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Online Encyclopedia
Originally appearing in Volume V14, Page 110 of the 1911 Encyclopedia Britannica.
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IRIII.Iu!111!I1Hi U U U 11 UU] a a its internal diameter 3 ft. 10 in. Normal speed 400 revs, per minute. Water is discharged into the wheel by a single nozzle, shown in fig. 202 with its regulating apparatus and some of the vanes. The water enters the wheel at an angle of 22° with the direction of motion, and the final angle of the wheel vanes is2o°. The efficiency on trial was from 75 to 78%. § 199. Theory of the Impulse Turbine.—The theory of the im- pulse turbine does not essen- tially differ from that of the re- action turbine, except that there is no pressure in the wheel oppos- ing the discharge velocity with which the water relative velocity v—u making an angle a with the direction of the vane's motion. Combining this with the velocity u of the vane, the absolute velocity of the water leaving the vane will bew = Bc. The component of w in the direction of motion of the vane is Ba = Bb -ab d c =u—(v—u) cos .a. Hence if Q is the quantity of water reaching the vane per second the change of momentum per second in the direction of the vane's motion is (G /g)[v—(uv u) cos a}] = ((iQ/g) (v —u) (I -i- c o s a). I f a= o°, cos a= I, and the -change of momentum per second, which is equal to the effort driving the vane, is P=2(GQ/g)(v—u). The work done on the vane is Pu=2(GQ/g)(v—u)u. If a series of vanes are inter-posed in succession, the quantity of water imping- ing on the vanes per second is the total discharge of the nozzle, and the energy expended at the nozzle is GQv'-/2g. Hence the efficiency of the arrangement is, when a=o°, neglecting friction, ,i=2Pu/GQv'=4(v—u)u/v2, which is a maximum and equal to unity if u= iv. In that case the whole energy of the jet is usefully expended in driving the series of vanes. In practice a cannot be quite zero or the water leaving one vane would strike the back of the next advancing vane. Fig. 203 shows a Pelton vane. The water divides each way, and leaves the vane on each side in a direction nearly parallel to the direction of motion of the vane. The best velocity of the vane is very approximately half the velocity of the jet. § 202. Regulation of the Pelton Wheel.—At first Pelten wheels were adjusted to varying loads merely by throttling the supply. This method involves a total loss of part of the head at the sluice or throttle valve. In addition as the working head is reduced, the relation between wheel velocity and jet velocity is no loner that of greatest efficiency. Next a plan was adopted of deflecting the jet so that only part of the water reached the wheel when the load was reduced, the rest going to waste. This involved the use of as equal quantity of water for large and small loads, but it had, what in some cases is an advantage, the effect of preventing any water hammer in the supply pipe due to the action of the regulator. In most cases now regulation is effected by varying the section of the jet. A conical needle in the nozzle can be advanced or withdrawn so as to occupy more or less of the aperture of the nozzle. Such a needle can be controlled by an ordinary governor. § 203. General Considerations on the Choke of a Type of Turbine.—The circumferential speed of any turbine is necessarily a fraction of the initial velocity of the water, and therefore is greater as the head is greater. In reaction turbines with complete admission the number of revolutions per minute becomes inconveniently great, for the diameter cannot be increased beyond certain limits without greatly reducing the efficiency. In impulse turbines with partial admission the diameter can be chosen arbitrarily and the number of revolutions kept down on high falls to any desired amount. Hence broadly reaction. turbines are better and less costly on low falls, and impulse turbines on high falls. For variable water flow impulse turbines have some advantage, being more efficiently regulated. On the other hand, impulse turbines lose efficiency seriously if their speed varies from the normal speed due to the head. If the head is very variable, as it often is on low falls, and the turbine must run at the same speed whatever the head, the impulse turbine is not suitable. Reaction turbines can be constructed so as to overcome this difficulty to a great extent. Axial flow turbines with vertical shafts have the disadvantage that in addition to the weight of the turbine there is an unbalanced water pressure to be carried by the footstep or collar bearing. In radial flow turbines the hydraulic pressures are balanced. The application of. turbines to drive dynamos directly has involved some new conditions. The