PUMP CHARACTERISTICS AND APPLICATIONS PDF

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Pump Characteristics And Applications Pdf

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o Pumps capable of feeding energy to water in com- other applications of pumps can be given, but they Each of the characteristics is explained in the next. Download Pump Characteristics and Applications 3rd Edition By Michael Volk Easily In PDF Format For Free. Centrifugal Pump Performance Characteristics for Domestic Application. Article ( PDF Available) in MATEC Web of Conferences

Each case study includes background on the pumprelated problem, an analysis of the problem, and the resulting solutions and lessons learned.

In addition to the above, the entire book was updated to reflect the latest thinking on pumps. A number of helpful new sections were added, such as the ten steps to determine total head that are summarized in Chapter 2 and the mechanical seal piping plans that are discussed and illustrated in Chapter 5. With this third edition we have switched to a downloadable demo of the software.

Use of a software tool to design or analyze a piping system should be considered if the system is complex e. Using a software tool to size pipes and determine pump total head allows the piping and pumps to be matched more accurately to the expected demands of the system, and helps to keep the pump from being oversized, thus saving energy and reducing pump maintenance costs.

This website is in compliance with the Digital Millennium Copyrights Act. Series Application. Booster and main pump are operated in series. The following points are to be kept in mind while considering the series operation. In other words, a single pump will operate at a lower head than if it were working in series.

This means the driver rating is usually higher as lower head operation means higher flow and higher power requirement. Parallel Application. Theoretically they can be used on large as well as small pumps of all specific speeds. Stepanoff gives a complete description of single-volute casing design. In all volute pumps the pressure distribution around the periphery of the impeller is uniform only at the BEP.

This pressure equilibrium is destroyed when the pump is operating on either side of the BEP, resulting in a radial load on the impeller. This load deflects the pump shaft and can result in excessive wear at the wearing rings, seals, packing, or bearings, In extreme cases, shaft breakage due to fatigue failure can result. The magnitude of this radial load is given by: For a specific single-volute pump it reaches its maximum at shutoff and will vary between 0. The effect of the force will be most pronounced on a single-stage pump with a wide b2 or a large-sized pump.

It is safe to say that with existing design techniques, single-volute designs are used mainly on low capacity, low specific speed pumps or pumps for special applications such as variable slurries or solids handling. Double-Volute Casing Designs A double-volute casing design is actually two single-volute designs combined in an opposed arrangement. The total throat area of the two volutes is identical to that which would be used on a comparable singlevolute design. Double-volute casings were introduced to eliminate the radial thrust problems that are inherent in single-volute designs.

Test measurements, however, indicate that while the radial forces in a double volute are greatly reduced, they are not completely eliminated. Radial thrust factor. For this reason, the pressure forces around the impeller periphery do not precisely cancel, and a radial force does exist even in double-volute pumps.

Values of the constant K have been established experimentally by actually measuring the pressure distributions in a variety of double-volute pumps. The data presented in Figure apply to conventional singlestage double-volute pumps and indicate substantial reductions in the magnitude of K. Tests on multistage pumps with completely symmetrical double-volute casings indicate that the radial thrust is nearly zero over the full operating range of the pump. The hydraulic performance of double-volute pumps is nearly as good as that of single-volute pumps.

Tests indicate that a double-volute pump will be approximately one to one and one-half points less efficient at BEP, but will be approximately two points more efficient on either side of BEP than a comparable single-volute pump. Thus the double-volute casing produces a higher efficiency over the full range of the head-capacity curve than a single volute. Double-volute pump casings should not be used in low-flow below GPM single-stage pumps.

The small liquid passages behind the long 54 Centrifugal Pumps: Design and Application dividing rib make this type of casing very difficult to manufacture and almost impossible to clean.

In large pumps double-volute casings should be generally used and single-volute designs should not be considered, The Double-Volute Dividing Rib Splitter The dividing rib or splitter in double-volute pumps causes considerable problems in the production of case castings.

This is particularly true on small-capacity pumps where flow areas are small and a large unsupported core is required on the outside of the dividing rib splitter. The original double-volute designs maintained a constant area in the flow passage behind the splitter. This concept proved to be impractical due to casting difficulties.

