Pumps

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Information about Pumps

Published on December 23, 2008

Author: jacobrajeev

Source: slideshare.net

14: Pumps Centrifugal pumps 418 How to estimate the head for an average centrifugal pump 455 Speed torque calculation 430 Find the reciprocating pump capacity 455 How to estimate the Kp required to pump at a given Pulsation Control for Reciprocating Pumps 431 rate at a desired discharge pressure 455 Rotary pumps on pipeline services 439 Nomograph for determining reciprocating pump capacity 456 Key Centrifugal Pump Parameters and How Nomograph for determining specific speed of pumps 457 Nomograph for determining horsepower requirement of pumps.... 458 They Impact Your Applications—Part 1 444 How to select motors for field-gathering pumps 458 Key Centrifugal Pump Parameters and How Reciprocating pumps 459 Understanding the basics of rotary screw pumps 468 They Impact Your Applications—Part 2 450 Estimate the discharge of a centrifugal pump at various speeds 454

Centrifugal pumps Centrifugal pumps are best suited for large volume The first step in selecting centrifugal pumps is to analyze applications or for smaller volumes when the ratio of the pipeline system and determine its characteristics such as volume to pressure is high. The selection of the proper the pressure required to move the desired flow rate. Initial pump will depend on the system throughput, viscosity, and future conditions should be evaluated so that pumps specific gravity, and head requirements for a particular with sufficient flexibility to handle both conditions with application. Where a variety of products with different minor changes may be selected. Refer to Section 13—Liquid characteristics are handled, it may be necessary to utilize Hydraulics for data on calculating pipeline pressure drop. multiple pumps either in series or parallel, or a series- Pump manufacturers publish pump performance maps parallel combination, or pumps with variable speed capabil- similar to the one shown in Figure 1 to aid in pump ities to achieve the desired throughput at least cost. selection. The performance map shows the range of Economics will ultimately determine the type and number differential head and capacity that will be developed by a of pumps to be used. family of pumps. Once the system data have been ONE STAGE CAPACITY - Gallons per minute PUMP PERFORMANCES AT 3600 RPM UNLESS THESE PERFORMANCES ARE NOMINAL USE FOR OTHERWISE NOTED PRELIMINARY SELECTION ONLY DIFFERENTIAL HEAD IS SHOWN FOR Ont Stoa* Figure 1. Typical performance map for centrifugal pumps. Courtesy United Centrifugal Pumps.

NPSH in Feet NPSH for Water @ £ Impeller Total Differential Head in Feet Per Stage BHP/STG 11quot; 1st STG - 1 0 quot; Series Gallons per Minute une*«ige Figure 2. Typical performance curve for centrifugal pumps. Courtesy United Centrifugal Pumps. established, a pump may be selected from the map. The material selection will be determined by the system selection is further refined by referring to a curve such as requirements and the fluid to be pumped. that shown in Figure 2, which shows a typical pump It is extremely important that sufficient net positive manufacturer's catalog curve for a specific pump showing suction head (NPSH) above the vapor pressure of the the performance for various impeller size, as well as the liquid being pumped is available for safe and reliable pump brake horsepower, NPSH requirement, and pump effi- operation. NPSH is always specified in feet of head above ciency. Pump performance curves are plotted in feet and for vapor pressure required at the center line of the impeller. a specific gravity of 1.0. When pumping a nonviscous liquid, The NPSH required will increase as the capacity of the the pump will develop the same head shown on the pump increases, which makes it necessary for the user to performance curve; however, the pressure will change with make a best effort to furnish the pump manufacturer with an specific gravity, as will the horsepower required to drive the accurate range of flow rate conditions. The pump suction pump. Pressure may be calculated by multiplying the pump nozzle geometry, pump speed, and the impeller inlet will differential head by the specific gravity and by 0.4329. The determine the NPSH that will be required. If too much horsepower required to drive the pump will also vary directly margin of safety is included in the pump specifications, the with the specific gravity of the fluid being pumped. pump will probably be oversized to obtain a low NPSH and The viscosity of the fluid will influence the performance of will not operate at its best efficiency point. Figures 8, 9, and the centrifugal pump. This is discussed in more detail later. 10 show three examples of how to calculate available NPSH. Centrifugal pumps are available in a number of different As a rule of thumb, the NPSH available should be at least configurations such as horizontal, vertical, radial split, and 10% greater than the NPSH required by the pump. It is axial split casings. The final choice will depend on the system difficult and costly to correct NPSH problems after the hydraulic conditions (pressure and flow rate), the ease of pump is placed in service. NPSH is discussed in detail in maintenance desired, and amount of space available for Centrifugal Pumps: Design and Application by Ross and installing the pump. The pressure rating of the case and Lobanoff, Gulf Publishing Co., and it's recommended