electrical engineer generally desires a high speed of rotation, and a very constant speed at all times. The reaction turbine is generally more suitable than the impulse turbine. As the diameter of the turbine depends on the quantity of water and cannot be much varied without great inefficiency, a difficulty arises on low falls. This has been met by constructing four independent reaction turbines on the same shaft, each having ofcourse the diameter suitable for one-quarter of the whole discharge, and having a higher speed of rotation than a larger turbine. The turbines at Rheinfelden and Chevres are so constructed. To ensure constant speed of rotation when the head varies considerably without serious inefficiency, an axial flow turbine is generally used. It is constructed of three or four concentric rings of vanes, with independent regulating sluices, forming practically independent turbines of different radii. Any one of these or any combination can be used according to the state of the water. With a high fall the turbine of largest radius only is used, and the speed of rotation is less than with a turbine of smaller radius. On the other hand, as the fall decreases the inner turbines are used either singly or together, according to the power required. At the Zurich waterworks there are turbines of 90 h.p. on a fall varying from Io2 ft. to 41 ft. The power and speed are kept constant. Each turbine has three concentric rings. The outermost ring gives 90 h.p. with 1o5 cub. ft. per second and the maximum fall. The outer and middle compartments give the same power with 14o cub. ft. per second and a fall of 7 ft. lo in. All three compartments working together develop the power with about 250 cub. ft. per second. In some tests the efficiency was 74% with the outer ring working alone, 75.4% with the outer and middle ring working and a fall of 7 ft., and 80.7 % with all the rings working. § 204. Speed Governing.- vVhen turbines are used to drive dynamos direct, the question of speed regulation is of great importance. Steam engines using a light elastic fluid can be easily regulated by governors acting on throttle or expansion valves. It is different with water turbines using a fluid of great inertia. In one of the Niagara penstocks there are 400 tons of water flowing at io ft. per second, opposing enormous resistance to rapid change of speed of flow. The sluices of water turbines also are necessarily large and heavy. Hence relay governors must be v this opens an aperture a in. in-diameter, made in a brass screw plug b. The hole is reduced to 3I8 in. in diameter at the outer end of the plug and is closed by a small valve opening inwards. Through this, during the rebound after each stroke of the ram, a small quantity of air is sucked in which keeps the air vessel supplied with its elastic cushion of air. During the recoil after a sudden closing of the valve d, the pressure below it is diminished and the valve opens, permitting outflow. In consequence of the flow through this valve, the Water in the supply pipe acquires a gradually increasing velocity. The upward flow of the water, towards the valve d, increases the pressure tending to lift the valve, and at last, if the valve is not too heavy, lifts and closes it. The forward momentum of the column in, the supply pipe being destroyed by the stoppage of the flow, the water exerts a pressure at the end of the pipe sufficient to open the delivery valve o, and to cause a portion of the water to flow into the air vessel. As the water in the supply pipe comes to rest and recoils, the valve d opens again and the operation is repeated. Part of the energy of the descending column is employed in compressing, the air at the end of the supply pipe and expanding the pipe itself. This causes a recoil of the water which momentarily diminishes the pressure in the pipe below the pressure due to the statical head. This assists in opening the valve d. The recoil of the water is sufficiently great to enable a pump to be attached to the ram body instead of the direct rising pipe. With this arrangement a ram working with muddy water may be employed to raise clear spring water. Instead of lifting the delivery valve as in the ordinary ram, the momentum of the column drives a sliding or elastic piston, and the recoil brings it hack. This piston lifts and forces alternately the clear water through ordinary pump valves. Pumps § 206. The different classes of pumps correspond almost exactly to the different classes of water motors, although the mechanical details of the construction are somewhat e different. They are properly reversed water motors. Ordinary reciprocating pumps corre- spond to water-pressure engines. Chain and bucket pumps are in principle similar to water wheels in which the water acts by weight. Scoop wheels are similar to undershot water wheels, and centrifugal pumps to turbines. Reciprocating Pumps are single or double acting, and differ from water-pressure engines in that the valves are moved by the water instead of by automatic machinery. They may be classed thus:--- r. Lift Pumps.: The water drawn through a foot valve on the ascent of the pump bucket is forced through the bucket i valve when it descends, and lifted by the bucket when it reascends. Such pumps give an intermittent discharge. 