In addition the consistently small flow areas caused high friction losses in one of the volutes, which in turn produced an uneven pressure distribution around the impeller. Most modern designs have an expanding area in this flow passage ending in equal areas on both sides of the volute rib. The effects of volute rib length on radial thrust are shown in Figure Note that the minimum radial thrust was achieved during Test 2 for which the dividing rib did not extend all the way to the discharge flange.

Also note that even a short dividing rib Test 4 produced substantially less radial thrust than would have been obtained with a single volute. Triple-Volute Casings Some pumps use three volutes symmetrically spaced around the impeller periphery. Total area of the three volutes is equal to that of a comparable single volute.

The triple volute casing is difficult to cast, almost impossible to clean, and expensive to produce. We do not see any justification for using this design in commercial pumps. Quad-Volute Casings Approximately 15 years ago a 4-vane quad volute was introduced.

The discharge liquid passage of these pumps is similar to that of a multi-stage crossover leading to a side discharge. There is no hydraulic advantage to this design. The only advantage of this design is its reduced material cost. The overall dimensions of quad-volute casing are considerably smaller than those of a comparable double-volute pump.

Volute Design 55 56 Centrifugal Pumps: Design and Application Circular-Volute Casings Several pump manufacturers have conducted tests to evaluate the hydraulic performance of pumps with circular volutes.

A study of the results of these tests reveals that circular volutes improve the hydraulic performance of small high head or low specific speed units and impair the performance of high specific speed pumps. Specifically, pump efficiency is improved below specific speeds of This can be explained by remembering that in a conventional volute a uniform pressure and velocity distribution exists around the impeller periphery only at the BEP. At off-peak capacities, velocities and pressures are not uniform.

For circular volutes the opposite is true. Uniform velocity and pressure exist only at zero flow. This uniformity is progressively destroyed as the capacity is increased, Therefore, at BEP the casing losses are generally greater than those of the conventional volute. For low specific speed pumps, however, there is some gain in efficiency due to a circular volute since the benefits of the Improved surface finish in the machined volute outweigh the problems created by the nonuniform pressure distribution.

A comparison of the efficiency of circular and conventional volutes is shown in Figure The use of a circular volute design should be considered in the following instances: General Design Considerations It was pointed out previously that the casing itself represents only losses and does not add anything to the total energy developed by the pump. In designing pump casings it is therefore important to utilize all available means of minimizing casing losses.

However, commercial considerations dictate some deviations from this approach, and experience Volute Design 57 Figure Efficiency comparison of circular and conventional volutes, has shown that these do not have a significant effect on casing losses.

The following design rules have shown themselves to be applicable to all casing designs: Constant angles on the volute sidewalls should be used rather than different angles at each volute section.

Experience has shown that these two approaches give as good results and the use of constant wall angles reduces pattern costs and saves manufacturing time.

The volute space on both sides of the impeller shrouds should be symmetrical. Design and Application 3. All volute areas should be designed to provide a smooth change of areas. Circular volutes should be considered for pumps below a specific speed of Circular volutes should not be considered for multistage pumps.

The total divergence angle of the diffusion chamber should be be tween 7 and 13 degrees.

The final kinetic energy conversion is obtained in the discharge nozzle in a single-stage pump and in both the discharge nozzle and crossover in a multi-stage pump. In designing a volute, be liberal with the space surrounding the impeller. In multi-stage pumps in particular, enough space should be provided between the volute walls and the impeller shroud to allow one-half inch each way for end float and casting variations, A volute that is tight in this area will create axial thrust and manufacturing problems.

The Use of Universal Volute Sections for Standard Volute Designs It has been noted that when the volute sections of different pumps are factored to the same throat area; their contours are almost identical.

Any differences that do exist can be traced to mechanical considerations or the designer's whim, rather than any important principle of hydraulic design.

Similarly, factoring the impeller width and the radial gap between the impeller and the cutwater reveals that the values of these parameters also lie in a very narrow random range. In other words, the entire discharge portion of the pump casing when viewed in cross section and factored to a common throat area has only minor variations throughout the entire specific speed spectrum. This fact enables us to eliminate the usual trial-and-error method of designing volute sections while still consistently producing casings to a high standard of hydraulic design.

To facilitate this process we have prepared a set of "universal" volute drawings on which the typical volute sections described above have been laid out for a 10 sq in.

