Post on 06-May-2015
description
Bachelor Thesis
Centrifugal Pumps
by Christian Allerstorfer
Supervised by
Univ.‐Prof. Dipl.‐Ing. Dr.mont. Franz Kessler
Directory I
Contents
1) Abstract ..................................................................................................................................................... 1
2) Abstract [German] .................................................................................................................................... 1
3) Introduction .............................................................................................................................................. 2
4) Definition .................................................................................................................................................. 5
5) The principal of centrifugal pumps ........................................................................................................... 5
6) Pump design .............................................................................................................................................. 6
7) Pump assembly ......................................................................................................................................... 8
Casing ............................................................................................................................................... 8
Impeller ............................................................................................................................................ 9
Shaft ................................................................................................................................................ 11
Bearings .......................................................................................................................................... 11
Sealing ............................................................................................................................................ 12
8) Pump parameters and selection ............................................................................................................. 13
Total dynamic head (TDH) .............................................................................................................. 13
Flow rate (Q) ................................................................................................................................... 13
Net positive suction head (NPSH) ................................................................................................... 13
Specific speed (ns) ........................................................................................................................... 14
Power and Efficiency (P, η) ............................................................................................................. 15
Pump characteristic curve .............................................................................................................. 16
Affinity laws .................................................................................................................................... 17
System characteristic curve ............................................................................................................ 19
Pump selection ............................................................................................................................... 20
Example 1 ....................................................................................................................................... 22
Example 2 ....................................................................................................................................... 28
9) Problems at centrifugal Pumps ............................................................................................................... 31
Cavitation ........................................................................................................................................ 31
Solids and slurry handling (abrasive medias) ................................................................................. 32
corrosion ......................................................................................................................................... 33
10) Comparison centrifugal pumps vs. Piston pumps ................................................................................... 33
11) Standards ................................................................................................................................................ 35
12) Conclusion ............................................................................................................................................... 35
13) References .............................................................................................................................................. 36
Directory II
Table of Figures and Equations
Figure 1 – Pump categories ..................................................................................................................................... 2
Figure 3 – Mud Cleaning Unit (NGM Technologies) ................................................................................................ 3
Figure 2 – Mud circulation rotary drilling ............................................................................................................... 3
Figure 5 – Pump station, by Warren pumps in west china (left) and Trans Alaska Pump Station (right) ............... 4
Figure 4 – Trans‐Alaska‐Pipeline topographic map ................................................................................................. 4
Figure 6 – Principle of a centrifugal pump .............................................................................................................. 5
Equation 1 – Bernoulli principle .............................................................................................................................. 6
Figure 7 – Single and double suction pump ............................................................................................................ 6
Figure 8 ‐ Multistage pump (Goulds pumps ‐ model 3600) .................................................................................... 7
Figure 9 – Deep well pump (Goulds pumps ‐ model VIT‐FF) ................................................................................... 7
Figure 10 – horizontal splitted casing of a double suction pump (lower part) ....................................................... 8
Figure 11 – open impeller ....................................................................................................................................... 9
Figure 12 ‐ loss compensation .............................................................................................................................. 10
Figure 13 – enclosed impeller ............................................................................................................................... 10
Figure 14 – pump’s crank shaft ............................................................................................................................. 11
Figure 15 – bearing properties .............................................................................................................................. 11
Figure 16 – mechanical single seal ........................................................................................................................ 12
Equation 2 – total dynamic head .......................................................................................................................... 13
Equation 3 – flow rate ........................................................................................................................................... 13
Equation 4 ‐ NPSH ................................................................................................................................................. 14
Equation 5 – specific speed ................................................................................................................................... 14
Figure 17 – Impeller design over specific speed ................................................................................................... 14
Equation 6 –power ................................................................................................................................................ 15
Equation 7 ‐ efficiency ........................................................................................................................................... 15
Figure 18 ‐ Pump characteristic sheet (Gould pumps – model 3196) ................................................................... 16
Equation 8 – affinity laws (constant impeller diameter) ....................................................................................... 17
Equation 9 – affinity law (constant rotation speed) ............................................................................................. 17
Figure 19 ‐ approximate pump characteristic curve (Goulds pumps – model 3196 at different RPMs) ............... 18
Figure 20 – Examples of hydraulic systems ........................................................................................................... 19
Figure 21 ‐ System characteristic curves ............................................................................................................... 20
Figure 22 – System & Pump Characteristic curve ................................................................................................. 21
Figure 23 ‐ Borehole .............................................................................................................................................. 22
Figure 24 – friction coefficient for OSTWALD fluids .............................................................................................. 23
Figure 25 – pressure loss in manifold systems ...................................................................................................... 25
Figure 26 – Pump selection software, criteria definition ...................................................................................... 26
Directory III
Figure 27 – example of results provided by goulds pumps pums selection tool .................................................. 26
Figure 28 ‐ Hydrocyclone, working principle ......................................................................................................... 28
Figure 29 – recommended manifold system ........................................................................................................ 28
Figure 31 – regions of impeller cavitation ............................................................................................................. 31
Figure 30 – bubble collapse .................................................................................................................................. 31
Figure 32 – typical impeller wear due to cavitation .............................................................................................. 32
Figure 33 – piston pump ....................................................................................................................................... 33
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1) Abstract
Aim of this thesis is to give an overview on centrifugal pumps in general and especially in applications within the petroleum industry. There is a wide range of pumps available but as the radial pump is by far the most prolific member of the pump family so this paper will concentrate on them. It will first explain the principal of centrifugal pumps; its types of construction, which bandwidth of pressures and flow rates are available and how to choose the right pump for a specific application. Also some comparison with another big family of pumps, the piston pumps, is made. Later chapters deal with typical problems when using centrifugal pumps such as cavitations and corrosion. Note that this is my first bachelor thesis during my studies of Petroleum Engineering. It is meant as a literature research to scientifically handle a specific topic and to define the state of the art. All sources are listed at the end of the document in the chapter references.
