Reliability Issues †Centrifugal Slurry Pumps Essay Example for Free

Reliability Issues – Centrifugal Slurry Pumps Essay Introduction Pumps were probably the first machine ever developed, and are now the second most common machine in use around the world, out-numbered only by the electric motor. The very earliest type of pump is now known as a water wheel, Persian wheel or â€Å"noria†, consisting of a wheel of buckets that rotates to pick up water from a stream and dump it into a trough. Another early pump was the â€Å"Archimedean screw†, similar to the modern screw conveyor except that the flights were often fixed to the tube so that the whole arrangement would turn together. Both of these devices are still used, most commonly in basic agricultural applications. Pumps are now produced in an enormous range of types and sizes, for a very wide scope of applications, and this makes it difficult for any individual reference document or organisation to cover â€Å"pumps and pumping† as a general topic. So the broad field of pumping is classified into sub-divisions and then dealt with at that level. In the mining industry, the upper end of the pump scale includes impellers with diameters over 2.5m, slurry lines 10km long, particle size up to 100mm, flow rates handling more than 7000tph, and motors over 10MW. Finer slurries of around 1mm particle size are pumped for hundreds of kilometres in some operations. There are many ways to classify pumps. This just one of them. This document only addresses centrifugal pumps, with a focus on single-stage radial-flow slurry pumps. Centrifugal pumps are capable of meeting duties of up to 1.4 m /s at 30MPa, and higher volumes at lower 3 pressures. The maximum flow rate at low discharge pressure is about 180 m /s. Industrial applications requiring high delivery pressures generally use reciprocating fixed-displacement pumps, but they are limited in the amount of flow they can put out per unit. In general purpose applications, where different types of pumps could all deliver the performance sought, centrifugal pumps are usually the preferred choice due to lower lifecycle costs. Basic Requirements for Reliability Assuming correct pump manufacture and installation, the basic requirements for reliable long-term operation of centrifugal pumps are: 1. Continuous operation at best-efficiency point (BEP) 2. Adequate net positive suction head (NPSH) 3. Low velocity fluid flow within the pump and throughout the system 4. Processing of fluids that are benign ie: a) Chemically and physically stable b) At near-ambient temperatures c) Free of particles likely to cause wear or blockage Pumps of a basic design satisfying all these requirements have run for 50 years and more without major component replacement. The first three requirements are satisfied by matching pump performance to expected duty. Where item 4 cannot be addressed through pre-treatment of the fluid, the pump configuration, geometry and materials must be optimised to give best results. Obviously, item 4.c) is a dominating issue for slurry pumps as it cannot be eliminated and must be managed. Centrifugal Pump Construction Centrifugal pumps have two main sub-assemblies – the rotating parts (impeller, shaft, bearings), and the fixed parts (casing, piping connections, stand, foundations. Pumps of all types may be single stage or multi-stage. Multiple stages are used where it is not practical to generate the necessary discharge pressure using a single impeller. The simplest way to imagine a multi-stage pump is as one pump with its discharge feeding straight into the suction of a second pump so that the overall discharge pressure is increased while the flow rate stays the same. However, this arrangement is properly described as â€Å"single stage pumps in series†. A true multi-stage pump consists of multiple impellers mounted on a single shaft, positioned in a single casing made up of multiple chambers. Multi-stage pumps of this type are not used with slurries, but sometimes slurry pumps are mounted in series. Casing There are two types of casing designs â€Å"volute† and â€Å"diffuser†. A volute casing has a snail’s shell shape, while a diffuser casing has internal vanes. Diffuser casings are rarely used on single-stage radial pumps, and are not commonly used for handling slurries due to the flow restriction and high wear rates that would result. Slurry pumps have volute casings which house the impeller and have a spiral-shaped outer volume that extends 360 degrees and increases in cross-sectional area as it approaches the discharge flange. At full circle the volute overlaps itself, creating the cut-point, also known as â€Å"cut-water point† or â€Å"tongue†. The ideal shape is to have a steady linear increase in cross-sectional area for 360 degrees around the circumference starting from the cut-water point, but this can be difficult to manufacture. Compared to a clear water pump, a slurry pump has a much larger radial gap between cut-water point and impelle r, to reduce risk of blockage. Where a pump is identified as oversize for its duty, and is suffering high recirculation wear, it may be possible to fit liners with an extended cut-water point that throttles the flow. In theory, when a pump operates at its best efficiency point (BEP), the pressure acting on the impeller and casing are uniform. However, in practice the pressure is rarely completely uniform, and if a pump is operating away from its BEP the imbalanced in the radial forces acting on the impeller become significant. These forces are larger for bigger pumps operating at higher pressures. Running a large pump below rated capacity can create unbalanced radial forces that may (over time) damage the bearings or snap the