Head Ef f. BHP NPSH Differential Head - Ft. Efficiency - % Gallons Per Minute Figure 3. Centrifugal pump performance data. reading for those who are involved in the selection and impeller diameter was chosen from the curve shown in application of centrifugal pumps. Figure 2. In this case, an impeller diameter of 11-in. was Table 1 gives some conversion factors and formulas that chosen. will be useful in problems dealing with pumps. Centrifugal pumps may be operated singly, in series, in AU centrifugal pumps, except those pumps that are parallel, or in a combination of series and parallel. When specifically designed for low suction pressure applications pumps are operated in series, the heads are additive at a such as tank farm booster pumps, should be protected by a given rate of flow. When pumps are operated in parallel, low suction pressure shut down switch. The value of flow is additive at a given value of head. This is illustrated in pressure to be used to set the switch will be determined Figure 3. When pumps are operated in a series-parallel by the amount of NPSH required at the maximum expected combination, these same rules will apply to determine the operating flow rate, the vapor pressure of the fluid being total head developed by the combination of pumps. Figure 4 pumped, and the atmospheric pressure at the pump location. shows performance data for one pump and for series and Once the maximum expected operating flow rate is known, parallel operation of two pumps. the amount of NPSH required may be obtained from the Pumps connected in series have the discharge of one pump performance curve. The NPSH curve is plotted in feet pump connected to the suction of the downstream of head, and this may be converted to psi for switch setting pump. Pumps connected in parallel use a common purposes as follows: suction line and common discharge line. (See Figure 4a.) Pumps that are to be operated in parallel must be PSIG = NPSH x 0.433 x specific gravity matched so that both develop the same head at the + vapor pressure — atmospheric pressure same flow rate. They should also have constantly rising head curves. Pumps with a head curve that has a hump Figure 3 shows the performance data for a centrifugal (one that develops less head at shutoff than at some pump that was chosen from the performance map. The other point on the head curve) should be avoided since

Table 1* Conversion factors and formulas specific gravity where N = speed in rpm N3 = specific speed in rpm S = suction specific speed in rpm Q = capacity in gpm specific gravity P = pressure in psi H = total head in ft H1 = head per stage in ft hsv = net positive suction head in ft hv = velocity head in ft whp = water horsepower bhp = brake horsepower U = peripheral velocity in ft per second g = 32.16 ft per second (acceleration of gravity) mgd = million gallons per day cfs = cubic ft per second bbl = barrel (42 gallons) C = specific heat psi = Ib/in.2 gpm = gallons per minute e = pump efficiency in decimal D = impeller diameter in in. V = velocity in ft per second T = torque in ft-lb t = temperature in 0 F tr = temperature rise in 0 F A = area in in.2 *Courtesy United Centrifugal Pumps. One Pump 2 Pumps Series 2 Pumps Parallel a Pumps in Series Head - Feet 2 Pumps in Parallel O D E Pmnp Gallons Per Minute Figure 4. Series/parallel combinations—centrifugal pumps.

power. If as many as three series units can be justified, the units should be sized as follows: One unit-^ull head One unit—/4 of the head of the full head unit One unit—% of the head of the full head unit This will allow the following pump combinations to be used: Units Total head Units Total head SUCTION DISCHARGE Figure 4a. Pumps connected in parallel/in series. this could possibly preclude the second pump from pumping. Increments of head can be achieved up to the last 1A The high cost of energy will in almost all cases justify the head increment. This results in maximum flexibility when additional cost of multiple pumping units or variable speed constant rpm drivers are used. drivers. The use of multiple units will allow most pumping Figure 5 shows the performance of series versus parallel conditions to be met without throttling and therefore wasting operation of pumps. In this particular example, series One Pump 2 Pumps Series 2 Pumps Parallel Sy stem Cur/e System Curve 2 Pumps in Series Head - Feet 2 Punps in Parallel nnp PM np Gallons Per Minute Figure 5. Series/parallel combinations with system curve—centrifugal pumps.