2. Plunger or Force Pumps, in which the water drawn through the foot valve is displaced by the descent of a solid plunger, and forced through a delivery valve. They have the advantage that used, and the tendency of relay governors to hunt must be overcome. In the Niagara Falls Power House No. I, each turbine has a very sensitive centrifugal governor acting on a ratchet relay. The governor puts into gear, one or other of two ratchets driven by the turbine itself. According as one or the other ratchet is in gear the sluices are raised or lowered. By a subsidiary arrangement the ratchets are gradually put out of gear unless the governor puts them in gear again, and this prevents the over correction of the speed from the lag in the action of the governor. In the Niagara Power House No. 2, the relay is an hydraulic relay similar in principle, but rather more complicated in arrangement, to that shown in fig: 2o6, which is a governor used for the '250 h.p. turbines at Lyons. The sensitive governor G opens a valve and puts into action a plunger driven by oil pressure from an oil reservoir. As the plunger moves forward it gradually doses the oil admission valve by lowering the fulcrum end f of the valve lever which rests on a wedge w attached to the plunger. If the speed is still too high, the governor re-opens the valve. In the case of the Niagara turbines the oil pressure is 1200 lb per sq. in. One millimetre of movement of the governor sleeve completely opens the relay valve, and the relay plunger exerts a force of 50 tons. The sluices can be completely opened or shut in twelve seconds. The ordinary variation of speed of the turbine with varying load does not exceed 1%. If all the load is thrown off, the momentary variation of speed is not more than 5%. To prevent hydraulic shock in the supply pipes, a relief valve is provided which opens if the pressure is in excess of that due to the head. § ao5. The Hydraulic Ram. The hydraulic ram is an arrangement by which a quantity of water falling a distance h forces a portion of the water to rise to a height k1, greater than h. It consists of a supply reservoir (A, fig. 207), into which the water enters from some natural stream. A pipe s of considerable length conducts the water to a lower level, where it is discharged intermittently through a self-acting pulsating valve.. at d. The supply pipe s may be fitted with a flap valve for stopping the ram, and this is attached in some cases to a, float, so that the ram starts and stops itself automatically,; according as the supply cistern fills or empties. The lower float is just'sufficient 'to keep open the flap after it has been raised by the action of the upper float. The length of chain is adjusted so that the upper float opens the flap when the level in the cistern is at the desired height. If the water-level falls below the lower float the flap closes. The pipe s should be as long and as straight as possible, and as it is subjected to considerable pressure from the sudden arrest of the motion of the water, it must be strong and strongly jointed. a is an air vessel, and e the delivery pipe leading to the reservoir at a higher level than A, into which water is to be pumped. Fig. 208 shows in section the construction of the ram itself. d is the pulsating discharge valve already mentioned, which opens in-wards and downwards. The stroke of the valve is regulated by the cotter through the spindle, under which are washers by which the amount of fall can be regulated. At o is a delivery valve, opening outwards, which is often a ball-valve but sometimes a flap-valve. The water which is pumped passes through this valve into the air vessel a, from which it flows by the delivery pipe in a regular stream into the cistern to which the water is to be raised. In the vertical chamber behind the outer valve a small air vessel is formed, and into the friction is less than that of lift pumps, and the packing round the plunger is easily accessible, whilst that round a lift pump bucket is not. The flow is intermittent. 