shaft. If it is known that a pump may need to occasionally operate well away from its BEP, the manufacturer should include an oversize shaft arrangement in the design, but with commercial competition driving purchase decisions this may have to be specifically requested. Another option for reducing imbalanced radial forces is to use a twin-volute design, which consists of a wall splitting the volute in half for about half its circumference, ending after the cut-point but before the discharge flange. This is not practical for most slurry applications. Casings must be designed to allow the impeller to be installed inside, and so are manufactured in at least two parts. Solid casings have a removable cover, either on the suction side or shaft side or both, but the volute shape is a one-piece casting. Casings may also be split, either axially or radially. Axially split housings make inspection easier because the upper piece can usually be removed without disturbing the shaft or piping too much. Split casings may tend to â€Å"breathe† at high pressures, resulting in leakage, air entrainment, vibration, misalignment etc. Casings are normally provided with ribbing at the location of highest stresses, to minimise this. Open or semi-open impellers require close clearances against the casing to ensure pumping efficiency. The casings generally include a side-plate that can be adjusted for minimal clearance using jacking screws or shims, especially in wearing applications eg slurries. Impeller Impellers are classified according to their design features ie: ï‚ · Suction flow orientation o Single suction ie inlet on one side only o Double suction ie inlet on both sides ï‚ · The direction of exit flow relative to the shaft axis ie: o Radial flow o Axial flow o Mixed flow ï‚ · Vane shape ie: o Single curvature vanes, also called straight vanes – the impeller surfaces that accelerate the fluid are straight and parallel to the axis of rotation o Francis or screw vane – the surfaces that accelerate the fluid are curved in relation to the axis of rotation ï‚ · Mechanical construction o Enclosed ie with side walls or â€Å"shrouds† o Open ie no shrouds o Semi-open ie shroud on one side only o Partially shrouded ie shroud not extending to impeller tips The open area through which the fluid flows into the impeller is called the suction eye. For a closed-shroud impeller, this is simply the hole in the shroud. The suction eye area is an important featur e of the pump design. The area taken up by the shaft, if it protrudes through the eye, is deducted when calculating eye area. Impellers can be single suction or double suction. A single suction impeller has an inlet eye on one side only, with the shaft extending out the opposite side so the impeller overhangs. A double suction impeller can be thought of as two mirror-image single suction impellers mounted back-to-back. They accept fluid from both sides and usually have a shaft that extends straight through the impeller with bearings providing support on both sides. Double suction impellers are usually fed fluid from a single inlet flange, with the fluid flow being split into two streams by channelling inside the casing. Double suction units provide advantages in reduced fluid velocity at the impeller eye, and better balancing of axial hydraulic forces, while single suction units are simpler in design, manufacture and maintenance. Most if not all slurry pumps are single suction type. Some pumps may have an inducer, which is an axial flow impeller with a few blades installed between the suction inlet and the main impeller, intended to improve the suction head seen by the main impeller. Impeller shrouds often incorporate thin â€Å"pump-out vanes† cast into the outside of the shrouds. Their purpose is to help clear any solids from the back hub of the impeller (opposite the inlet eye), reduce pressure at the seal area, reduce axial thrust, and discourage recirculation. Some impellers have similar vanes on the eye side as well as the shaft side – in this case, those on the shaft side are usually called â€Å"expeller vanes†. In clear water pumps, a cylindrical ring is usually cast or machined into the outside surface of the shrouds, coinciding with a matching feature in the casing, to help seal off the discharge fluid from the suction fluid and prevent internal circulation. Clearances here are tight in order to ensure pumping efficiency – typically around 0.25mm on radius for most common sizes of industrial pumps. In larger pumps the casing (and sometimes also the impeller) is usually protected at this point by replaceable â€Å"wear rings†, which may be high-wear items, and need to be replaced before efficiencies fall too low. It is good practice to replace wear rings once the clearance reaches twice the original specification. Wear rings are provided in a wide range of designs and materials according to the pressures, speeds and fluids involved. The wear rings on impeller and casing are often made from differing materials that are not subject to galling, to reduce problems should contact occur. Wear ring features may include labyrinths, water injection, inspection ports, adjustment mechanisms etc. Pumps handling light slurries may make use of wear rings, sometimes with water injection to reduce wear from the slurry. Pumps handling heavier slurries usually just use pump-out vanes. Slurry pump impellers must be designed to resist wear and tear, and this requires some pumping efficiency features to be sacrificed. For example, vane edges will be blunter, vanes and shrouds will be generally thicker, and the number of vanes will be decreased in order to open up the channels between them. Passages through slurry pumps, including impeller