2) Abstract [German]
Ziel dieser Arbeit ist es einen Überblick über Zentrifugalpumpen im allgemeinem und besonders in Anwendungen der Erdölindustrie zu vermitteln. Für industriele Anwendungen sind heutzutage viele verschiedene Pumpentypen verfügbar, aufgrund der weiten Verbreitung von Zentrifugalpumpen wird sich diese Abhandlung auf diese konzentrieren. Zuerst wird auf Aufbau, Prinzip und Konstruktionsvarianten ebenso wie auf verfügbare Bandbreiten in Druck und Durchfluss sowie Pumpenwahl eingegangen. Weiteres werden Zentrifugalpumpen den Kolbenpumpen gegenübergestellt. Spätere Kapitel behandeln typische Probleme welche beim Betrieb dieser Pumpen auftreten wie Kavitation und Korrosion. Man beachte das dies die erste Bakkalaureatsarbeit während meines Petroleum Engineering Studiums ist. Es wird als Literaturrecherche verstanden und dient dazu sich mit dem wissenschaftlich bearbeiten eines vorgegebenen Themas zu befassen. Alle Quellen sind am Ende des Dokuments im Kapitel “References” angeführt.
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3) Introduction
A pump is a machinery or device for raising, compressing or transferring fluid. A fluid can be gasses or any liquid. Pumps are one of the most often sold and used mechanical devices and can be found in almost every industry. Due to this there is a wide range of different pumps available. In general, the family of pumps is separated into positive displacement and kinetic pumps. A subcategory of kinetic pumps are centrifugal pumps which are again separated into radial pumps, mixed flow pumps and axial pumps. But even at the axial end of the spectrum there is still a part of the energy coming from centrifugal force unless most of the energy is generated by vane action. On the other hand side in radial pumps almost all the energy comes from centrifugal force but there is still a part coming from vane action. There are also several pumps combining both principles placed somewhere in between the two extremes in the centrifugal pump spectrum known as mixed flow impellers. Characteristic for radial pumps are low specific speeds. As shown in the diagram below there are many options in pump design, which will be discussed in detail in later chapters.
Figure 1 – Pump categories
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Figure 2 – Mud circulation rotary drilling1
Within the petroleum industry pumps are necessary to process fluids especially hydrocarbons. Another important application within the petroleum industry is in the mud circuit on a drilling rig. On drilling rigs, mud which consists mainly of water and bentonite as well as of several different additives depending on many different factors is used. The heart of the mud circuit is the mud pump which is in general a high pressure piston pump. It provides the major part of head to overcome the systems resistance. The mud is pumped through a piping system to the derrick and through the standpipe to a certain high. Now through the kelly hose via the gooseneck into the upper kelly cock. It flows through the kelly and the lower kelly cock into the drill string down the borehole. At its end, the mud leaves the drilling collars through the drilling bit. The mud pressure is increased by its nozzles and released into the borehole (fig.21). The mud cools the bit and collects the cuttings to transport them up to the surface where the mud is cleaned. It leaves the borehole and is forced through the BOP Stack and the chock manifold system. Now bigger cuttings are removed in the shall shaker and the mud is collected in the settling pit. It is now pumped though a degasser to remove any gasses collected from the borehole to avoid explosions. After degassing, the sand is removed in a desander and the mud is processed to the mud cleaner. It consist of several desilters. Here small cuttings even smaller than 74µm, are removed. Desander and desilter are so called hydrocyclones of different sizes, commonly charged by centrifugal pumps. At the end of the mud conditioning circuit, a centrifuge is located to remove anything left. The mud is now stored in tanks and kept in motion by nozzles or agitators. Finally the mud is sucked through the hopper to the mud pump by another centrifugal pump. To sum up, centrifugal pumps can be found on several locations within the mud circuit of a drilling rig like to charging degasser, desander, mud cleaner as well as the mud pump. On rigs centrifugal pumps can also be found as fuel or cooling water pumps for e.g. diesel engines
Figure 3 – Mud Cleaning Unit (NGM Technologies)2
1 www.q8geologist.com (modified) 2 Neftegazmash‐Technologies (modified)
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Other typical applications for centrifugal pumps are pipeline applications. Pipelines are used for economical transport of hydrocarbons like oil and gas over long distances. At the beginning of a pipeline system, in most cases huge storage tanks can be found to ensure a continuous flow through the pipeline. The oil is forced through the pipe by a few powerful centrifugal pumps in serial. On its long way, pumping stations are required to overcome the resistance and heights. These pumping stations are distributed over the whole length of the pipeline, but can be found especially before mountains. As an example, the 1280km long Trans‐Alaska pipeline has 11 pumping stations with 4 pumps each. Usually only 7 stations are in operation and provide the head to overcome height differences and the fluid – pipe friction. The other 4 pump stations are on standby and are activated if necessary to ensure sufficient head at peak loads. The pipeline has a maximum capacity of around 330.000m³ per day. So it is obviously that pipelines are a perfect application of high capacity pumps like centrifugal pumps. There are also several valves to control the flow or to shut in the pipeline in case of an accident along it. On the map (fig.41) it can be easily seen that the pump stations are not distributed regularly over the pipeline’s length. At the end of a pipeline, usually a distributing station like a major harbour or refineries can be found. In case of the Trans‐Alaska Pipeline, it is the harbour in Valdez to distribute the oil from the Prudhoe Bay Oil Field to up to four tankers simultaneously.
Figure 5 – Pump station, by Warren pumps in west china (left) and Trans Alaska Pump Station (right)2
These are just examples for the wide range of applications of centrifugal pumps within the petroleum industry. Also important are applications within the hydrocarbon processing industry and on offshore rigs or distributing stations at harbours.