Once the designer has chosen his throat area, he can quickly produce the required volute sections by factoring the sections shown for the "universal" volute. Sections for a single-volute pump are shown in Figures and , and sections for a double volute pump are shown in Figure Typical single-volute layout.

The universal volute sections for such a design are shown in Figure A rectangular volute casing requires the same throat area as a standard volute casing and should be laid out according to the principle of constant velocity. Rectangular volutes are widely used in small single-stage and multistage pumps.

The benefits of the rectangular volute are strictly economical. The simple volute section yields a considerable cost savings due to reduced pattern costs and production time. Over the range of specific speeds where it is used the hydraulic losses are negligible. The Design of Circular Volutes The details of a typical circular volute casing design are shown in Figure , The ratio between the impeller diameter, D2, and the volute di- 60 Centrifugal Pumps: Universal volute sections for single-volute pump.

Universal volute sections and typical layout for trapezoidal doublevolute pump. Volute Design 61 Figure Universal volute sections and typical layout for rectangular double volute pump. The volute width, t, should be chosen to accommodate the widest maximum flow impeller that will be used in the casing. The capacity at BEP can be controlled by the choice of the volute diameter, D4. Generally, the best results are obtained by selecting the volute width and diameter for each flow requirement.

To minimize liquid recirculation in the volute, a cutwater tongue should be added.

General Considerations in Casing Design There are several considerations in the casing design process that apply to all volute types. These are as follows: This area together with the impeller geometry at the periphery establishes the pump capacity at the best efficiency point.

The throat area should be sized to accommodate the capacity at which the utmost efficiency is required, using Figure Where several impellers in the same casing 82 Centrifugal Pumps: The use of these figures will save the designer time and introduce consistency into the design process.

Typical layout for circular volute pump. Volute Design S3 Manufacturing Considerations Casings, particularly of the double-volute design, are very difficult to cast. In small- and medium-sized pumps the volute areas are small and the liquid passages are long, requiring long unsupported cores.

In volutetype multi-stage pumps the problem is more pronounced since there are several complicated cores in a single casing. Casing Surface Finish To minimize friction losses in the casing, the liquid passages should be as clean as possible.

Since cleaning pump casings is both difficult and time consuming, an extreme effort to produce smooth liquid passages should be made at the foundry. The use of special sand for cores, ceramic cores, or any other means of producing a smooth casting should be standard foundry practice for producing casings.

Particularly with multi-stage pumps, however, even the best foundry efforts should be supplemented by some hand polishing at points of high liquid velocities such as the volute surfaces surrounding the impeller and the area around the volute throat. Both of these areas are generally accessible for hand polishing. In addition, both cutwater tongues should be sharpened and made equidistant from the horizontal centerline of each stage.

The same distance should be maintained for each stage in a multistage pump. Casing Shrinkage Dimensional irregularities in pump casings due to shrinkage variations or core shifts are quite common.

Shrinkage variations can even occur in castings made of the same material and using the same pattern. The acceptance or rejection of these defects should be based upon engineering judgments. However, knowing that shrinkage and core shifts are quite common, the designer should allow sufficient space for rotating element end float. The allowance for total end float should be a minimum of onehalf inch.

Conclusion Although it is often claimed that casings are very efficient, this is misleading, since the hydraulic and friction losses that occur in the casing can only reduce the total pump output and never add to it. It is the designer's responsibility to do his utmost to minimize these losses. Total impeller width including shrouds at D2 in. Experimental constant Specific gravity Reference Stepanoff, A.

This type of pump offers the following features: Many pumps in service are operating at 3, to 4, psi discharge pressure.

Only when shaft diameter is too big or rotating speed too high, will Kingsbury-type bearings be required.

Pump Characteristics and Applications 3rd Edition By Michael Volk

The impeller design for the multi-stage pump is the same as that for a one-stage unit, as described in Chapter 3. The double-volute design is also the same as that for a one-stage pump as described in Chapter 5, 65 Figure The term "crossover" refers to the channel leading from the volute throat of one stage to the suction of the next. Crossovers leading from one stage to the next are normally referred to as "short" crossovers and are similar to return channels in diffuser pumps.