vane spacing, are larger than for clear water pumps. Open impellers are sometimes used for very stringy materials, but tend to be weak and wear quickly, and so are not very common. Vane shape is obviously a major element of impeller design. Two critical factors are the blade entry angle (ß1) and blade exit angle (ß2), as measured between the centre-line of the vane and a tangent to the inner or outer diameter (respectively) drawn from their tips, in the oppo site direction to rotation. Most modern pumps have impellers with ß2 smaller than ninety degrees – ie backward-curved blades. Theoretically, a forward-curved blade would give higher head, but at less efficiency. Some pumps have ß2 at ninety degrees, and these are sometimes referred to as â€Å"expellers†. Many clear-water impeller designs rely on close running clearances between vane tips and casing to minimise recirculation from one â€Å"vane chamber† to the next, and maximise efficiency. Even small amounts of vane tip wear can have an effect on head and overall efficiency. The outer and inner vane tips should be sharp, not rounded or chamfered. Replacing a pump which is too large for its duty can be a major exercise. It usually requires changes to the foundations, drive arrangement and piping, spares holdings, and so on. A model of the ideal size may be just not available. As an alternative, in some cases it may viable to install a reduced-diameter impeller without changing other components. If done correctly, trimming the impeller will move the pump’s BEP to match the actual system operating point. The efficiency at the new BEP will be lower than the BEP with the original impeller, but higher than was being achieved in practice when operating well away from the original BEP. The performance variation can be estimated using the â€Å"affinity laws† which often apply to a specific impeller before and after machining: Flow rate: Pump head: Motor power: Q1 / Q2 = n1 D1 / n2 D2 H1 / H2 = (n1 D1 / n2 D2) P1 / P2 = (n1 D1 / n2 D2) 2 So if running at the same speed, trimming an impeller by a certain proportion will result in a corresponding drop in flow rate, a greater decrease in head produced, and an even greater decrease in the motor power consumed. However, these equations are based on several assumptions and some caution is called for. Impellers are complex three-dimensional objects and their effects on the liquid are due to other factors that are also affected by machining, beyond just the outside diameter – eg open area, discharge blade angle and so on. The following considerations should apply. ï‚ · Diameter reductions should not exceed 10%. Reductions beyond 20% are generally considered extreme. Some references state 30% as the maximum reduction advisable. ï‚ · Some overlap in the vanes should be retained. ï‚ · The angle between the vane centreline and the tangent to the outer diameter drawn at its tip should be restored to original by filing, with most filing occurring on the trailing si de of the vane.   The vanes will probably be thicker after cutting, and should be filed back to original shape, by filing on the traling side of the vane. ï‚ · Vane tips should be kept sharp, not rounded or chamfered. Outer tips should be sharpened by filing on the trailing side, and inner tips by filing mostly on the leading side.   Inefficiencies will take the form of increased disc friction, increased flow path length within the casing, and more recirculation across vane tips. Impellers apply forces to the fluid and are subject to the equal and opposite forces themselves. The typical single-suction impeller engages with fluid entering the pump and at first accelerates it axially into the pump, before diverting it into the radial direction. The impeller pushes the fluid into the pump, and at the same time pushes itself axially back toward the inlet point. Another way of looking at this effect is to consider that the impeller is mostly exposed to pressurised fluid all over the shroud surfaces, but not at the eye on the suction side. The thrust on the impeller must be resisted by the shaft arrangement, which must always include bearings capable of serious thrust loading. Double-suction pumps typically have less axial loading, but can still experience axial thrust, especially if flow is restricted more on one side due to internal differences in the pump, or restrictions in fluid supply on one side. Clean water pump designs may incorporate features to reduce this imbalance, such as having wear rings on both sides of the impeller, with the pressure within t he volume they enclose largely equalised by â€Å"balancing holes† passing right through the impeller. Another method is the use of a balancing disc. This is a disc mounted on the shaft in a separate chamber, with a geometry and clearances designed to counterbalance thrust effects. However, these are not practical for slurry pumps, which may use pump-out vanes instead, to lower the pressure toward the inner area of the non-suction shroud. Axial thrust loads usually consist of a steady state component plus dynamic fluctuations. Heavy axial loading is often associated with recirculation. Where failure occurs it is usually a result of overloading and over-heating of bearing components. Measures to correct excessive axial loading include:   Restoring BEP operating conditions (which may include selecting a more appropriate pump size or trimming the impeller)   Ensuring internal clearances / wear are not excessive ï‚ · Verifying correct bearing type and installation including clearances / pre-load To further complicate this issue of axial thrust, single-suction pumps handling fluids with a high suction head may experience thrust on the impeller in the opposite