3. The Double-acting Force Pump is in principle a double plunger pump. The discharge fluctuates from zero to a maximum and back to zero each stroke, but is not arrested for any appreciable time. 4. Bucket and Plunger Pumps consist of a lift pump bucket combined with a plunger of half its area. The flow varies as in a double-acting pump. 5. Diaphragm Pumps have been used, in which the solid plunger is replaced by an elastic diaphragm, alternately depressed into and raised out of a cylinder. As single-acting pumps give an intermittent discharge ,three are generally used on cranks at 120°. But with all pumps the variation of velocity of discharge would cause great waste of work in the delivery pipes when they are long, and even danger from the hydraulic ramming action of the long column of water. An air vessel is interposed between the pump and the delivery pipes, of a volume from 5 to 100 times the space described by the plunger per stroke. The air in this must be replenished from time to time, or continuously, by a special air-pump. At low speeds not exceeding 30 ft. per minute the delivery of a pump is about 90 to 95 °,o of the volume described by the plunger or bucket, from 5 to 1o% of the discharge being lost by leakage. At high speeds the quantity pumped.occasionally exceeds the volume described by the plunger, the momentum of the water keeping the valves open after the turn of the stroke. The velocity of large mining pumps is about 140 ft. per minute, the indoor or suction stroke being sometimes made at 250 ft. per minute. Rotative pumping engines of large size have a plunger speed of 90 ft. per minute. Small rotative pumps are run faster, but at some loss of efficiency. Fire-engine pumps have a speed of 18o to 220 ft. per minute. The efficiency of reciprocating pumps varies very greatly. Small reciprocating pumps, with metal valves on lifts of 15 ft., were found by Morin to have an efficiency of 16 to 40%, or on the average 25%. When used to pump water at considerable pressure, through hose pipes, the efficiency rose to from 28 to J7%, or on the average, with 50 to 100 ft. of lift, about 50%. A large pump with barrels 18 in. diameter, at speeds under 6o ft. per minute, gave the following results: Lift in feet 143 34 47 Efficiency .46 •66 •70 The very large steam-pumps employed for waterworks, with 150 ft. or more of lift, appear to reach an efficiency of 90%, not including the friction of the discharge pipes. Reckoned on the indicated work of the steam-engine the efficiency may be 8o%. Many small pumps are now driven electrically and are usually three-throw single-acting pumps driven from the electric motor by gearing. It is not convenient to vary the speed of the motor to accommodate it to the varying rate of pumping usually required. Messrs Hayward Tyler have introduced a mechanism for varying the stroke of the pumps (Sinclair's patent) from full stroke to nil, without stopping the pumps. § 207. Centrifugal Pump.—For large volumes of water on lifts not exceeding about 6o ft. the most convenient pump is the centrifugal pump. Recent improvements have made it available also for very high lifts. It consists of a wheel or fan with curved vanes enclosed in an annular chamber. Water flows in at the centre and is discharged at the periphery. The fan may rotate in a vertical or horizontal plane and the water may enter on one or both sides of the fan. In the latter case there is no axial unbalanced pressure. The fan and its casing must be filled with water before it can start, so that if not drowned there must be a foot valve on the suction pipe. When no special attention needs to be paid to efficiency the water may have a velocity of 6 to 7 ft. in the suction and delivery pipes. The fan often has 6 to 12 vanes. For a double-inlet. fan of diameter D, the diameter of the inlets is D/2. If Q is the discharge in cub. ft. per second D = about o•6 v Q in average cases. Theand the disk is keyed on the driving shaft C. The casing K has a spirally enlarging discharge passage into the discharge pipe K. A cover L gives access to the pump. S is the suction pipe which opens into the pump disk on both sides at D. Fig. 210 shows a centrifugal pump differing from ordinary centrifugal pumps in one feature only. The water rises through a suction pipe S, which divides so as to enter the pump wheel W at the centre on each side. The pump disk or wheel is very similar to a turbine wheel. It is keyed on a shaft driven by a belt on a fast and loose pulley arrangement at P. The water rotating in the pump disk presses outwards, and if the speed is sufficient a continuous flow is maintained through the pump and into the discharge pipe D. The special feature in this pump is that the water, discharged by the pump disk with a whirling velocity of not inconsiderable magnitude, is allowed to continue rotation in a chamber somewhat larger than the pump. The use of this whirlpool chamber was first suggested by Professor James Thomson. It utilizes the energy due to the whirling velocity of the water which in most pumps is wasted in eddies in the discharge pipe. In the pump shown guide-blades are also added which have the direction of the stream lines in a free vortex. They do not therefore interfere with the action of the, water when pumping the normal quantity, but only prevent irregular motion. At A is a plug by which the pump case is filled before starting. If the pump is above the water to be pumped, a foot valve is required to permit the pump to be filled. Sometimes instead of the foot valve a delivery valve is used, an air-pump or steam jet pump being employed to exhaust the air from the pump case. § 208. Design and Proportions of a Centrifugal Pump.