1 www.Nationalatlas.gov (modified) 2 Warren Pumps (left), Howard C. Anderson (right)
Figure 4 – Trans‐Alaska‐Pipeline topographic map1
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4) Definition
Symbol Unit Definition Symbol Unit Definition
D m Impeller diameter ρ kg/m³ density
z m height (pos. upwards from PCL) η ‐ efficiency
ps bar pressure suction flange Q m³/s flow rate
p d bar pressure discharge flange pv bar vaporize pressure
p e bar pressure environment (1bar) P W electric power
g m/s² acceleration of gravity (9,81m/s²) NPSHA m NPSH ‐ available
v m/s velocity NPSHR m NPSH ‐ required
n 1/min rotation per minute H m Head
Shortcut Description
TDH total dynamic head
N PSH net positive suction head
BEP best efficiency point
PCL pump centre line
index: 1,2 suction side, discharge side
5) The principal of centrifugal pumps
A centrifugal pump is a rotodynamic pump that uses a rotating impeller to increase the pressure of a fluid. The fluid enters the pump near the rotating axis, streaming into the rotating impeller. The impeller consists of a rotating disc with several vanes attached. The vanes normally slope backwards, away from the direction of rotation. When the fluid enters the impeller at a certain velocity due to the suction system, it is captured by the rotating impeller vanes. The fluid is accelerated by pulse transmission while following the curvature of the impeller vanes from the impeller centre (eye) outwards. It reaches its maximum velocity at the impeller’s outer diameter and leaves the impeller into a diffuser or volute chamber (fig.6).
Figure 6 – Principle of a centrifugal pump1
1 ITT – Goulds Pumps (modified)
6
So the centrifugal force assists accelerating the fluid particles because the radius at which the particles enter is smaller than the radius at which the individual particles leave the impeller. Now the fluid’s energy is converted into static pressure, assisted by the shape of the diffuser or volute chamber. The process of energy conversation in fluids mechanics follows the Bernoulli principle (eqn.1) which states that the sum of all forms of energy along a streamline is the same on two points of the path. The total head energy in a pump system is the sum of potential head energy, static pressure head energy and velocity head energy.
2 21 1 2
1 22 2v p v pz z 2
g g g gρ ρ+ + = + +
⋅ ⋅ ⋅ ⋅ Equation 1 – Bernoulli principle
As a centrifugal pump increases the velocity of the fluid, it is essentially a velocity machine. After the fluid has left the impeller, it flows at a higher velocity from a small area into a region of increasing area. So the velocity is decreasing and so the pressure increases as described by Bernoulli’s principle. This results in an increased pressure at the discharge side of the pump. As fluid is displaced at the discharge side of the pump, more fluid is sucked in to replace it at the suction side, causing flow.
6) Pump design
Back in 1475, the Italian Renaissance engineer Francesco di Giorgio Martini describes a water or mud lifting machine in one of his treatises that can be characterised as the first prototype of a centrifugal pump. The first true centrifugal pump was invented by the French physician Denis Papin in 1689, when he was experimenting with straight vane impellers. British inventor John Appold introduced the first curved vane impeller in 1851. Nowadays only curved impellers are used in 3 different types. There are pumps with open, semi‐open and enclosed impellers. Open impellers only consist of blades attached to its eye as semi‐open ones are constructed with a disc attached to one side of the vanes. Enclosed impellers have discs attached to both sides of the vanes. Impellers are also classified based on the number of points where the fluid can enter the pump. There are single suction, which allow the fluid to enter its centre from only one side, as well as double suction impellers which can be entered by fluid from both sides simultaneously. These types of construction are also known as overhung impeller pumps and impeller between bearings pumps.
Figure 7 – Single and double suction pump1
1 ITT – Goulds Pumps
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Another option in centrifugal pump design is single stage and multistage design (fig.9). Single stage pump is the standard centrifugal pump design, equipped with only a single impeller. Multistage pumps on the other hand consist of two or more impellers fitted to the same shaft in a single casing. Multistage pumps work like two or more pumps operating in serial. Therefore multistage pumps are most suitable in low flow rate and high TDH applications.
Figure 8 ‐ Multistage pump (Goulds pumps ‐ model 3600)1
Centrifugal pumps can also be separated into horizontal pump design and vertical pump design (fig.8). Vertical centrifugal pumps are especially used as submerged or in well pumps. Another point when talking about centrifugal pumps is priming. Every centrifugal pump has to be primed as it is not able to suck any fluid as long as the impeller is filled with air. This is because air is approximately 1000 times lighter than for example water. So to suck water into the pump to prime itself, for every meter it would have to be able to produce a TDH of 1000m.Due to the fact that conventional centrifugal pumps are not able to produce a TDH in that order of magnitude, most centrifugal pumps have to be primed either with an extra device, for example a vacuum pump or a special design of the pump casing. Due to the wide range of design variations where most of them are combinable in many different ways there is a huge range of centrifugal pumps available beginning with standard single suction, single stage, non self priming pumps up to double suction high flow rate, multistage pumps for high pressures or self priming pumps for special applications.
1 ITT – Goulds Pumps
Figure 9 – Deep well pump (Goulds pumps ‐ model VIT‐FF)1
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7) Pump assembly
In this chapter, the main parts, a centrifugal pump consist of are discussed. These are the casing, the impeller, shaft, bearings and seals.