direction, away from the inlet. And then there are pumps with highly variable duties and suction conditions that may experience impeller thrust in different directions at different times. Shaft The shaft transmits mechanical power to the impeller from the driving motor or engine. It must also support the impeller and restrict its axial and radial movement. The loads on the shaft include self-weight of the rotating components, torque, and forces transmitted to / from the fluid. Design of a shaft requires consideration of maximum allowable deflection, the span or overhang, the location and direction of all loads, any temperature variations, and the critical speed. Loads are normally at their maximum on start-up. All objects have a natural frequency at which they will vibrate after being struck. Machines made of several components with complex shapes normally have several natural frequencies, some of which dominate. In the case of pumps, if the rotational speed of the impeller matches a dominant natural frequency, small imbalances may be amplified to a level where they interfere with operation and/or reliability. These are known as â€Å"critical speeds†. Steady operating speeds between 75% and 120% of the first critical speed should be avoided. Pumps with longer overhang on the shafts have lower critical speeds. Shafts are referred to as rigid or flexible, according to whether the running speed is lower or higher than the first critical speed. Pumps with a flexible shaft must pass through a critical speed on each start-up. This is not usually a problem because frictional forces with the fluid and the bearings act as dampers for a period sufficient for transition through the critical speed. Pumps with speeds below 1750rpm, which includes most slurry pumps, are usually of the rigid-shaft design. The shaft must be designed so that any deflection will not bring moving parts into contact, for example at wearing rings, or cause non-concentricity in critical areas such as the shaft seal. As a general rule, shaft deflection should not exceed 0.15mm even under the most extreme conditions. Deflection and critical speed are related stiffening a shaft to reduce deflection will also raise its critical speed. For pumps with overhung impellers, as is the case for most slurry pumps, this often results in the shaft diameter between bearings being quite large. The fluid passing through a pump creates a hydrodynamic bearing effect, known as the â€Å"Lomakin Effect†. That is, to some extent, the impeller rotating in the casing with fluid present is like a shaft rotating in a journal bearing with oil present. The result is that the shaft is better supported when running than when idle, so that the shaft deflection will be less, and the critical speed of the shaft assembly will be higher. However, the Lomakin Effect varies with pump head and internal clearances, both of which diminish with wear. Therefore the effective critical speed may be expected to decrease with time in service. To allow assembly, shafts step up in diameter from coupling to bearing to impeller, so tha t any torque problems are very likely to appear first at the coupling rather than the impeller, at least in single stage pumps. Shaft Seal and Sleeve The shaft connects the drive to the impeller, and so must pass through the pressurised casing. Achieving a reliable seal between shaft and casing is one of the most problematic areas in pumping. Centrifugal pumps have two types of seals – mechanical seals and packing seals. Many designs of mechanical seals have been attempted for slurry pumps, without comprehensive success, and the remainder of this discussion concentrates mainly on packing seals and stuffing boxes. Note, however, that packing is only suitable within pressure and temperature limitations. Depending on pump design and duty, the seal may need to prevent either air ingress into the casing, or fluid egress out of the casing or both of these at different times, if operation is variable. Many casings are designed with the seal area built into a compartment configured to improve sealing performance. For mechanical seals, this compartment is usually referred to as the â€Å"seal chamber†, while for packing seal s, it is known as the â€Å"stuffing box†. Slurry pump seals usually consist of several rings of packing fitted in a stuffing box around the shaft, often with provision for grease lubrication or water injection to reduce friction and provide additional sealing (particularly for when the pump is stopped). There are many stuffing box design variations and many types and configurations of packing. Stuffing boxes will accept a number of rings of packing, with a packing ring or throat bush preventing extrusion into the casing, and a gland (sometimes called a â€Å"follower†) used to adjust packing compression. A lantern ring may be substituted for one of the packing rings, to cater for injection of grease or sealing water, water being particularly required if air would otherwise be sucked into the fluid stream at this point. Sealing water (or an alternative clean liquid) is usually required for: Slurries   Liquids for which leakage is not acceptable   Liquids that are not suitable for sealing purposes   Suction lifts greater than 4.5m (air ingress may interfere with priming)   Discharge pressures above 70kPa The packing must be placed under some compression and this tends to result in wear on the shaft, which is often sleeved to avoid having to replace the entire shaft once wear is advanced. There are numerous designs of shaft sleeves. The shaft sleeve must be resistant to friction and heat, and several different materials and surface treatments are available – eg hardened high-chrome stainless