—The design of the pump disk is very simple. Let r, ro be the radii of the inlet and outlet surfaces of the putnp disk, d;, do the clear axial width at those radii. The velocity of flow through the pump may be taken peripheral speed is a little greater than the velocity due to the lift. Ordinary centrifugal pumps will have an efficiency of 40 to 6o%. The first pump of this kind which attracted notice was one exhibited by J. G. Appold in 1851, and the special features of his pump have been retained in the best pumps since constructed. Appold's pump raised continuously a volume of water equal to 1400 times its own capacity per minute. It had no valves, and it permitted the passage of solid bodies, such as walnuts and oranges, without obstruction to its working. Its efficiency was also found to be good. Fig. 209 shows the ordinary The pump disk and vanes B form of a centrifugal pump. are cast in one, usually of bronze, A the same as for a turbine. If Q is the quantity pumped, and H the lift, ui=0'25'/2 H. 2aridi = Q/ui. di=I.2ri . . ri ='2571 J (Q/J H). Usually ro = 2ri, and do=di or Id; according as the disk is parallel-sided or coned. The water enters the wheel radially with the velocity ui, and uo =Q/z,rrodo. (3) Fig. 211 shows the notation adopted for the velocities. Suppose the water enters the wheel with the velocity vi, while the velocity of the wheel is Vi. Completing the parallelogram, v,; is the relative velocity of the water and wheel, and is the proper direction of the wheel vanes. Also, by resolving, ui and wi are the cornponent velocities of flow and velocities of whir of the velocity vi of the water. At the outlet surface, vo is the final" velocity of discharge, and the rest of the notation is similar to that for the inlet surface. Usually the water flows equally in all directions in the eye of the wheel, in that case vi is radial. Then, in normal conditions of working, at the inlet surface, vi = ui w, =o (4) tan 9= u;AV; v,i = u; cosec e =Al u;2+Vi2 If the pump is raising less or more than its proper quantity, B will not satisfy the last condition, and there is then some loss of head in shock. At the outer circumference of the wheel or outlet surface, v,o = us cosec ¢ 1 wo =V —u, cot (5) vo=J luo +(Vo—uo cot s6)2} Variation of Pressure in the Pump Disk.—Precisely as in the case of turbines, it can be shown that the variation of pressure between the inlet and outlet surfaces of the pump is ho—hi = (Vo —Vi2)/2g — (14o2— vri2) /2g. Inserting the values of v, ,, v,i in (4) and (5), we get for normal conditions of workingho—hi = (Vo —Vi2)/2g—ua cosec2cp/2g+(ui2+V;2)/2g =Vo /2g—u 2 cosec 24/2g+1ii2/2g• (6) Hydraulic Efficiency of the Pump.—Neglecting disk friction, journal friction, and leakage, the efficiency of the pump can be found in the same way as that of turbines (§ 186). Let M be the moment of the couple rotating the pump, and a its angular velocity; wo, r° the tangential velocity of the water and radius at the outlet surface; wi, ri the same quantities at the inlet surface. Q being the discharge per second, the change of angular momentum per second is (GQ/g) (worn—wir;). Hence M = (GQ/g)(woro—wiri). In normal working, wi =o. Also, multiplying by the angular velocity, the work done per second is Ma = (GQ/g)woroa. But the useful work done in pumping is GQH. Therefore the efficiency is =GQH/Ma =gH/woroa =gH/woV0. (7) § 209. Case 1. Centrifugal Pump with no Whirlpool Chamber.