These are normally designed in right hand or left hand configurations, depending upon the stage arrangement, Crossovers that lead from one end of the pump to the other or from the center of the pump to the end are normally referred to as "long" crossovers. The stage arrangements used by various pump manufacturers are shown schematically in Figure Arrangement 1 minimizes the number of separate patterns required and results in a minimum capital investment and low manufacturing costs.

However, with this arrangement a balancing drum is required to reduce axial thrust. Arrangement 2 is used on barrel pumps with horizontally split inner volute casings. Arrangement 3 is the most popular arrangement for horizontally split multi-stage pumps and is used by many manufacturers. Finally, with Arrangement 4 the series stages have double volutes while the two center stages have staggered volutes.

This design achieves a balanced radial load and an efficient final discharge while requiring only one "long" crossover, thereby reducing pattern costs and casing weight. General Considerations in Crossover Design The principal functions of a crossover are as follows: Velocity cannot be efficiently converted into pressure if diffusion and turning are attempted simultaneously, since turning will produce higher velocities at the outer walls adversely affecting the diffusion process.

Furthermore, a crossover channel that runs diagonally from the volute 68 Centrifugal Pumps: Multi-stage pump stage arrangements. For these reasons, the multi-stage pumps of 25 or more years ago were designed with high looping crossovers.

To achieve radial balance these crossovers were in both the top and bottom casing halves. This design, referred to as the "pretzel" casing, was very costly, difficult to cast, and limited to a maximum of eight stages. These problems prompted a study to evaluate the performance of various crossover shapes. A 4-in. For these tests the pump hydraulic passages were highly polished micro-inches , ring clearances were minimized and component crossover parts were carefully matched using a template.

Configuration 1 was designed with a total divergence angle of Design of Multi-Stage Casing 69 Figure Configurations evaluated during crossover performance study. From this point the area was held constant to the impeller eye. To prevent prerotation a splitter was added to the suction channel. Crossovers 2 and 3 were designed maintaining the same areas at sections A, B, and C with the same divergence angle but progressively reducing the radial extent of the crossover.

The "U" bend on Crossovers 1 and 2 were cast separately from the casing and highly polished before welding. Crossover 3 was cast as a single piece, and the "U" bend polished only in the accessible areas.

The results of testing all three crossover configurations are shown on Figure The tests indicated that Crossover 1 yielded a peak efficiency four points higher than Crossover 3.

Subsequent testing of commercial units, however, indicated the difference to be only two points. The difference in improvement was attributed to the poor quality of the commercial castings and the use of normal ring clearances. The two-point efficiency loss associated with Crossover 3 was deemed commercially acceptable and was incorporated in multi-stage pumps of up to fourteen stages by all the West Coast manufacturers.

These pumps were suitable for higher pressures, easily adaptable to any number of stages, odd or even, and readily castable even in double-volute configurations. Design and Application Specific Crossover Designs A successful multi-stage pump development should produce a product that has excellent hydraulic performance, low manufacturing cost, and requires a minimum initial capital investment.

These three items become the basic design requirements during the layout of horizontally split multi-stage pumps. Hydraulically, the pump design should achieve the best possible efficiency, as well as the highest head per stage, thereby minimizing the number of stages required. The best available technology should therefore be utilized to produce the most efficient volutes and impellers. Although crossover design has only a secondary effect on pump efficiency, it too should use every available "trick" to achieve the best possible results.

Figure shows short and long configurations of the two basic types of crossovers normally used on multi-stage pumps. Both have been tested by the West Coast pump companies. Results of these tests indicate that the radial diffusion type is approximately one point more efficient than the diagonal diffusion type.

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Results of crossover performance study. Design of Multi-Stage Casing 71 Figure Radial and diagonal diffusion crossovers.

This point should be reached before the "U" bend to the suction channel. The suction channel should be sized to accommodate the largest capacity impeller that will be used in the pump. The area of the suction return channel should be consistent immediately after the "U" bend. Some designers prefer to decelerate slightly at the impeller eye; however, recent tests indicate that better efficiency is obtained if the liquid is accelerated as it approaches the impeller eye.

However, the cost of adding this splitter is generally prohibitive, and it is not generally used. From a theoretical viewpoint, crossover channels should have a circular cross section to minimize friction losses. However, all the designs on the market today have rectangular shapes for practical reasons.