3 (Parallel) 3 (Series) PRESSURE DROP - PSIG 1 (Parallel) 2 (Series) Sys. Friction 1 (Series) FLOW RATE - B/D Figure 5a. Series and parallel operation—centrifugal pumps. operation will provide a higher pumping rate than provided The use of multiple pumps affords some protection with a parallel operation, and two pumps in parallel will against throughput reductions due to one or more pumps pump only slightly more than one single pump. The being out of service. maximum flow rate with two pumps in series will be Parallel pump operation is usually best suited for systems 2,594 gpm or 89,439 b/d. The maximum flow rate for one where the residual static pressure is high compared to the pump will be 1,917 gpm or 66,092 b/d. line friction loss. Such systems usually have a flat system For system rates between 66,092 b/d and 89,439 b/d, curve. An example of this is shown in Figure 5a. pressure throttling will be necessary to match the pump For this system, three parallel pumps would offer the most performance to the system curves, unless the pumps are flexibility, since one unit will always be able to overcome the equipped with variable speed drives. If throttling is necessary, system friction and static and would pump almost as much as it may be minimized by operating at a high flow rate for part two series pumps. One could make the argument that one of the day and at a lower rate for the balance of the day. If a large pump could be selected to meet the 2 (series) condition system rate of 70,000 b/d is desired with the system shown in shown in Figure 5a and a V^-head pump could be used for Figure 5, one pump operating 20 hours per day and two the maximum volume condition. This argument is valid as pumps operating 4 hours per day would achieve this rate; long as the large pump is available. When it is not available, however, this may not be the most economical method of the system will be down, since the ^-head unit will not operation. It will be necessary to calculate and compare the develop enough head to overcome the static and friction cost of throttling pressure versus the cost to operate the losses. Three pumps operating in parallel will offer the most second unit for 4 hours. System rates less than 66,092 b/d will flexibility for the system shown in Figure 5a. require pressure throttling unless the pump speed may be Changes in system throughputs may require resizing the varied. It is not uncommon for pipeline pump stations to pump impellers to meet the new conditions. Some types of contain several pumps, each of which may be sized multi-stage pumps may be destaged or upstaged by adding differently to increase the amount of flow rate flexibility. or taking out impellers, depending upon the number of

Table 2* pump impeller(s) or adding volute inserts. The pump Formulas for recalculating pump performance with manufacturer should be contacted for details on how to impeller diameter or speed change accomplish this. The pump affinity laws shown in Table 1 may be used to 0 , , H , , b h p , , D, and N 1 » Initial Capacity, Head, Brake Horsepower, predict the performance of the pump at different operating Diameter, and Speed. conditions. These laws are commonly used to determine new Q2 . H 2 . bhp2 , D2 and N 2 * New Capacity, Head. Brake Horsepower, impeller diameters required for nominal changes in flow Diameter, and Speed. rates or performance at different speeds. The affinity laws Diameter Chonge Only Speed Change Only Diameter 8 Speed Change will allow the translation of a point on the performance curves but will not define the operation of the pump in the * • * ( & • ) « • • * ( * ) * • « ( * • » system. A new pump performance curve may be constructed using these laws and then applied to the system curve to quot;••quot;•©; quot;••••& +•>&•& determine the actual operation of the pump. bhp2 «bhp, (B^) bhp2 -bhp, ( ^ bhp2 - b h p , ^ x ^-J The system shown in Figure 5, with both pumps configured with 11-in. diameter impellers, will pump a * Courtesy United Centrifugal Pumps. maximum of 89,439 b/d. The system throughput require- ments have now changed such that the maximum desired throughput is 87,931 b/d, and it is desired to trim the impellers in one of the pumps to meet this new condition. stages in operation and the amount of head change required. The new H-Q condition for one pump will be H 2 = 955 ft If it is contemplated that it may be necessary in the future to and Q 2 = 2,550 gpm. In order to establish the new impeller destage or upstage, this should be specified when the pump diameter, it will be necessary to translate this point to the is purchased. It is also possible to optimize the efficiency existing pump curve to determine Hi and Qi. Select flow point by chipping the pump volutes and under-filing the rates on both sides of the new rate and calculate the 11-iich dia D i f f e r e n t i a l Head - Ft. N W Opsratiig Point 2 Gallons Per Minute Figure 6. Impeller sizing—centrifugal pumps.