Casing The pump’s casing (fig.91) houses the hole assembly and protects is from harm as well as forces the fluid to discharge from the pump and convert velocity into pressure. The casings design does not influence TDH but is important to reduce friction losses. It supports the shaft bearings and takes the centrifugal forces of the rotating impeller and axial loads caused by pressure thrust imbalance. Most of all centrifugal pumps are of simple spiral casing and are not equipped with a guide vane aperture. Even if this would increase efficiency due to the simplicity of spiral casings, this is the preferred type of construction. Only extraordinary big or multistage pumps do have guide vanes. The spiral pump casing has to be carefully designed to avoid turbulences resulting in a decrease in efficiency. The shape of the casing is defined by several factors; these are profiles angles, diameter and width. The whole amount of fluid flows through the discharge cross section, the amount of fluid is decreasing when going backwards in the spiral, from point of view of flow direction. Therefore the area of the profiles is decreasing continuously as well, to fit the flow rate in the specific point of the pump casing. The result is a spiral shaped casing. The optimum properties of the spiral were found in experiments and expressed in formulas and diagrams. The fluid velocity is not constantly distributed over a certain profile section. Modern Pumps are designed for a constant pressure and constant mean velocity in every profile section at the BEP. Apart from the BEP, the radial forces are out of balance resulting in a total radial force different to zero. This is important because the radial force bends the pump’s shaft and results in higher wear at seals and could lead to shaft fatigue. To reduce most radial forces the pump casing can be designed as a double spiral casing. In this case the flow is spitted into two parts. Due to symmetry reasons almost all radial forces chancel each other out. Another important part of the pump’s casing are elbows in multistage pumps to deflect the flow from the previous stages discharge side to the suction side of the following. If a multistage pump is equipped with guide vanes, no elbows are necessary. As already mentioned, guide vane construction is only common at big or horizontal multistage pumps. Guide vanes work as a diffuser and convert the increased fluid velocity into pressure. It consists of extending channels arranged around the impeller. To ensure adequate pump life time, the pump’s casing material should be selected carefully. Standard pump casings are made of cast iron but due to the fact that cast iron is not that resistant against cavitation, many pumps are coated or made from more wear resistant materials. Due to vibrations the casing should have good damping properties. Pump casings are splitted either axial or radial to allow assembling and maintenance.
1 www.rumfordgroup.com
Figure 10 – horizontal splitted casing of a double suction pump (lower
part)1
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Impeller The impeller is the essential part of a centrifugal pump. The performance of the pump depends on the impeller diameters and design. The pump’s TDH is basically defined by the impeller’s inner and outer diameter and the pump’s capacity is defined by the width of the impeller vanes. In general, there are three possible types of impellers, open, enclosed and semi open impellers, each suitable for a specific application. Standard impellers are made of cast iron or carbon steel, while impeller for aggressive fluids and slurries require high end materials to ensure a long pump life.
o Open impeller Open impellers (fig.91) are the simplest type of impellers. They consist of blades attached to the hub. This type of impeller is lighter than any of the other type at the same diameter. Weight reduction leads to less force applied to the shaft and allows smaller shaft diameters. These results in lower costs compared to equivalent shrouded impellers. Typically, open impellers operate at higher efficiency because there is no friction between the shrouds and the pump casing. On the other hand side, open impellers have to be
carefully positioned in the casing. The gap between the impeller and the surrounding casing should be as small as possible to maximise efficiency. As the impeller wears the clearance between the impeller and the front and back walls open up, what leads to a dramatic drop in efficiency. A big problem when using a pump with an open impeller are abrasives. Due to the minimized clearance between blades and casing, high velocity fluids in close proximity to the stationary casing establish vortices that increase wear dramatically.
o Semiopen impeller Semi‐open impellers can be seen as a compromise between open and enclosed impellers. A semi‐open impeller is constructed with only one shroud, usually located at the back of the impeller. It usually operates at a higher efficiency than an equivalent enclosed one due to reduced disc friction as there is only one shroud. A big advantage of semi‐open impellers compared to open ones is that the impellers axial position can be adjusted to compensate for wear. A problem is that the entire backside of the impellers shroud is under full impeller discharge pressure as the front side is under suction pressure increasing along the impeller radius due to centrifugal force. The differential between these pressures causes an axial thrust imbalance. Manufactures try to reduce this effect by applying vanes to the back side of the impeller. But the efficiency of these so called “pump out vanes” decreases if the impeller is moved forward to compensate for wears. A better option to compensate the loss of efficiency is an adjustable wear plate, so that clearance adjustments can be made. Semi‐open impellers are also easily to manufacture as all sides of the impeller are easy accessible for manufacturing processes as well as for applying surface hardening treatments. In combination with wear compensation applications, semi‐open impellers can be used for intermediate abrasive fluids. Another advantage if using semi closed impellers in combination with an adjustable wear plate compared to an open impeller equipped with the same wear compensation system is vane support. This prevents the vanes from collapsing or deformation when using it with fluids contaminated by solids. This justifies the application of semi‐open impellers even thought it seems logically to use an open impeller due to its reduced weight.
1 Hwww.mcnallyinstitute.com (left), ITT – Goulds Pumps (right)
Figure 11 – open impeller1
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o Enclosed impeller Enclosed impellers (fig.121) consist of blades covered by a front and back shroud. The fluid steams through the impeller without interacting with the stationary pump casing. In a well designed enclosed impeller, the relative velocity between the fluid and the impeller walls at any given radius is rather small. The disc friction of the shrouds rotating in close proximity to the pump casing causes a lower efficiency as comparable semi‐open or open impellers. A problem when dealing with enclosed impellers is leakage between the impeller
shrouds and the pump casing back to the suction side of the pump. There are two common ways for controlling leakage in enclosed impeller pumps (fig.132). One are wear rings in combination with impeller balance holes. But the tight clearance between the rotating and the stationary wear ring causes high fluid velocities and therefore a high wear rate. Wear ring lifespan is unacceptable short in an abrasive environment. If wear rings reach the end of their intended lifespan, it has to be replaced because if it is not the high velocity zone can shift from the wear ring into the impeller thrust balance holes. This could cause significant damage to the impeller and may result in an expensive repair or replacement of the impeller. So this is only an option when dealing with moderate abrasive fluids with light solids only. The other possibility to control wear and axial thrust balance are pump‐out vanes. These pump‐out vanes cause much lower local velocities spread over a bigger area resulting in lower wear. It is not uncommon, that pump‐out vane lifespan equals or exceeds the main impeller’s lifespan. The major disadvantage of pump out vanes is their power consumption what leads to a lower efficiency. Overall pump‐out vanes provide a good pump characteristic when dealing with abrasive solids. Another problem when operating an enclosed impeller in combination with fluids contaminated by large solids like rocks is that it may happen that a piece of solid gets caught in the impeller eye outlet. This may cause a mechanical or hydraulically imbalance and has the potential to damage the pump. In an open or semi‐open impeller this rock would be broken by the grinding between the rotating impeller and the stationary casing. To remove the blockage disassembling of the pump would be necessary.