—When no special provision is made to utilize the energy of motion of the water leaving the wheel, and the pump discharges directly into a chamber in which the water is flowing to the discharge pipe, nearly the whole of the energy of the water leaving the disk is wasted. The water leaves the disk with the more or less considerable velocity v,, and impinges on a mass flowing to the discharge pipe at the much slower velocity v,. The radial component of vs is almost necessarily wasted. From the tangential component there is a gain of pressure (w,' —v,2)/2g—(wo—v,)2/2g =v,(wo—v,)lg, which will be small, if v, is small compared with w,. Its greatest value, if v, =two, is ,w 2/2g, which will always be a small part of the whole head. Suppose this neglected. The whole variation of pressure in the pump disk then balances the lift and the head u; /2g necessary to give the initial velocity of flow in the eye of the wheel. Also in practice Hence, (1) (2) ui /2g+H Vo2/2g—uo2 cosec 24,/2g+ui2/2g, H =V 2/2g—uo cosec 24,/zg (8) or V. = J (2gH +uo2 cosec 2¢ . and the efficiency of the pump is, from (7), =gH/Vowo =gH/tV (Vo—no cot 0)J, _(V02—uo cosec 24,)lt2Vo(V—uocot0}, (9) For ¢=90°, (V02—u62)/2V02, which is necessarily less than 2. That is, half the work expended in driving the pump is wasted. By recurving the vanes, a plan introduced by Appold, the efficiency is increased, because the velocity vo of discharge from the pump is diminished. If q5 is very small, cosec ¢ = cot ¢; and then = (V o+uo cosec ds) /2Vo, which may approach the value 1, as ¢ tends towards o. Equation (8) shows that us cosec 4, cannot be greater than Vs. Putting U. =0'254 (2gH) we get the following numerical values of the efficiency and the circumferential velocity of the pump:— 4, 77 V 900 0.47 I.0342gH 45,0 0.56 i.o6 „ 30° 0.65 I'I2 200 0.73 1.24 Io° 0.84 1'75 ,, ¢ cannot practically be made less than 20°; and, allowing for the frictional losses neglected, the efficiency of a pump in which ¢=2o° is found to be about •6o. § 210. Case 2. Pump with a Whirlpool Chamber, as in fig. 210.—Professor James Thomson first suggested that the energy of the water after leaving the pump disk might be utilized, if a space were left in which a free vortex could be formed. In such a free vortex the velocity varies inversely as the radius. The gain of pressure in the vortex chamber is, putting re, r„ for the radii to the outlet surface of wheel and to outside of free vortex, jjI — to '2\ r I 2g\ r I -2g \ —-k2 if k = ra/r°,. The lift is then, adding this to the lift in the last case, H = { V 2—u02 cosec24+vo (1-k2))/2g. But sot=V, 2Vouo cot (19+142 cosec'4); .'.H ={(2 —k2)V. -2kVouo cot eb —k2u 2 cosec2¢)/2g. (to) Putting this in the expression for the efficiency, we find a considerable increase of efficiency. Thus with 0=9o° and k=2, n=8 nearly, 43 a small angle and k = , =1 nearly. With this arrangement of pump, therefore, the. angle at the outer ends of the vanes is of comparatively little importance. A moderate angle of 30° or 40° may very well be adopted. The following numerical values of the velocity of the circumference of the pump have been obtained by taking k= and 14=0-254 (2gH). 4) V. 90° •76242gH 45° .842 30° '911 „ 2o° 1.023 The quantity of water to be pumped by a centrifugal pump necessarily varies, and an adjustment for different quantities of water can-not easily be introduced. Hence it is that the average efficiency of pumps of this kind is in practice less than the efficiencies given above. The advantage of a vortex chamber is also generally neglected. The velocity in the supply and discharge pipes is also often made greater than is consistent with a high degree of efficienei. Velocities of 6 or 7 ft. per second in the discharge and suction pipes, when the lift is small, cause a very sensible waste of energy; 3 to 6 ft. would be much better. Centrifugal pumps of very large size have been constructed. Easton and Anderson made pumps for the North Sea canal in Holland to deliver each 67o tons of water per minute on a lift of 5 ft. The pump disks are 8 ft. diameter. J. and H. Gwynne constructed some pumps for draining the Ferrarese Marshes, which together deliver 2000 tons per minute. A pump made under Professor J. Thomson's direction for drainage works in Barbados had a pump disk 16 ft. in diameter and a whirlpool chamber 32 ft. in diameter. The efficiency of centrifugal pumps.when delivering less or more than the normal quantity of water is discussed in a paper in the Proc. Inst. Civ. Eng. vol. 53. § 211. High Lift Centrifugal Pumps.—It has long been known that centrifugal pumps could be worked in series, each pump overcoming a part of the lift. This method has been perfected, and centrifugal pumps for very high lifts with great efficiency have been used by Sulzer and others. C. W. Darley (Proc. Inst. Civ. Eng., supplement to vol. 154, p. 156) has described some pumps of this new type driven by Parsons steam turbines for the water supply of Sydney, N.S.W. Each pump was designed to deliver r million gallons per twenty-four hours against a head of 240 ft. at 3300 revs. per minute. Three pumps in series give therefore a lift of 720 ft. The pump consists of a central double-sided impeller 12 in. diameter. The water entering at the bottom divides and enters the runner at each side through a bell-mouthed passage. The shaft is provided with ring and groove glands which on the suction side keep the air out and