corresponding head. These data are plotted and the resulting Since the pump affinity laws are not quite exact for a line will cross the existing pump head curve as shown in diameter change, it will be necessary to apply a correction Figure 6. Determine the head and flow rate at the point factor to the calculated value of D 2 to determine the final where this line crosses the existing head curve, and these diameter. The correction factor is determined by calculating values will be Hi and Qi. For this calculation, flow rates the ratio of the calculated diameter to the original diameter. were selected as follows: This value, expressed as a percentage, is then used to enter Figure 7 to determine the actual required diameter as a Selected Q Calculated H percentage of the original diameter. 2,800 1,151 2,600 993 D calc/D orig = 10.34/11 = 94% 2,400 846 The head was calculated as follows: Enter the chart with 94% and read 94.5% as the actual 2 required diameter as a percentage of the original diameter. H = H 2 x(Q/Q 2 ) The new diameter will therefore be 0.945x11.0 = This line crosses the existing head curve at 2,700 gpm and 10.395 in. The impellers would probably be trimmed to 1,080 ft, and these values will be the values to use for Qi and 107Z16 -in. Hi, respectively. Once the new diameter is established, the new pump head Use the pump affinity law to calculate the new diameter as curve may be calculated by using the affinity laws to follows: calculate new H-Q values. The pump affinity laws are exact for a speed change. H 2 /Hi = (D 2 /Di) 2 Figure 7a shows the effect of a speed change on the pump D 2 = D 1 XUVH 1 ) 1 7 2 performance. The pump manufacturer should be consulted D 2 = 11.0 x(955/l,080) 1/2 for data on the range of speeds over which the pump may be D 2 = 10.34 in. efficiently operated. CaIc. Reqd. Dia. % of Orig. Actual Required Dia. % of Orig Figure 7. Impeller trim correction.

H-Q 4000 RPM Differential Head - Ft. H-Q 3550 RPM Efficiency - % EFF. 3550 RPM EFF. 4000 RPM BHP 4000 RPM BHP 3550 RPM Gallons Per Minute Figure 7a. Centrifugal pump performance data for speed. Very viscous liquids will have an impact on the performance of centrifugal pumps. Figure 11 compares NPSH CALCULATION pump performance when pumping water to pumping a FOR SUCTION LIFT liquid with a viscosity of 6,000 SUS. The effects of viscosity LINE LOSS on pump performance may be predicted by the use of SPECIFIC GRAVlTYOF WATER=IO Figures 12 and 13. These charts are used as follows. Refer to Figure 11 and select the water capacity at bep (best efficiency point). In this example, the water capacity at bep NPSH ABSOLUTE AVAILABLE ~ PRESS1FT VAPOR _ LINE + DIFFERENCE PRESS FT. LOSS1FT quot; I N ELEV 1 FT is 23,000 gpm. Refer to Figure 12 and enter the chart with 23,000 gpm and move vertically to intersect the desired ATMOSPHERE impeller diameter, 27in., and then horizontally to intersect 14.7 PSIA WHERE PRESSURE IN FEET = (PRESSURE, PSIAX23I) (SPECIFIC GRAVITY) the 6,000SUS line. Move vertically to intersect the line representing the water efficiency (85%), then horizontally to NPSH AVAILABLE read the viscous efficiency 62% (not a correction factor). The 0.5 PSIA previous vertical line is extended to intersect the head (WATER AT 80° F) FEET correction and capacity correction curves. The capacity correction factor is 93%, and the head correction factor is 95%. The viscous head at shutoff will be the same as the NPSH AVAILABLE MUST BE GREATER THAN shutoff head for water. The viscous head will be 617 ft at NPSH REQUIRED BY THE PUMP 21,400 gpm. Plot this point and then draw the head-capacity Figure 8. NPSH calculation for suction lift.

GAUGE READING 37 5 PSI NPSH CALCULATION FOR GAUGE READING IO PSI NPSH CALCULATION LIQUID AT BOILING POINT FOR PRESSURED DRUM SPECIFIC GRAVITY OF N-BUTANE AT 100° F = 0 5( SPECIFIC GRAVITY OF WATER=IO A8S0LUTE PRESSURE = GAUGE PRESSURE + ATMOSPHERIC PRESSURE ABSOLUTE PRESSURE » GAUGE PRESSURE + ATMOSPHERIC PRESSURE = GAUGE PRESSURE + 14 7 -GAUGE PRESSURE + 14.7 GAS NPSH = ABSOLUTE _ VAPOR _ LINE +DIFFERENCE NPSH m ABSOLUTE _ VAPOR _ LINE +DIFFERENCE PRESSURE AVAILABLE PRESS1FT PRESS1FT LOSS,FT. ~ IN ELEV 1 FT. AVAILABLE PRESS..FT PRESS1FT LOSS,FT quot; IN ELEV, FT WHERE PRESSURE IN FEET = (PRESSURE , PSIA) (2 31) AIR PRESSURE (SPECIFIC GRAVITY) WEE P E S R , F E E T - ^ ^ . g HR RSU E 52 2 PSIA (N- BUTANE NPSH NPSH AVAILABLE 0.5 PSIA AVAILABLE (WATER AT 800F) NPSH AVAILABLE MUST NPSH AVAILABLE MUST BE GREATER THAN BE GREATER THAN LINE LOSS NPSH REQUIRED BY THE PUMP NPSH REQUIREDBY THE PUMP LINE LOSS Figure 9. NPSH calculation for liquid at boiling point. Figure 10. NPSH calculation for pressured drum. Water Eff. Water Head Ft. Vise. Head D i f f . Head - Efficiency - % Vise. Eff. GPM X 1. 0 0 0 Figure 11. Effects of viscosity on centrifugal pump performance.