Design and Application To attaie the best hydraulic performance, anti-rotation splitters must be added to the suction channel at the impeller eye. The best overall results are obtained by placing two splitters at the casing parting split as shown In Figure The long crossover is identical to the short or series configuration yp to the area "B," where the long channel that traverses the pump begins, This long channel should be designed with a "window" at the top for cleaning and a properly shaped plate matched to the crossover opening before welding.

The configuration of the long crossover is also shown in Figure Crossovers with Diagonal Diffusion Sections The diagonal diffusion type crossover shown in Figure leads the liquid from the volute throat to the suction of the next impeller while traveling diagonally around the periphery of the volute.

This design has one long radius turn as compared to the "U" bend used in the radial diffusion type. Other than these differences, both types of crossover have the same diffusion, area progression etc.

Volk M. Pump Characteristics and Applications

Figure also shows the configuration in which the long crossover channel climbs over the short crossover. Even though this crossover has only a single long-radius turn, it is not as efficient as the radial type. This can be attributed to the diagonally located channel, which imparts a spiral motion to the fluid leaving the volute throat, resulting in hydraulic losses larger than those in the "U" bend.

Mechanical Suggestions In previous chapters, we have described design procedures for centrifugal pump impellers and volutes applicable to one-stage or multi-stage units and in this chapter crossovers for multi-stage pumps only. However, hydraulic considerations alone for multi-stage pumps are not sufficient to complete a final unit. Mechanical details must be considered.

This refers to patterns, foundry methods, mechanical bolting, and quality controls, Patterns Multi-stage casings have quite complex shapes of liquid passages, including crossovers, crossunders, double volutes at each stage, etc. For this reason, pattern equipment must be of high quality sectional design to allow for variations in number of stages. Normal practice is to make the first pattern a four stage of hard wood with each stage being a separate Design of Multi-Stage Casing 73 section.

For additional stages, the sections are duplicated in plastic material. For each stage combination, plastic sections should be assembled on their own mounting boards, This arrangement will allow several pumps of different stages to be produced at the same time. Foundries The core assembly for multi-stage pumps is very complex as shown in Figure It shows a stage 4-in. In order for the rotating element to fit into the pump casing, each volute core must be assembled perpendicular to the shaft centerline.

To assure perpendicularity, a special gauge should be made for this purpose.

It is also vitally important to cast casing with wet area surfaces as smooth as possible. For this reason, the casing cores should be made from "green sand" or ceramic materials. The major hydraulic loss in multi-stage pumps is friction loss. To minimize this, the as-cast-surface roughness of the internal passages should be a minimum of micro inches. The smoother the wet areas, the less the cost of hand polishing or grinding will be. Core assembly. Design and Application Casing Mismatch On horizontally split pump casings mismatching between the upper and lower casing halves is quite common.

This mismatch is normally corrected by hand filing using a template Figure The same template is used by the machine shop to define bolt locations and also by the foundry. It is quite important that the fluid passages in both halves of the casing match in both the horizontal and vertical planes since any mismatching will adversely affect pump performance.

Check for Volute interference After the pump casing is completely finished to satisfy quality specification, the assembled rotating element should be installed in the casing as shown in Figure to check for interference against volute walls. Also check the element for total end float, which should be no less than one-half inch total or one-quarter inch on each end. At each stud at assembly the metal will rise.

The relief will allow the gasket to lay flat, reduce gasket area, and increase gasket unit pressure. These can be installed recessed at the casing split or cast on casing hub rings.

Pressure up to 6, psi hydro can be obtained by special design of bolting and case. It is essential to prevent stage-to-stage bypassing to have all bolts located as close as possible to the open area. The high pressure differential in opposed impeller design multi-stage pumps is between two center stages and at the high pressure end. Particular attention to bolting should be paid in these two areas. It is sometimes advantageous to provide elevated bosses to bring bolting closer to open area.

Design of Multi-Stage Casing Figure Split case template. Multi-stage rotating element. As a guide to selecting bolt size use: S Allowable bolt stress psi AR Root area of bolt thread sq in.

Applications range from light duty building trade pumps to heavy duty pipeline injection pumps Figure The double-suction is a very simple machine whose initial cost is relatively low. All modern double-suction pumps are designed with double-volute casing to maintain hydraulic radial balance over the full range of the head-capacity curve.