steel, ceramic, plasma spray or tungsten carbide coating etc. To prevent chipping, coatings should not extend to the edges of the sleeve. The sleeve does not contribute to strength, so the shaft itself must be large enough to carry all the loads, and this means that including a sleeve in the design enlarges the seal diameter. For small pumps, this may decrease pumping efficiency and raise the purchase cost to the point that sleeves may be abandoned and a stainless steel shaft used instead. Glands may be solid, or split to allow replacement without disassembly of pump or bearing assembly. They are usually made of bronze, cast iron or steel. Special designs are used to improve safety if the fluid is hazardous. The leakage of fluid past the packing is controlled by tightening the gland, compressing the packing axially and expands it radially so that leakage paths along the shaft sleeve are constrained. However, some fluid flow between packing and sleeve is usually needed to avoid overheating the packing and damaging the sleeve surface. Once the sleeve surface is damaged, the sealing efficiency decreases and more tightening is required, further damaging the sleeve, and so on. The secret is to provide a configuration of packing and seal water injection that suits the application, and then avoid over-adjustment. To further reduce the pressure at the shaft seal area, where the rear pump-out vanes are not sufficient, some slurry pumps are fitted with a second smaller open-faced impeller, usually called an â€Å"expeller†. Many different designs have been tried. If sealing water is used, there will be a design intention regarding the ratio of water to pass in to the volute compared to out past the gland follower. This can be controlled using the number of packing rings on each side of the lantern ring, but the lantern ring must be installed at the injection point. For clean water pumps, this seal water is sometimes provided from the pump discharge. Clean water must be used to avoid contaminating the packing with grit – filtration or cycloning may be necessary if the water contains some grit. When managing sealing arrangements, thought must be given to what happens when the pump is stopped. The pressure in the stuffing box changes to static conditions, which may result in slurry leaking into the packing and contaminating it, causing rapid sleeve wear on re-starting. But if sealing water continues to be applied, the slurry may be diluted, and eventually a sump can be filled with sealing water if left idle for a long time. For prolonged stoppages, sumps may be best dropped, for various reasons. On restarting, sealing water supply should start before the pump starts. Stuffing boxes in extreme applications may be provided with galleries through which cooling water can pass to prevent excessive temperatures around the packing. In applications where leakage must be more precisely controlled, or where elevated temperatures in the seal area must be avoided (for example where the fluid is volatile), mechanical seals may be suitable, provided that the fluid is not damaging to the seal components. A comparison between mechanical seals and packing seals is given below. ï‚ · Packing seals: o Low initial cost o Tend to deteriorate gradually o Easily replaced when necessary o Can handle large axial shaft movements o Always some leakage required o Require regular adjustment o Not suitable for hazardous / volatile fluids o Often cause progressive shaft sleeve wear o Can result in significant shaft power losses o Limited to low pressures and speeds ï‚ · Mechanical seals: o Minimal or zero leakage o No adjustments required o Suitable for hazardous / volat ile fluids o No shaft wear o Do not consume significant shaft power o Can handle high pressures and speeds o Tend to fail suddenly o Replacement requires pump disassembly o High initial cost Packing seals work as a result of axial compression, so that the packing rings extrude outward and apply radial pressure to the adjacent components, these being the static surface of the stuffing box, and the rotating shaft sleeve. A dynamic seal is formed between the packing rings and the sleeve surface, with some fluid flow between the two being necessary for lubrication and cooling. For clean water pumps, this fluid may be supplied from the inner end of the stuffing box, or from the discharge pipe via small diameter piping. In the case of slurries, grit in the fluid would add to friction and wear, so the lubricating and cooling fluid is usually injected from a separate clean water supply. The injection pressure should be 10 to 25psi greater than that at the inside end of the stuffing box, and this figure should be available from the pump designer. A rule of thumb is to set the gland feed water pressure to between 35 and 70kPa above pump discharge pressure. Pressure regulation is often helpful. In theory, some slurry pumps should operate with a pressure at the inside of the stuffing box which is below atmospheric pressure, so that the packing is required only to prevent air ingress into the pump. However, when the pump is turned off, or in abnormal operating conditions, slurry can pass back into the seal and contaminate the packing with grit, so these situations still call for water injection. Grease or oil may be used instead of water in some applications. Packing material must be able to withstand the operating environment and remain resilient to perform satisfactorily despite minor shaft misalignment, run-out, wear and thermal expansion / contraction. Packing is available in a huge range of materials (lubricant, binder and fibre / matrix) and in many sizes, shapes, and constructions, to suit different applications – particularly size, shaft