Viscous Correction Factors Capacity - Head - Percent (CH) (CQ) Viscous Efficiency - Percent (Not a correction Factor) (Ev) Centistokes Impeller DIA - Inches Water Capacity in 1000 GPM (at BEP) Figure 12. Viscosity correction chart—capacities above 10,000GPM. Courtesy United Centrifugal Pumps.

Capacity and Efficiency Head Correction Factors Centistokes Head in Feet (First Stage) Viscosity-SSU Capacity in 100 GPM CORRECTION APPLIES TO PEAK EFFICIENCY ONLY. FOR OTHER CAP DRAW CORRECTED H.R CURVE PARALLEL TO WATER H.R CURVE AND CALCULATE EFFICIENCIES BASED ON ASSUMED HEAD - CAP CURVE. ADJUST AS REQUIRED TO OBTAIN A SMOOTH CURVE. USE CORRECTIONS OBTAINED FOR MAXIMUM DIAMETER FOR CUT DIAMETERS. Figure 13. Viscosity correction chart—capacities up to 10,000GPM. Courtesy United Centrifugal Pumps. curve through this point. The viscous brake horsepower Figure 13 is used for pumps with capacities up to is calculated at the bep (62% at 21,400 gpm). Draw 10,000 gpm, and Figure 12 is used for pumps with capacities the viscous brake horsepower curve through this point. greater than 10,000 gpm. The slope of the line will follow the slope of the water brake There may be instances where it is necessary to horsepower curve. The viscous efficiency will be 62% at calculate the pump speed torque requirements to deter- 21,400 gpm, and the viscous efficiency curve may now be mine if the pump driver will be able to accelerate the calculated using the viscous brake horsepower, viscous head pump to operating speed. Example 1 illustrates this and capacity, and the viscous specific gravity, which in this calculation procedure, and the results are shown in example is 0.9. Figure 14.

Speed torque calculation The formula for calculating the speed-torque relation for a pump with a given specific gravity and viscosity is: 5 2 5 0 x H P Torque V H (lb-ft.) = RPM The following data is required to plot the speed-torque curve: 1.HP@Shutoff. 2. HP @ Open Valve (use one of the following): 2a. HP @ 120% of peak capacity. 2b. HP rating of driver if greater than HP @ 120% of peak capacity. 2c. Depending on pump application maximum horsepower on performance curve which is greater than dri- ver rated horsepower. NOTE: Using quot;2a,quot; or quot;2bquot; will cover most of the applications. However on some applications quot;2cquot; must be considered. 3. Pump RPM. 4. Torque @ Shutoff. 5. Torque @ quot;2a,quot; quot;2bquot; or quot;2cquot;. From the above data the torque with the valve closed (using HP @ Shutoff), and the torque with the valve open (using HP @ quot;2a,quot; quot;2bquot; or quot;2cquot;) at full speed can be computed. The torque varies as the square of the speed; therefore to obtain the torque at: 3/4 Speed—multiply full speed torque by—0.563 1/2 Speed—multiply full speed torque by—0.250 1/4 Speed—multiply full speed torque by—0.063 1/8 Speed—multiply full speed torque by—0.016 From the figures a speed-torque curve for both closed and open valve, covering the complete speed range, can be plotted. Example HP @ Shutoff = 80 Torque @ Shutoff = 5250 x 80 _ 1 1 8 j b _ft @ 3 5 5 0 R P M 3550 HP @ 120% of peak capacity = 150 Pump RPM = 3550 Torque @ 120% of peak cap. = 5250x150 = 2 2 2 |b _ft @ 3 5 5 0 RPM 3550 Speed-RPM Torque-Valve Open (150 HP) Torque-Valve Closed (80 HP) Full-3550 Full-222 FuIM 18 3/4-2660 0.563-125 0.563-67 1/2-1775 0.250-56 0.250-30 1/4-885 0.063-14 0.063-8 1/8-445 0.016-4 0.016-2 0-0 0.050-11 Torque (Ib-ft) Speed-RPM Figure 14. Courtesy United Centrifugal Pumps.