Having a double-suction impeller, the pumps are theoretically in axial balance. Double-suction pumps often have to operate under suction lift, run wide open in a system without a discharge valve, or satisfy a variable capacity requirement.

These pumps may be quite large, pump high capacities, and handle pumpage with gas or entrained air.

In spite of all this, they are expected to operate without noise or cavitation. Suction passage design should therefore be based on the best available technical "know-how," and liberties should not be taken during the design process. Double-Suction Pump Design Pump Casing The double-suction, double-volute casing is designed in identical manner to the single-suction pump, as described in Chapter 5.

The specific speed, Ns, of a double-suction pump is identical to the single-suction unit. Do not divide calculated pump specific speed by the square root of 77 Figyre Experience shows that this procedure will give misleading design factors and unfavorable test results.

Design of the suction approach to the pump impeller for double-suction pumps will differ from the single-suction design. This will be covered in detail in the following paragraphs.

Double-Suction Impeller The method for calculating impeller diameter, impeller width, number of vanes, and vane angularity is identical to the procedure for the singlesuction impeller described in Chapter 3. The method for impeller layout will also follow Chapter 3, with a double-suction impeller being considered two single-suction impellers back to back.

Side Suction and Suction Nozzle Layout The importance of hydraulic excellence in the design of liquid passage areas from suction nozzle to the impeller eye or eyes is quite often minimized or unfavorably adjusted for economic reasons. Experience shows that this approach leads to many field NPSH problems. The current trend in industry is one of reducing NPSHA; therefore, it is essential for optimum NPSHR that the design of the suction approach to the impeller eye be carefully controlled.

We know from experience that in the design of the side suction inlet a certain amount of prerotation of the incoming liquid is desirable. To obtain this condition, the baffle or splitter is provided. The splitter will locate the radial section of zero flow, and the areas will progressively increase in both directions away from it.

The following drawings and information must be available to design a side suction. Volute layout. Shaft or sleeve diameter at the impeller. Suction nozzle size. Layout of the laterally displaced side suction should be done in two parts: Sketch an approximate end view and profile Figures and using the following guidelines: Double-suction layout—end view. Dimension D, x 0,84 D, X 0. Double-suction layout—profile.

For linear dimensions of Sections 1 through 7, and A-B, use Table Area progression from nozzle to impeller should follow Figure The range suggested permits impellers of different suction specific speeds to be used with a common suction. Areas of Sections 3 through 6 measured normal to the flow are shown in Table Make final layout after checking all areas and dimensional location of suction nozzle.

Draw a circle DI at point O. Locate Sections 1 through 7 from Table For length of A-B use Table Locate nozzle dimensions L and S and mark off nozzle diameter. Dimension L should be only long enough for gradual area progression and clearance behind the flange bolting. This clearance becomes more critical on horizontally split pumps, where nozzle bolting may interfere with the parting flange, 5.

Connect all points freehand. Layout volute metal line. Suction Layout Profile See Figure for a diagram of the profile. Layout impeller shape, including shaft and sleeve. Layout volute shape and metal line at intersection of chord Pa3. Layout volute shape in dotted lines at location X to observe maximum blockage. Curve metal line around volute sections into the impeller eye, maintaining minimum metal thickness.

Mark off chord lengths PI,? Locate nozzle diameter. Connect outer wall points freehand with ample curvature into impeller eye.

Lay in Sections 1 through 7. In this example only Section 1 is shown. Sections can be divided into any number of increments e.

Transfer these chords from the end view to the profile, measure dimensions P' - A' P' - B' and transfer to section layout. Suction area progression. Double-suction layout—sections. Design and Application Repeat Step 10 for Sections C-D or any other sections deemed necessary. Close sections with appropriate radii.

These should be liberal for castability and should follow a smooth transition in the end view. Example Double Suction Pump: This NPSH advantage of the double suction model makes it possible to be adopted in addition to the standard horizontally split, single-stage pump to many different types of centrifugal pumps such as: Probably more has been written on NPSH than any other subject involving pumps.

With so much literature available, one might assume that NPSH and its relationship to cavitation is well understood. Nothing could be further from the truth. To this day, NPSH is still misunderstood, misused, and misapplied, resulting in either costly over-design of new systems or unreliable operation of existing pump installations. As NPSHR increases with pump capacity, normal practice is to establish NPSHA at the operating condition, then add a reasonable margin to accommodate any anticipated increase in pumping capacity.