speed, temperature, pressure, and chemical resistance. The number of packing rings varies between applications, the most common arrangement being throat bush or ring, three inner packing rings, lantern ring, two more packing rings, and gland follower. The lantern ring may be placed further in, to reduce slurry ingress. Packing size is usually proportional to shaft / sleeve outer diameter, as follows: Shaft / Sleeve OD (mm) 15 to 30 30 to 50 50 to 75 75 to 120 120 to 305 Packing Size (mm) 6 8 10 12.5 16 Shaft sleeve finish needs to be at least 0.4micron CLA to avoid excessive rotational friction, and the finish in the stuffing box bore needs to be at least 1.65 micron CLA to allow even compression during adjustment. The sleeve must be harder than the packing, and chemically resistant to the fluid pumped and the injection fluid. Any coating on the sleeve must have a good thermal shock resistance. The lantern ring allows for entry and distribution of the lubricant or flushing fluid. Lantern rings are usually split to allow installation and removal without pump disassembly. They were traditionally made from metal such as stainless steel, but lubricant-impregnated plastics are now common. Gland followers are also usually split to allow easy replacement. They are usually bronze but may be steel or cast iron. Special purpose gland followers are used with volatile or hazardous materials, including capacity for diluting and safely flushing away leakage. The axial compression on the packing must be occasionally adjusted to control leakage. The correct leakage rate is one drip per second. Over-tightening should be avoided as it will result in over-heating and shaft wear. Most packing is supplied with impregnated lubricant, and over-tightening will press the lubricant out. Pumps need extra sealing provisions if pressure at the inner end of the stuffing box is greater than 75psi. The use of harder packing material on the inner rings may help. The procedure for replacing packing is: 1. Read the instructions provided by the pump manufacturer and packing supplier. 2. Loosen and remove gland follower. Inspect gland follower for wear, corrosion, warping etc. 3. Remove old packing rings using a packing puller, and the lantern ring. 4. Inspect shaft sleeve surface for deterioration, and clean up where possible. Replace if necessary. 5. Inspect bore of stuffing box for corrosion, wear, scaling etc, and clean up where possible. 6. Verify correct packing size to be used. 7. Tightly wrap the correct number of packing coils around a mandrel of equal diameter to the shaft sleeve. 8. Cut each ring at an oblique angle. 9. Install each ring, staggering the joins 90 degrees on subsequent rings. Suction / Intake Design Centrifugal pumps operate most efficiently when the liquid to be pumped flows into the inlet nozzle in a smooth, uniform manner with minimal turbulence. Suction systems need to be designed to ensure that this happens. The most common problems are: ï‚ · Uneven / turbulent flow ï‚ · Vapour collection ï‚ · Vortex formation Suction piping should be as short and straight as possible to minimise friction, and if unavoidably long, should be of large diameter. The suction line will normally be at least one pipe size larger than the pump inlet flange, requiring fitment of a reducer. A reducer should not change the pipe bore by more than 100mm. Fluid flow should be as uniform as possible right up to the pump inlet flange. There should not be any fittings likely to cause turbulence, sudden changes in flow direction or spin within ten pipe diameters of the pump inlet flange. There should be no short radius elbows at all, and no long radius elbows within three pipe diameters. All suction line connections need thorough sealing to prevent air being drawn in. For suction manifolds serving multiple pumps, all the above points apply, and branches should be angled at 30 or 45 degrees, rather than ninety degrees, and sized so that fluid flow is constant throughout. Flow should not exceed 0.9m/s. Improper suction conditions or designs can result in the fluid swirling as it approaches the pump through the suction pipe. This is called â€Å"pre-rotation†. It causes a drop in pumping efficiency because the pump is designed to process fluid that is entering without rotation, and can cause additional suction pipe wear. Sometimes a radial fin is fitted to the suction pipe or casing to reduce pre-rotation. The suction pipe design should cater for elimination of air from the suction line, and prevention of vapour pockets, in the simplest manner, meaning that: ï‚ · For pumps with the feed being drawn from a level below (eg a dam pump), o Suction pipe should have a slightly upward slope toward the pump o The eccentric reducer should have the flat side on top ï‚ · For pumps with the feed being drawn from a level above (eg a thickener underflow pump), o Suction pipe should have a slightly downward slope toward the pump Vortexing in feed tanks needs to be avoided to prevent air being drawn down into the pump. Baffles may need to be fitted to tank walls. The tank fluid level needs to be kept well above the suction inlet. Bearings Bearings provide axial and lateral restraint to the pump shaft and attached components, while allowing free rotation. Axial loading on pump shafts may be significant as discussed separately, and the bearing arrangement always includes some thrust capability. The bearings most commonly used are deep-groove single row ball bearings, and single or double row angular contact ball bearings. Pumps may be in overhung configuration, where the shaft is supported by bearings on one side only, or have a shaft that passes right throught the casing with bearings on both sides. Most