PULSATION CONTROL FOR RECIPROCATING PUMPS Suppress the surge in the suction and discharge systems of your positive displacement pumps By V. Larry Beynart, Fluid Energy Controls, Inc. This article reviews the specific requirements for the The source of fluid pressure pulsations selection and sizing of pulsation dampeners to reduce the fluid pressure pulsations generated by positive displacement It is almost mandatory that multiple pump installations reciprocating pumps. It includes analyses of the nature of have well-designed individual suction and discharge pulsation pressure pulsation and explains the advantages of using control equipment. While multiple pumps may be arranged pulsation dampeners and suction stabilizers. Sizing proce- to run at slightly different speeds, it is impossible to prevent dures for dampeners illustrate how their application them from frequently reaching an quot;in-phasequot; condition improves suction and discharge control in reciprocating where all the pump flow or acceleration disturbances occur pumps, reduces pressure pulsations, increases Net Positive simultaneously. Having more than one pump can multiply the Suction Head (NPSH), and makes the service life of pumps extent of the pipe vibration caused by such disturbances longer and more efficient. because the energy involved is similarly increased. The reduction of the hydraulic pressure pulses achieved by using pulsation control equipment is usually reported in total pulsation pressure quot;swingquot; as a percentage of average Combating pulsation pressure. This method is used widely within the industry and is often recommended as a standard. Referring to Figure 3, The primary purpose of pulsation control for reciprocating pumps is to attenuate or filter out pump-generated pressures 10 quot;Crank Movement At that create destructive forces, vibration, and noise in the piping system. Every reciprocating pump design has inherent, built-in pressure surges. These surges are directly related to the crank-piston arrangement (Figure 1). Fluid passing through a reciprocating pump is subjected to the continuous change 10 Crank Movement Al in piston velocity as the piston accelerates, decelerates, and End Ol Stroke Moves Piston 1% of Stroke (») stops with each crankshaft revolution. During the suction stroke, the piston moves away from the pump's head, Figure 1. The rotary motion of the crank movement is thereby reducing the pressure in the cylinder. Atmospheric transformed into the reciprocating motion of the piston. pressure, which exists on the surface of a liquid in a tank, forces liquid into the suction pipe and pump chamber. During the discharge stroke, the piston moves toward the Frequency, Hz pump's head, which increases the pressure in the cylinder Duplex Triplex Quint RPM Pump Pulsequot; Pulsequot; Pulsequot; and forces the liquid into the discharge pipe. This cycle is 50 0.8 3.2 2.4 4.0 repeated at a frequency related to the rotational speed of the 100 1.7 6.8 5.1 8.5 150 2.5 10.0 7.5 12.5 crank. 200 3.3 13.2 9.9 16.5 quot; Pubc = pump RPM x number of cylinders/60 At every stroke of the pump, the inertia of the column of liquid in the discharge and suction lines must be overcome to accelerate the liquid column to maximum velocity. At the end of each stroke, inertia must be overcome and the column decelerated and brought to rest. Figure 2 compares pump pulsation frequencies to pipe span natural frequencies of vibration. A key benefit of pulsation control is the reduction of fatigue in the pump's liquid end and expendable parts. Reducing the pressure peaks experienced by the piston will Figure 2. (a) Typical pump and pulse frequencies, (b) Typical in turn reduce the power-end peak loading. piping natural frequencies.