1 http://knowledgepublications.com (left), www.engineersedge.com (right) 2 Lawrence Pumps Inc., RunTimes jan.05 issue (modified)
Figure 13 – enclosed impeller1
Figure 12 ‐ loss compensation2
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Shaft The shaft is the connection between impeller and drive unit which is in most cases an electric motor but can also be a gas turbine. It is mainly charged by a radial force caused by unbalanced pressure forces in the spiral casing and an axial force due to the pressure difference between front and backside of the impeller. Most common pump shafts are made of carbon steel. There are several cranks to support the bearings and seals. A high surface quality and small clearances are required. Especially in the areas of the bearing’s, clearance and surface quality is important to ensure right positioning of the shaft in the casing and therefore close positioning clearances of the impeller. At the area of the seals, particularly the surface quality is important to ensure an adequate seal lifespan. In shaft design it is also important to avoid small radiuses at cranks to minimize stress in these areas which are susceptible for fatigue.
Figure 14 – pump’s crank shaft1
Bearings The bearings keep the shaft in place to ensure radial and axial clearance. Some approximate bearing properties can be seen in fig.142. The bearings lead radial and axial forces from the impeller into the casing. In double suction pumps bearings are located at both sides of the impeller as at single suction pumps all bearings are located behind the impeller. In horizontal process pumps, usually oil bath lubricated bearings are used. Medium and heavy duty process pumps are used in refineries, where highest reliability is required. In these pumps axial loads are supported by universal single row angular contact ball bearings. In heavy duty process pumps, also matched taper roller bearings with steep contact angles, arranged face to face or back to back are used to support combinations of high radial and axial loads. In very high duty service and slurry pumps, spherical roller bearings can be used to support very high radial loads. A spherical thrust bearing is used to support axial loads. It is usually spring preloaded to ensure that sufficient load is applied during start up or pump shutdown. At vertical pumps, the thrust bearing can be a ball bearing with a spherical outer ring raceway, with the centre of the radius located on the bearing axis, providing a self‐alignment
1 ITT – Goulds pumps (modified) 2 Pump User’s Handbook (by Heinz P. Bloch, Allan R. Budris)
Figure 15 – bearing properties2
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capability. It is equipped with a 45° contact angle that enables the bearing to support large axial loads and moderate radial loads. If the pump is operated at its BEP, the bearing will only have to carry the rotating assemblies’ weight, the stress due to interference fit of the shaft and in some cases manufacturer dependent preloads. Unfortunately, many bearings are overloaded because of wrong interference fit, shaft bend, solids, unbalanced rotating elements, vibrations, axial thrust and many more. This leads to increased stress and temperature and therefore to a decrease in lifespan. It is also important for the bearing’s lifespan to protect it from fluid by adequate seals.
Sealing To protect the bearings against fluid and prevent leakage, there are several seals fitted into the casing. Nowadays, rotary pumps are equipped with mechanical seals (fig.141). A mechanical seal consists of primary and secondary sealing. In most cases the primary part, which is fitted to the casing, is made of a hard material like silicon carbide or tungsten carbide. The other, the rotating part of the primary seal is made of a soft material like carbon. Both parts are pressed against each other by e.g. a spring. The secondary sealings are not rotating relative to each other and provide a fluid barrier. Mechanical seals can be separated into pusher/non‐pusher seals, seal driving/spring compression, balanced/unbalanced and inside/outside mounting. Pusher seals will have a tendency to “hang up” when handling fluids which crystallize because the secondary seal member is not able to accommodate for travel. Whether applying a balanced or unbalanced seal will effect seal performance. Unbalanced seals see a high pressure at the impeller side and therefore have a reduced fluid film between the seal faces. This leads to overheating, rapid face wear and seal fatigue at early stages. To simplify maintenance many seals are available in cartridges which are pre‐packed seal assemblies. To avoid any leakage when handling hazard fluids, double or tandem seals can be applied. In these seals, a secondary so called containment seal is placed after the primary one. The space in between is filled with a natural fluid called barrier or buffer fluid. These seals are very common in the petroleum industry. The difference between a tandem and a double seal is that in a double seal the barrier fluid is pressurised. Due to this, in case of primary seal fatigue the pressurised barrier fluid streams into the pumps instead of the hazard fluid into the atmosphere. The seal materials must fit the fluid to ensure accurate seal lifespan. The standards of modern mechanical seals are widely defined by API Standard 682 ‐ Shaft Sealing Systems for Centrifugal and Rotary Pumps.
1 US patent 2951719
Figure 16 – mechanical single seal1
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8) Pump parameters and selection
There are several parameters depending on impeller design, diameter, RPM etc., characterising a pump. In this chapter the most important pump parameters will be discussed as well as an example showing how to calculate losses in a system and to select a pump.
Total dynamic head (TDH) Head in general is used to define energy supplied to a liquid by a pump and is expressed in units of length. In absence of any velocity it is equal to the height of a static column of fluid that is supported by a pressure in the point of datum. Total dynamic head (TDH) is the difference between total dynamic discharge head and total dynamic suction head (eqn.2). Total dynamic discharge (suction) head is practically the pressure read from a gauge at the discharge (suction) flange converted to length units and corrected to the pump centre line plus the velocity head at the point of the gauge (eqn.2). These two values represent the total amount of energy of the fluid at the discharge and suction flange of the pump. Mathematically it is the sum of static discharge (suction) head and velocity at the discharge (suction) flange minus total friction head in the discharge (suction) line. The difference of these values gives you the THD which represents the energy added to the fluid. TDH does not depend on the delivered fluids density. A higher density only increases the pressure and therefore the required power at a constant flow rate.
d s2 2
2 1 2 12 1
TDH=h -h
(p -p ) (v -v )TDH=(z -z )+ +ρ×g 2×g
Equation 2 – total dynamic head
Flow rate (Q) (Volumetric) Flow rate is the volume of fluid passing through the pump per unit of time. It is calculated as area times fluid velocity (eqn.3). It depends on the impeller geometry and RPM. Impellers are optimised for highest outlet velocities. Multiplied by the useable impeller inlet area you will get the flow rate. An impeller is designed for a maximised flow rate at a specific speed depending on its diameter. This is called the point of best efficiency.