slurry pumps are of the overhung design. The bearings are usually rolling-element, but plain journal bearings are sometimes used on larger pump sizes. The bearings must be lubricated by grease injection or oil bath and may need provisions for cooling as well. This may be by having a cooling water jacket integral with the bearing housing, or by pumping the lubricating oil through a heat exchanger and filter. Oil lubrication is usually recommended rather than grease, if speed exceeds 5000rpm (which is very rare in a slurry pump). Grease-packed bearings should have one third of the chamber filled with grease. Oil baths should be filled to the centre point of the lowest rolling element. Inadequate loading of bearings can result in the rolling elements skating over the race instead of rolling, and this can cause heating and failure. To avoid this, bearing assemblies are usually designed with an assembly configuration, including preload, that ensures all bearings carry some load. Frame and Foundations For large pumps that are directly connected (ie no vee-belt drive), the motor and pump are usually mounted on the same bed-plate, which is fixed to the foundations in a way sufficient for eliminating looseness and distortion. This eliminates some misalignment issues at the source. Foundations including bed-plates should be checked occasionally for deterioration (corrosion, ground subsidence, concrete cracking, loose fasteners, missing grout etc), and the alignment between pump and motor should also be checked if there is any cause for concern. The framework should have provisions for drainage of any spillage and seal leakage etc, so that this does not become trapped and contribute to corrosion etc. Where pumps operate at high temperature (ie above around 100C) the pump casing should be supported at its axial centre-line, to help reduce thermal stresses. It is generally preferred that all suction and discharge piping have its own supports, so that the pump casing and foundations do not carry any significant static or dynamic piping loads, and so that pump components can be independently removed and replaced. Where this is not the case, extra pump and foundation attention may be needed at the design stage. Drive Arrangement Many drive arrangements are possible to suit the circumstances. Electric motor drive is the most popular, followed by internal combustion engines. Variable speed drives are sometimes necessary and often convenient, but always more expensive and less reliable. In minerals handling plants, slurry pumps are most often electric motor driven, with belt drives. Belt drives allow speeds to be changed through minor modifications – ie pulley changes. Short, low head slurry system designs usually provide motors that are 10 to 20% oversized, to cater for any under-estimates in slurry or system characteristics such as viscosity and friction, and to allow for minor system modifications during the service life. Instrumentation Pumps may be controlled to allow: ï‚ · Variation of flow rate, pressure, liquid level ï‚ · Protection against damaging operating conditions ï‚ · Flexibility in matching pumping performance to duty For centrifugal pumps, control is usually accomplished by speed setting (including turning off/on), or valve setting. This may be manual or automatic. For slurries, control by throttling valve is rare due to the wear rates that usually result. Typical instrumentation includes: ï‚ · Tank / sump level switches ï‚ · Pressure sensors ï‚ · Flow sensors ï‚ · Density sensors In each case, protection from damage by the slurry is critical. This is commonly achieved by using sensors that do not need to contact the slurry eg nucleonic density sensors mounted outside the pipe, with source on one side and detector on the other. Ideally, it is good to have instrumentation available, either permanently mounted or portable, to: Verify operation at BEP, by measuring the difference between suction and discharge pressure Determine flow Ensure that NPSH is sufficient to prevent cavitation Compare flow to motor amperage, to identify when impeller adjustment is needed Need to search more on valves for slurry applications. Notes on Material Selection Where there is some chance of parts coming into contact during pump operation, thought should be given towards minimising the damage that may result. An example of this is at the wear-ring / impeller interface. Studies have shown that damage can be minimised by manufacturing adjacent components from materials that:   Are dissimilar, except where known to be resistant to adhesive wear and galling   Have a difference in hardness of at least 10Rc, if either has hardness less than 45Rc Because it may be difficult to always prevent cavitation from occurring, impellers are usually made of cavitationresistant materials such as chrome-manganese austenitic stainless steel, carburised 12% chrome stainless steel, cast nickel-aluminium bronze, etc. Obviously corrosion resistance is another key selection factor that these materials satisfy. Slurry pumps are subject to heavy wear in the form of abrasion and erosion. The aggressiveness of the slurry is determined by the hardness of the particle s in the slurry, their shape (rounded or sharp), the pulp density, and the size distribution. Slurries can become less aggressive as they travel through a minerals processing plant as the sharp edges become rounded off. Velocity and angle of impingment are also very important factors affecting the resultant wear rates, with wear rate being proportional to velocity squared according to some references. The impingement angle associated with maximum wear rate seems to be dependent on the hardness and brittleness of the material being