the percent of residual pulsation pressure can be calculated system, amplification and excessive vibration occur. Ampli- as follows: fication factors can be as high as 40 for pulsation resonance and 20 for mechanical resonance. If the mechanical Case I: AP/Ave x 100 = 460/1000 x 100 = 46% resonance coincides with the acoustic resonance, a combined amplification factor of 800 is possible. Case IL AP/Ave x 100 = 70/100 x 100 = 7% It is very important that any reference to degree of Wear and fatigue. Pulsation can lead to wearing of pulsation should apply to the total excursion in terms pump valves and bearings and, if coupled with mechanical of pressure and percentage. In Case I, the total pulsation of vibration, will often lead to loosening of fittings and bolts and 460 psi (46%) implies that the excursion is from 680 psi (32% thus to leakage. In severe cases, mechanical fatigue can lead below the average) to 1,140psi (14% above the average). to total failure of components and welded joints, particularly when the liquid being pumped is corrosive. Reciprocating pumps introduce into the suction and discharge systems three apparently unrelated pressure disturbances, which are illustrated in Figure 4. These Cavitation. Under certain pumping conditions, areas of include: low system pressure can occur. If this pressure falls below the liquid's vapor pressure, local boiling of the liquid will • A low-frequency disturbance at A, based on the rate at occur and bubbles of vapor will form. If the liquid has maximum flow velocity pressures dissolved gases, they will come out of solution before the • A higher frequency due to maximum acceleration liquid boils. pressure at the beginning of each piston stroke at B • A pressure disturbance at the point of flow velocity change (valley) at C Pressure waves summary The earlier analysis of the source and nature of fluid Major problems caused by pressure pulsations pressure pulsations enables us to build the idealized pressure wave pattern (Figure 5). Although identical forces are acting on the liquid in the suction and discharge sides of a pump, Unsteady flow. The most obvious problem caused by the effects of these forces vary greatly. The major difference pulsation is that flow is not constant, which can lead to is that acceleration effects in the suction at B tend to problems in processes where a steady flow rate is required, separate the liquid into vapor or quot;cavitationquot; pockets that such as applications involving spraying, mixing, or metering. collapse with a high-magnitude pressure pulse. Acceleration Flow rates can be difficult to measure with some conven- on the discharge side of the piston at B, however, tends to tional types of flowmeters. compress the liquid and create an impulse pressure pulse. This requires separate consideration of the suction and Noise and vibration. Many reciprocating pump installa- discharge systems when choosing and installing pulsation tions suffer from problems that can lead to excessive control and suction stabilizing equipment. maintenance costs and unreliable operation. Typical examples are noise and vibration in the piping and the Pulsations in suction systems pump. Vibration can lead to loss of performance and failure of valves, crossheads, crankshafts, piping, and even pump The friction losses in a suction system are usually low barrels. High levels of pulsation can occur when the because of the relatively short length and large diameter of pulsation energy from the pump interacts with the natural the piping involved. Accordingly, the flow-induced A-type acoustic frequencies of the piping. A reciprocating pump pressures are of low magnitude compared to the accepted produces pulsation at multiples of the pump speed, and the magnified pulsation in the system is generally worse at multiples of the plunger frequency. Most systems have more than one natural frequency, so problems can occur at MAX 1140 PSI MAX 1021 PSI differing pump speeds. AVE 1000 AVE 1000 MIN 680 MIN 951 Shaking. Shaking forces result from the pulsation and ZERO ZERO cause mechanical vibration of the piping system. These forces are a function of the amplitude of the pulsation and the cross- Figure 3. Reporting the degree of pulsation and sectional area of the piping. When the exciting frequency of control, (a) Non-dampened waveform from flow variation the pulsation coincides with a natural frequency of the (Case I). (b) Dampened waveform from flow variation (Case II).