All too often, future operating problems begin here. It is not unusual for the ultimate user to add some anticipated increase in capacity, then for insurance the contractor designing the system adds even more. When the pump designer finally gets the data sheet, he designs the impeller inlet and suction nozzle geometry for operating capacity, plus user margin, plus contractor margin.

If these margins are not carefully con- 85 86 Centrifugal Pumps: Design and Application trolled, the result can be an over-sized pump that operates well to the left of BEP Figure Such over-designed pumps are vulnerable to surging, recircuiation, cavitation, noise, and vibration.

Pump Characteristics and Applications 2nd Edition By Michael Volk PDF Book Download

This is particularly true with high-suction specific speed pumps above 11, where the inlet geometry has already been extended for minimum NPSH. As this is not always possible or practical, pumps will often operate at lower flows. It is therefore important that the minimum flow for continuous trouble free operation be carefully considered by the pump designer. Minimum flow is influenced by physical pump size, margin between NPSHA and NPSHR, impeller inlet geometry, suction nozzle geometry, mode of operation, and last but not least, the liquid being pumped.

Margins of safety result in oversized pump. In considering NPSHR, it is necessary to understand that a centrifugal pump is designed as a hydraulic machine to move liquids. Any amount of entrained air or gas present will cause a deterioration in pump performance. Various tests substantiate the claim that a volume of only one percent air or gas will cause a loss of head and efficiency.

As liquid travels from suction nozzle to impeller eye, it will experience pressure losses caused by friction, acceleration, and shock at blade entry. If the summation of these losses permits vaporization of the liquid, vapor bubbles will form in the impeller eye, travel through the impeller, and upon reaching a high pressure region, collapse. This collapse or implosion of the vapor bubbles is classic cavitation, which can lead to impairment of performance and impeller damage Figures and Thus, predicting Figure Cavitation damage—looking into impeller eye.

Cavitation damage—at leading edge of one vane. NPSHR is in fact predicting the losses in the critical area between suction nozzle and the leading edge of the first-stage impeller blades Figure Moderate Speed Pumps One method used successfully for many years by pump designers will predict MPSHR with reasonable accuracy when the pump liquid is water.

Development of this chart is a result of acquisition of many years of pump test data. When velocities exceed those shown, the chart should not be extrapolated. The most commonly used method in determining the cavitation characteristics of a centrifugal pump is to cause a breakdown in the normal head capacity curve.

Pressure loss between suction nozzle and leading edge of impeller vane. Either of these methods will produce a breakdown in the head characteristic as shown in Figure , indicating a condition under which the performance of the pump may be impaired. Accurately determining the inception of cavitation requires extreme control of the test and involves sophisticated instrumentation. Loss in head during cavitation test.

Design and Application evidence that cavitation is present under test conditions. They fiirther caution, that the pump should be operated above the break-away sigma if noise and vibration are to be avoided.

Thus the same impeller when tested before and after a diameter trim will show different values of sigma yet will behave identically under cavitation. In comparing cavitation performance and predicting NPSHR, we prefer to use the suction specific speed parameter. This can be applied in the same manner as sigma yet has more significance as it relates only to the inlet conditions and is essentially independent of discharge geometries and pump specific speed.

Figure shows suction velocity triangles for NSs from 7, to 16, It graphically shows that as suction specific speed increases, normal vane entrance angle becomes flatter, C MJ relative to shaft becomes smaller, and peripheral velocity at impeller eye becomes greater. The ratio of CMI to LJt for higher suction specific speed, is very small and chances for cavitation are greatly increased. This cavitation would appear for the following reason: To move liquid from one point to another, a change of pressure gradient must take place.

When it becomes lower than the head required to overcome impeller entrance losses, the liquid will backflow creating cavitation and pump damage. Figure represents a test of a four-inch pump with eight different suction specific speed impellers.Torsional Critical Speed Analysis. Locate nozzle diameter. Constant BEP for different vane numbers and discharge angles by change in b2. Mechanical details must be considered.

Volk's services include pump training seminars; pump equipment evaluation, troubleshooting, and field testing; expert witness for pump litigation; witnessing of pump shop tests; pump market research; and acquisition and divestiture consultation and brokerage.

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