struck. For very hard / brittle materials it is between 65 and 90 degrees, while for more ductile materials it may be around 25 degrees. Pumps handling slurries with greater than 6mm particle size are usually lined with rubber. However, if impeller tip speed exceeds 28m/s, rubber becomes subject to thermal degradation, and this usually restricts the use of rubber to a maximum head of 30m per stage. Metal lined pumps may be used up to 55m head per stage. For wet end components, materials that may be used to resist wear include Ni-resist, carburised and hardened 12% chromium steel, etc. White iron slurry pump components, which includes Ni-Hard, are restricted to impeller tip speeds of about 36m/s to avoid maximum disc stresses. Steel components are softer but can run at higher speeds, up to a tip speed of 45m/s. Centrifugal pumps are subject to cyclic loads due to such things as imbalance, unbalanced radial forces, fluctuating axial thrust, the vibration induced as each vane passes the cut-point, and variations in upstream and downstream fluid pressure and flow. This sets the scene for fatigue loading, which becomes more of an issue if the slurry is corrosive. Fretting may occur between assembled components where looseness is allowed to develop. This is best avoided through the use of correct manufacturing dimensions and surface finishes, good fitting practice etc. The materials commonly used for pump components include:   Impellers (require castability, weldability, and resistance to corrosion, abrasion, and cavitation) o Bronze, for non-corrosive liquids below 120C o Nickel-aluminium bronze, for higher speed and mildly corrosive applications o Cast iron, for small low-cost applications o Martensitic stainless steel, where added resistance to cavitation, wear, corrosion (other than salt water) or high temperatures may be required o Austenitic stainless steel (mostly cast 316 grade), where a higher level of corrosion resistance is needed. Austenitic stainless steel with 6% molybdenum is often used for salt water pumping.   Casings (require strength, castability and machinability, weldability, and resistance to corrosion and wear) o Cast iron o Cast steel, where extra strength is required ie for pressures above 6000kPa (1000psi) and temperatures above 175C. o Austenitic cast irons with 15 to 20% nickel (Ni-Resist) may be used where abrasion and corrosion are issues. o Bronze, for water applications o Stainless steel, where corrosion is a major issue – martensitic for higher pressures in mildly corrosive fluids, austenitic for more aggressively corrosive fluids. ï‚ · Shafts (require resistance to fatigue and corrosion) o Mild steel, where corrosion and fatigue are minor issues Low alloy steel such as 4140 for added strength Martensitic stainless steel, where added strength and corrosion resistance are needed Shafts are usually chrome-plated, and care is needed to avoid this adding to the fatigue susceptibility through micro–cracking and hydrogen embrittlement. Shafts can be shot-peened prior to plating, and heat-treated afterward to reduce these effects. Wear rings (require castability and machinability, and resistance to corrosion, abrasion and galling) o Bronze for clean liquids and temperatures up to 120C o Stainless steel for applications with abrasion, corrosion and high temperatures – but steps must be taken to avoid galling should the rings come into contact eg increased clearances, hardness differences etc. o o o Impellers other than those made from martensitic stainless steel can usually be repaired by welding, although in some cases this needs to be followed by specific heat treatment processes. In all cases, more exotic (and expensive) materials may be used for specific applications. Material selection is often a balancing act between optimising purchase cost and maintenance / operations performance. Where high temperatures are involved, material selection must take into account differences in expansion rates. Unlined slurry pump impellers and casings are often made from abrasion-resistant cast irons as per ASTM A532, which includes Ni-Hard. These materials consist of a martensitic matrix with secondary hard phases of chrome and iron carbides that increase wear resistance. They cannot be machined or welded, and tend to be prone to corrosion, and breakage through mechanical impact and thermal shock. Brittleness may be reduced by annealing, but this reduces wear resistance. Slurry pump impellers and casings may be lined with softer materials like rubber, where high temperatures can be avoided. These can reduce wear rates by absorbing the impact energy of the particles, while resisting corrosion. Problems may arise in bonding of the rubber at the cut water point, and on the impeller. The lining reduces the thickness of the metal section of the component, so stronger materials are usually used eg steel rather than cast iron. Manufacturers develop their own specifications for ideal liner thicknesses based on experience, but one reference suggests a volute liner thickness of 4% to 6% of impeller diameter. Natural rubbers seem well suited for wear liners for use with slurries with less than 6mm particle size for the impeller, and 15mm particle size for the volute. Provided the base materials are suitable, patches of high wear on wet end parts can sometimes be repaired by welding / hard-facing. However, this increases the likelihood of cracking. Also if the welding results in uneven surfaces in critical points, the added turbulence can accelerate further wear. Many types and styles of surface coating have been tried, with some success. These include thermal spray coatings, diffusion surface treatments, spraying and trowelling of epoxies, etc.