is the most common case. Should always be given when pump takes its water from a well or from a river, provided pump is above water surface. Information about suction pipe needed. State if foot valve is provided. 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