1 1 2Q=A v =A v2⋅ ⋅
Equation 3 – flow rate
Net positive suction head (NPSH) NPSH is defined as total suction head above the suction nozzle and corrected to datum, less the vapour pressure of the fluid converted into length units. It analyses energy condition on the suction side of the pump to determine whether the liquid will vaporise at the lowest pressure point of the pump. Vapour pressure is a characteristic fluid property increasing with increasing temperature. It indicates the pressure at which a fluid starts boiling, causing bubbles which move along the impeller surface to an area of higher pressure were they collapse rapidly and cause significant harm to it. By decreasing the pressure the temperature at which this happens also decreases. So if the pressure is low enough it is possible to see this effect even at surrounding temperature. This effect is known as cavitations and should necessarily be avoided. It is obvious that in order to pump a fluid in an effective way we have to keep it liquid. Therefore NPSH required (NPSHR) is the total suction head required to prevent the fluid from vaporising at the lowest pressure point of the pump. NPSHR is a function of
14
pump design as the pressure at the impeller decreases by accelerating the fluid along the impeller. There are also pressure losses due to shock and turbulences as the fluid strikes the impeller. To overcome all these pressure drops in the pump and maintain the fluid above vapour pressure a certain positive suction head is required. NPSHR varies with flow rate and speed within any particular pump. The available NPSH is a function of the system in which the pump operates. To avoid cavitations NPSHA must be bigger than NPSHR. In practise the NPSHA can be determined by a gauge on the suction flange of the pump and the following formula (eqn.4). It is also common to add a certain safety value to the NPSHR to make sure that there is enough suction head to prevent the fluid from vaporising. In practice a safety value of 0,5m has turned out to be reasonable.
21
1 e v
A 1
A R
v ρp +p + -p2
NPSH = +Δzρ g
NPSH ³NPSH +0,5m
⎛ ⎞⋅⎜ ⎟⎝ ⎠
⋅ ,2
Equation 4 ‐ NPSH
Specific speed (ns) Specific speed (eqn.5) is a value to characterise the shape of a impeller. Low specific speed characterises a radial impeller and is increasing up to high specific speed at axial impellers. Impellers in between are known as Francis‐vane and mixed‐flow impeller (fig.11). Specific speed is only of designing engineering significance used to predict pump characteristics.
BEPs 3
4BEP
Qn =n
TDH⋅
Equation 5 – specific speed
Figure 17 – Impeller design over specific speed1
1 www.lightmypump.com (modified)
15
Power and Efficiency (P, η) The work performed by a pump is a function of THD, flow rate and the specific gravity of the fluid. Pump input (P) or brake horse power (bhp) is the actual power delivered to the pump shaft. Pump output (Phydr) or hydraulic horse power (whp) is the energy delivered to the fluid per time unit (eqn.6). Due to mechanical and hydraulic losses in the pump, Phydr is always smaller than P. Therefore efficiency is defined as Phydr divided by P (eqn.7). The impeller geometry is optimized to provide highest flow rate at a certain speed at a given diameter at its point of best efficiency (BEP). If operating a pump off its (BEP), losses due to increasing turbulences and recirculation will increase and reduce efficiency. These effects are caused by a mismatch of the pump’s design flow rate and the actual flow rate. The difference between inlet vane angle and approaching flow angle is increasing as moving away from the BEP as well as losses between impeller vane exit and the diffuser. Result of this is an increased flow between the impellers shrouds and the casing.
hydrP =ρ g Q TDH⋅ ⋅ ⋅
Equation 6 –power
hydrP ρ g Q TDHη= =P P
⋅ ⋅ ⋅
Equation 7 ‐ efficiency
16
Pump characteristic curve The pump characteristic curve shows the performance of a pump. It usually shows TDH, power, efficiency and NPSHR plotted over flow rate at a given RPM. There are absolute or dimensional and relative or non‐dimensional plots (fig.12). The difference is that a dimensional diagram shows absolute values, while a non‐dimensional plot shows the data in percent of their values at the pumps BEP. The first line in the diagram shows the pumps THD plotted over flow rate. Characteristic is the slightly decreasing THD at increasing flow rate. The efficiency graph is typically increasing until it reaches its peak at the pumps BEP and drops as flow rate is further increasing. The bhp line is more or less a straight line as it increases with increasing flow rate. It is also possible to plot these functions for several speeds at a given diameter or at different diameters for a given speed in one diagram. Result is a set of pump characteristic curves as provided by most manufactures. In these diagrams you can estimate pump behaviour at constant speeds and a range of impeller diameters. Constant horse power, efficiency, and NPSHR lines are plotted over the various head curves. The pump characteristic curve shown in fig.12 is an example for the what information you can get out of such a diagram. In this example, we assume that we have this pump with an impeller diameter of 7” operating at 3540RPM and a flow rate of 48m³/h. Therefore we can read from the diagram the pump’s current efficiency, head, required power as well as the NPSHR. In this case, our operating point is almost the pump’s BEP and we get THD of 60m, an efficiency of about 61%, required power of 13Hp and a NPSHR of 9ft.
Figure 18 ‐ Pump characteristic sheet (Gould pumps – model 3196)1
1 ITT – Goulds Pumps (modified)
17
Affinity laws These laws express relationships between several variables involved in pump performance such as flow rate, impeller diameter, head and power. There are two ways to express these relationships: either holding the impeller diameter (eqn.8) or the rotation speed (eqn.9) constant. Affinity laws apply to radial pumps as well as axial pumps.