on the pressure side prevent leakage. Some water from the pressure side leaks through the glands, but beyond the first grooves it passes into a pocket and is returned to the suction side of the pump. For the glands on the suction side water is supplied from a low-pressure service. No packing is used in the glands. During the trials no water was seen at the glands. The following are the results of tests made at Newcastle: I. II. III. IV. Duration of test hours 2 1.54 1.2 1.55 Steam pressure lb per sq. in. 57 57 84 55 Weight of steam per water h.p. hour lb 27.93 30.67 28.83 27.89 Speed in revs. per min. 3300 3330 3710 3340 Height of suction . . ft. II II II II Total lift . . . ft. 762 744 917 756 Million galls. per day pumped By Venturi meter 1.573 P499 1.689 1.503 1.623 1.513 1.723 P555 By orifice . . . . Water h.p. . . . . 252 235 326 239 In trial IV. the steam was superheated 95° F. From other trials under the same conditions as trial I. the Parsons turbine uses 15.6 lb of steam per brake h.p. hour, so that the combined efficiency of turbine and pumps is about 56%, a remarkably good result. § 212. Air-Lift Pumps.—An interesting and simple method of pumping by compressed air, invented by Dr J. Pohle of Arizona, is likely to be very useful in certain cases. Suppose a rising main placed in a deep bore hole in which there is a considerable depth of water. Air compressed to a sufficient pressure is conveyed by an air pipe and introduced at the lower end of the rising main. The air rising in the main diminishes the average density of the contents of the main, and their aggregate weight no longer balances the pressure at the lower end of the main due to its submersion. An up-ward flow is set up, and if the air supply is sufficient the water in the rising main is lifted to any required height. The higher the lift above the level in the bore hole the deeper must be the point at which air is injected. Fig. 212 shows an air-lift pump constructed for W. H. Maxwell at the Tunbridge Wells water-works. There is a two-stage steam air compressor, compressing air to FIG. 2I2. from go to loo lb per sq. in. The bore hole is 330 ft. deep, lined with steel pipes 15 diameter for 200 ft. and with perforated pipes 131 in. diameter for the lower 150 ft. The rest level of the water is 96 ft. from the ground-level, and the level when pumping 32,000 gallons per hour is r 20 ft. from the ground-level. The rising main is 7 in. diameter, and is carried nearly to the bottom of the bore hole and to 20 ft. above the ground-level. The air pipe is 22 in. diameter, In a trial run 31,402 gallons per hour were raised 133 ft. above the level in the well. Trials of the efficiency of the system made at San Francisco with varying conditions will be found in a paper by E. A. Rix (Journ. Amer. Assoc. Eng. Soc. vol. 25, Steel Tubes /5'.Diam. Rising Main 7Diam. Air Pipe 2#'Qiam Igoo). Maxwell found the best results when the ratio of immersion calculated on the work of compression only. It is zero for no disto lift was 3 to I at the start and 2.2 to r at the end of the trial. In these conditions the efficiency was 37% calculated on the indicated h.p. of the steam-engine, and 46% calculated on the indicated work of the compressor. 2.7 volumes of free aar were used to x of water lifted. The system is suitable for temporary purposes, especially as the quantity of water raised is much greater than could be pumped by any other system in a bore hole of a given size. It is useful for clearing a boring of sand and may be advantageously used permanently when a boring is in sand or gravel which cannot be kept out of the bore hole. The initial cost is small. § 213. Centrifugal Fans.—Centrifugal fans are constructed similarly to centrifugal pumps, and are used for compressing air to pressures not exceeding 10 to 15 in. of water-column. With this small variation of pressure the variation of volume and density of the air may be neglected without sensible error. The conditions of pressure and discharge for fans are generally less accurately known than in the case of pumps, and the design of fans is generally somewhat crude. They seldom, have whirlpool chambers, though a large expanding outlet is provided in the case of the important Guibal fans used in mine ventilation. It is usual to reckon the difference of pressure at the inlet and outlet of a fan in inches of water-column. One inch of water-column =64•4 ft. of air at average atmospheric pressure =5.21b per sq. ft. Roughly the pressure-head produced in a fan without means of utilizing the kinetic energy of discharge would be v2/2g ft. of air, or 0.00024 v2 in. of water, where v is the velocity of the tips of the fan blades in feet per second. If d is the diameter of the fan and t the width at the external circumference, then