pumps. The usual effects are wear of surfaces in contact due to erosion, slow increase in clearances, reduction in Discharge performance, noise, and vibration. Cavitation damage in reciprocating pumps can happen very quickly and be Suction catastrophic. A cavitating 60 hp pump can break a major component every 350 hours. Suction control. Suction control is an essential step in Degrees the design of any pump system and is equally important for reciprocating and centrifugal pumps. This process requires Figure 4. Triplex single-acting pump—points of induced pressure disturbances. the following major steps: 1. Compare Net Positive Suction Head available (NPSHa) to Net Positive Suction Head required (NPSHr). (NPSH percentage of change (Figure 4). For example, if the suction is a feature of the system layout and liquid properties. It is pressure is a static 20psi, the 23% variation of a triplex often referred to simply as NPSHa or NPSH available. single-acting pump would generate a theoretical A-type Pump manufacturers usually quote NPSHr. This is a pressure variation of only 9.2 psi. This is hardly enough function of the pump's performance and is related to the energy to set pipes in motion. The same flow variation at pump's susceptibility to cavitation. To calculate the 1,400 psi working pressure, however, creates more than NPSHa to the pump, it is necessary to take into account enough energy—a 644 psi A-type pressure variation in the the friction losses of the suction piping. NPSHa is then discharge line. the difference between the total pressure on the inlet side Even so, the forces of acceleration become overwhelming of the pump and the sum of the vapor pressure of the disturbances in the suction. Pressure pulses of more than liquid and the friction losses of the suction pipes.) 25 psi are often encountered in pumps—even in systems with 2. NPSHa is greater than NPSHr, the general suction condi- short suction pipes. A small amount of dampening of the tions are under control, and their improvement (with flow-induced pressure can reduce pulsation to a negligible respect to NPSHa) by installation of a suction pulsation amount, leaving the 25-psi C-type acceleration pulsations. dampener is not required. In this situation, the installa- These remaining pulsations can create damage by enabling tion of a pulsation dampener would be required only for cavitation (Figure 4). pressure fluctuation and noise reduction purposes. 3. If NPSHa is less than NPSHr and the piping rearrange- Cavitation. In practice, cavitation occurs at a pressure ment of a suction system is impossible or insufficient, slightly higher than the vapor pressure of the liquid. Because installation of a pulsation dampener before the pump's vapor pressure is a function of temperature, a system may inlet is essential. run without cavitation during the winter months but be 4. Choose the appropriate dampener design based on steps subject to cavitation problems during the summer when 1-3. Using this process will help you achieve maximum temperatures are higher. Cavitation is connected with the reduction of pressure fluctuation. presence of microscopic gas nuclei, which are in the solid materials bonded in the liquid. It is because of these nuclei that liquids cannot withstand tension. Without them, it is Pulsations in discharge systems estimated that water could transmit a tension of approxi- mately 150,000 psi. The same forces at work on the suction side also affect During the beginning stages of cavitation, the nuclei give discharge systems, but the pressure at A due to flow-induced rise to the formation of gas bubbles, which are swept along to pulsations becomes overwhelming at 460 psi (Figure 4). The an area where the pressure is higher. The bubbles then 25 psi contribution from acceleration at C is a small collapse, producing high-intensity pressure waves. The percentage (2.5%) of the total discharge pressure. One formation of successive bubbles quickly follows. This exception is when the pump is delivering into a low-friction, repeating cycle lasts for only a few milliseconds, but during high-pressure system such as a short vertical discharge system this time the localized pressures can be as high as 60,000 psi (Figure 6(a)). An example is a mine dewatering application, and temperatures can rise as much as 15000F. where the pump is delivering to an already pressurized Cavitation can occur at a pump's inlets, inside the barrels, system—a pressurized pipeline—through a short connecting at the draft tubes of turbines and on propellers, mixing pipe. Another exception is when the pump is delivering to blades, Venturi meters, and other places where changes in already pressurized systems such as hydraulic press accumu- fluid velocity occur. The cavitation effects in rotary positive lators (Figure 6(c)). In those two cases, the acceleration displacement pumps are similar to those in rotodynamic pressures can become the overwhelming disturbance,

particularly if the piping system is relatively long compared to should be installed as close as possible to the pump inlet, and a suction system (but considerably shorter than a quot;pipelinequot;). the nitrogen in the bladder should be precharged to 60% of the pump's inlet pressure. Figure 8 shows a large capacity Control of pulsation with dampeners version of this design. The same type of suction pulsation dampener (150psi) Successful control of damaging pulsation requires careful was installed on a 120 gpm and 30 psi Wheatley triplex pump selection of pulsation dampeners and proper location of with 3quot; piston bore, 3-1/2quot; stroke. Another two discharge those dampeners in a piping system. The earlier pressure 1,500 psi pulsation dampeners were used with a 263 gpm and pulsation analysis showed that dampeners will reduce the 642 psi Wheatley quintuplex pump with 3-3/8quot; bore, 4-1/2quot; level of pulsation and consequently pressure fluctuations. stroke. With lower pipe vibration and noise, wear at the pump, risk of cavitation, and metal fatigue will be less as well. (There Flow-through dampeners. This design should be are several OSHA standards dedicated to the control of installed on a pump's suction system when inlet pressure industrial noise in the workplace. One of the most practical (NPSH) is very low and the fluid can contain entrained gases and inexpensive ways to meet OSHA noise level require- such as air, carbon dioxide, sulphur dioxide, etc. Flow- ments and achieve the overall performance improvement of through units have significant advantages over the appendage reciprocating pumps is installation of pulsation dampeners design for a few reasons. First, the inlet and outlet diameters on suction and discharge systems). of the dampener are equal to the pipe diameter. Second, the in-line installation enables reduction of the inlet fluid Dampener types. After evaluating the causes and types velocity into the pump. In addition, the bladder response of pressure pulsation effects in reciprocating pumps, we can time is faster than an appendage dampener, due to the review which available fluid pressure pulsation control minimum distance between the shell inlet and pipe center- equipment is best for specific conditions. line. The flow-through design can also be used to improve Hydro-pneumatic pulsation dampeners are gas-type the NPSHa of systems with low NPSH because it acts as a energy-absorbing design that uses a gas-filled bladder to suction stabilizer. absorb the quot;Aquot; flow on peaks and to give back that flow on Figure 9 shows an actual installation of a discharge flow the quot;lows,quot; thereby reducing flow variations beyond the through 2.5 gallon capacity dampener. This carbon steel, pump and consequently reducing the flow-

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