2 3
1 1 1 1 1 1
2 2 2 2 2 2
Q n TDH n P n= =Q n TDH n P n
⎛ ⎞ ⎛ ⎞⎜ ⎟ ⎜ ⎟⎝ ⎠ ⎝ ⎠
=
Equation 8 – affinity laws (constant impeller diameter)
2 3
1 1 1 1 1 1
2 2 2 2 2 2
Q D TDH D P D= =Q D TDH D P D
⎛ ⎞ ⎛ ⎞⎜ ⎟ ⎜ ⎟⎝ ⎠ ⎝ ⎠
=
Equation 9 – affinity law (constant rotation speed)
As an example, assume operating a pump at BEP at n1, we can calculate the BEP for any other n2 (or any other diameter). The efficiency remains almost constant at speed and small impeller diameter changes. At first; we have to determine flow rate, TDH and power for the pumps BEP at 3540RPM from the pump characteristic curve (fig.12). With this knowledge; it is possible to calculate the BEP for 4000RPM and plot a new pump characteristic curve.
31 1
22 2
2
1 12
2 2
3
1 12
2 2
Q n m= Q =42,9Q n h
TDH n= TDH =76,6mTDH n
P n= P =18,75HpP n
→
⎛ ⎞→⎜ ⎟
⎝ ⎠
⎛ ⎞→⎜ ⎟
⎝ ⎠
18
By performing this calculation for several points; we get the pump characteristic curve for the new speed (fig.13).
Figure 19 ‐ approximate pump characteristic curve (Goulds pumps – model 3196 at different RPMs)
This shows that with a change in speed or in impeller diameter; the pumps characteristic can be optimized to fit the system it is operated in.
19
System characteristic curve A system characteristic curve represents the behaviour of the system in which the pump is operated. It defines the point on the pump characteristic curve on which the pump operates. Plotting the system and pump characteristic curve in the same diagram, the point of intersection is the operation point of the pump, operated at a certain speed in a given system. It is also possible to predict the behaviour of the pump during a change either in system or pump properties.
Figure 20 – Examples of hydraulic systems1
System A is a typical piping system with a centrifugal pump to deliver fluid to a higher tank. The difference in system B is that in this case almost all the piping is vertical. This is important because the main losses are caused by friction between the fluid and the pipe’s inner surface. Therefore, losses in system B are smaller than in system A. The system characteristic curves corresponding to the examples (fig 14) are shown below (fig.10). Due to this dependence of friction from velocity, the blue curve, representing system B, is much flatter than the characteristic curve of system A. The red line shows the energy that is required (TDH) to pump the fluid from the lower to the upper tank which are both under ambient pressure. The system characteristic curve is of parabolic shape because it is plotted over flow rate and friction is function of squared velocity. So if flow rate is increasing, also velocity is increasing the same way and losses due to friction increase. Therefore, more energy is required to compensate losses and deliver fluid to the upper tank. Obviously, a throttled valve causes additional resistance and therefore additional losses resulting in more energy required to deliver fluid to the upper tank at the same flow rate or a lower flow rate at constant power.
1 Reinhütte Pumpen (modified)
20
Pump selection First of all, and this is properly the most important part, we have to take a close look at the application of the pump. There should be as much details about the system available as possible, to ensure choosing the right pump. Important selection parameters are required TDH, flow rate, NPSHA, fluid and flexibility of the system. It is also important to know the fluid. Parameters like pH‐value, viscosity, abrasives, fluid and surrounding temperature range as well as quantity, size and shape of solids. If we know that a centrifugal pump is the right pump for the application, we can go into detail searching for a potential pump model. Most manufactures provide a pump selection software, but there are also various manufacturer‐independent software packages available (e.g. www.pump‐flo.com). A pump selection software gives you a choice of pumps, fitting the specification made at the beginning. Many detailed specifications can be made to characterise the application. Most pump selection software are quite powerful tools that also provide calculators for NPSHA and TDH determination. There are also a lot of administration tools like PDF or excel export and file management features implemented. But in general, to select a pump it is useful to plot the curve, characterising the system and the characteristic curve of a potential pump into the same diagram. The point of intersection of the two head curves indicates the operation point of the pump in the system. It is also possible to make predictions how the pump will behave when changing system parameters. Obviously, the operation point should be as close to the BEP as possible. A common rule when selecting a pump is to choose a pump with at least 25% more head available than required by the system. Another common practice is to choose at most the second largest impeller diameter available in a pump series. This is reasonable in the case of a changing system or if there has been made a mistake during pump selection. So it is possible to change the impeller to the next larger size without changing the casing.
Figure 21 ‐ System characteristic curves
21
Figure 22 – System & Pump Characteristic curve
22
Example 1 In this example, the whole calculation for a selected application will be shown. Aim of the calculation is to calculate the required pump parameters and to select a matching pump as well as a proper drive unit. We assume the following equipment and hole (fig.181) properties:
Down‐hole:
part type dimension
hole depth 1100m
Bit diameter 9 12 ” (0,2413m)
Nozzles 3
tubing diameter 10 34 ” (0,2731m)
length 820m
wall thickness 10,16mm
drilling rod Type 6 58 ” FH
wall thickness 8,38mm
outer diameter 0,169m
drilling collar outer diameter 8 12 ” (0,2159m)
inner diameter 2 78 ” (0,07m)
length 120m
Manifold:
part diameter [mm] length [m]
main pipe 100 35
standpipe 100 14
mud hose 75 17
mud head 75 2
Kelly 100 12
Drilling fluid properties:
property value
Type OSTWALD‐fluid
density 1250kg/m³
KFactor 0,28Ns/m²
fluid index n 0,64
recommended speed 0,63m/s
1 Die Bohrspülung (by Gerd‐Ulrich Lotzwick)
Figure 23 ‐ Borehole1
23
First of all, it is required to calculate the flow rate, depending on the cross sections and recommended fluid speed. Therefore it is necessary to determine the cross sections one to three and to calculate the flow rate for the given speed by the in chapter 7 introduced formula.