rdt is the discharge area of the fan disk. If Q is the discharge in cub. ft. per sec., u =Q/rdt is the radial velocity of discharge which is numerically equal to the discharge per square foot of outlet in cubic feet per second. As both the losses in the fan and the work done are roughly proportional to u2 in fans of the same type, and are also proportional to the gauge pressure p, then if the losses are to be a constant percentage of the work done u may be taken proportional to sl p. In ordinary cases u = about 2211 p. The width t of the fan is generally from 0.35 to 0.45d. Hence if Q is given, the diameter of the fan should be: For t=o•35d, d =0.2011 (Q/,l p) For t=o•45d, d=o•i811 (Q/'p) If p is the pressure difference in the fan in inches of water, and N the revolutions of fan, v=rdN/6o ft. per sec. N = 12301/ p/d revs. per min. As the pressure difference is small, the work done in compressing the air is almost exactly 5•2pQ foot-pounds per second. Usually, however, the kinetic energy of the air in the discharge pipe is not inconsiderable compared with the work done in compression. 'If w is the velocity of the air where the discharge pressure is measured, the air carries away w2/2g foot-pounds per lb of air as kinetic energy. In Q cubic feet or o•o8o7QIb the kinetic energy is 0.00125 Qw2 foot-pounds per second. The efficiency of fans is reckoned in two ways. If B.H.P. is the effective horse-power applied at the fan shaft, then the efficiency reckoned on the work of compression is n = 5.2pQ/55oB. H. P. On the other hand, if the kinetic energy in the delivery pipe is taken as part of the useful work the efficiency is n2 = (5.2PQ+o•oo125Qw2)/550B.H.P. Although the theory above is a rough one it agrees sufficiently with experiment, with some merely numerical modifications. An extremely interesting experimental investigation of the action of centrifugal fans has been made by H. Heenan and W. Gilbert (Prot. Inst. Civ. Eng. vol. 123, p. 272). The fans delivered through an air trunk in which different resistances could be obtained by introducing diaphragms with circular apertures of different sizes. Suppose a fan run at constant speed with different resistances and the compression pressure, discharge and brake horse-power measured. The results plot in such a diagram shown in fig. al. The less the resistance to discharge, that is the larger the opening In the air trunk, the greater the quantity of air discharged at the given speed of the fan, On the other hand the compression pressure diminishes. The eltrve marked total gauge is the compression pressure+the velocity head in the discharge pipe, both in inches of water. This curve falls, bgtuot nearly so much as the compression curve, when the resistance a the air trunk is diminished. The brake horse-power increases as tiffs resistance is diminished because the volume of discharge in-creases very much. The curve marked efficiency is the efficiencycharge, and zero also when there is no resistance and all the energy given to the air is carried away as kinetic energy. There is a discharge for which this efficiency is a maximum; it is about half the discharge which there is when there is no resistance and the delivery pipe is full open. The conditions of speed and discharge corresponding to the greatest efficiency of compression are those ordinarily taken as the best normal conditions of working. The curve marked total efficiency gives the efficiency calculated on the work of compression and kinetic energy of discharge. Messrs Gilbert and Heenan found the efficiencies of ordinary 'fans calculated on the compression to be 40 to 6o % when working at about normal conditions. Taking some of Messrs Heenan and Gilbert's results for ordinary fans in normal conditions, they have been found to agree fairly with the following approximate rules. Let pa be the compression pressure and q the volume discharged per second per square foot of outlet area of fan. Then the total gauge pressure due to pressure of compression and velocity of discharge is approximately: p=pa-l-o•0004g2 in. of water, so that if pc is given, p can be found approximately. The pressure p depends on the circumferential speed v of the fan disk- p=0.00o25v2 in. of water v=63dp ft. per sec. The discharge per square foot of outlet of fan is--q=15 to 18s/p cub. ft.`per sec. The total discharge is Q =rdtq =47 to 56 Tits/ p For t=.35d, d=o•22 to 0.251/ (Q/1Ip) ft. t =.45d, d =0.20 to 0.221/ (Q/1/ p) ft. N =12031f p/d. These approximate equations, which are derived purely from experiment, do not differ greatly from those obtained by the rough theory given above. The theory helps to explain the reason for the form of the empirical results. (W. C. U.)
End of Article: IRIII

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