2 2o i
21
22 2
23
πA= (D -D )4
A =0,02775m
A =0,04341m V=A v
m³A =0,02924m V=98,478h
⋅
In practice, it is common to introduce a safety factor of 20% to the flow rate.
m³V=118,17h
Now we calculate the pressure losses for all parts of the down‐hole assembly as well as in the manifold.
Pressure losses inside the drilling rod:
V m= =1,873A s
v
By calculating the modified Reynolds number, it is possible to determine the friction coefficient from the diagram (fig.191)
2-n n
m mv D ρRe = Re =5875
K 1+3n88 4n
λ=0,027
n⋅ ⋅
⎛ ⎞⎛ ⎞⋅⎜ ⎟⎜ ⎟⎝ ⎠⎝ ⎠→
2
1
1
L vΔp =λ ρd 2
Δp =0,402MPa
⋅ ⋅
1 Die Bohrspülung (by Gerd‐Ulrich Lotzwick)
Figure 24 – friction coefficient for OSTWALD fluids1
24
Pressure losses inside the drilling collar:
The pressure losses in the inside of the drilling collar are calculated by the same way as shown above.
2Δp =1,201MPa
Pressure losses between bore hole and drilling rod:
When calculating fluid flow in an annular section, it is important to use the hydraulic diameter in the formula for the pressure loss and Reynolds number.
hydr o i
M
D =D - D
Re =1910
As the determined Reynolds number is smaller than the critical Reynolds number (3600), we can assume laminar flow.
M
3
Reλ= λ=0,03564
Δp =0,043MPa
Pressure losses between tubing and drilling rod:
After checking the Reynolds number, the calculation can be done either as shown above or like the first one but in both cases with the hydraulic diameter.
4Δp =0,157MPa
Pressure losses in the drilling bit:
To effectively support the drilling process the discharge speed of the drilling fluid from the nozzles should not be below 105m/s. Therefore it is possible to calculate the maximum nozzles cross section area by the law of continuity.
2Σnozzles discharge ΣnozzlesV=A v A =3,13cm⋅ →
To ensure enough fluid speed, a jet nozzle with a diameter of 7/16” (11,11mm) and a flow coefficient α of 0,95 would be suitable.
dischargeΣnozzles
V mv 112,8A s
= =
Now it is possible to calculate the pressure losses in the drilling bit.
discharge5 52
v ρΔp = Δp =8,814MPa
2 α⋅
→⋅
25
Pressure losses in the manifold system:
The manifold system can be separated in to four main groups. In case of this example the type is given and it is possible to read the pressure loss at the earlier determine flow rate from the diagram below (fig.201).
6Δp =0,25MPa
The diagram is only suitable for a fluid density of 1000kg/m³ (water). So if this diagram is used to determine the pressure loss in the manifold system with the fluid in the example it has to be corrected by the density factor.
drilling fluidF F
water
6 F 6
ρρ = ρ =1, 25
ρ
Δp =Δp×ρ Δp =0,31MPa
→
→
Total pressure losses:
To select a pump, now all the pressure losses are summed up and 10% is added to ensure sufficient head.
6
ii=1
(Δp )+10%=12,02MPa∑
So for this application a pump with a flow rate of about 120m³/h and a TDH of 1202m is needed. By entering this information into the previous mentioned pump selection software we get a number of matching pumps. To finally decide which pump fits the application best some other factors like acquisition costs, maintaining costs, energy consumption and electricity or fuel prices must be taken in account.
1 Die Bohrspülung (by Gerd‐Ulrich Lotzwick)
Figure 25 – pressure loss in manifold systems1
26
Figure 26 – Pump selection software, criteria definition1
Figure 27 – example of results provided by goulds pumps pums selection tool1
A possible selection can be saved as pdf‐file or plotted directly as shown in the example on the next page.
1 ITT – Goulds Pumps
27
27
28
In this case maybe a piston pump would be more suitable due to a higher efficiency at high pressures like a centrifugal pump. Problems with the selected pumps might be its operation far off it’s BEP and the fact that there is fairly no option to increase pressure or flow rat if necessary. Therefore centrifugal pumps are rarely found as mud pump. But as already mentioned they can be better used to suck the mud to a piston pump. In that application the needed TDH is much lower so this application matches the centrifugal pumps area of application, which is low TDH and high flow rate, much better.
Example 2 Another possible application on a drilling rig would be to charge a desander. A desander is, in most
cases, one or more hydrocyclones1. The mud enters the hydrocyclone tangentially into the upper cylindrical section. The mud is forced to move downwards into the conical segment. Due to centrifugal force the heavier solids are pressed against the outer wall. The inner phase of the mud can enter the inner cylindrical part at a certain point to flow upwards and discharge at the top. The solids leave the desander at its lower end. So solids down to about 80‐70µm are removed. To ensure proper operation it is important to guarantee the required velocity at the inlet of the desander. It mainly depends on the size of the desander. In this example the pump has to charge a tri‐flo model 1000‐2 desander unit with the properties shown below. It consists of 2x10“ hydrocyclones seperating solids down to 70µm.
model
Feed Rate
recomanded
operating
pressure
Length
Width
Height
Weight
Header
Diameter
Inside
10002 1000 gpm 25 psi 48“ 80“ 38“ 760 lbs 8
The manifold system recomanded by the manufacturer can be seen in fig.292. Over all, the desander is charged over a 6” (~150mm) pipe with a total length of 117” (~4m). Total hight difference is about 2m. The properties of the mud can be seen in the table below.
1 http://glwww.mst.dk 2 Tri‐flo desander 2x10“ operating manual
property value
Type OSTWALD‐fluid
density 1250kg/m³
KFactor 0,28Ns/m²
fluid index n 0,64
Figure 28 ‐ Hydrocyclone, working principle1
Figure 29 – recommended manifold system2
29
First calculating the fluid speed from the required flow rate and the given cross section area.
V m= =3,6A s
v
By calculating the Reynolds number it is possible to determine the friction factor from the diagram.
2-n n⋅ ⋅m m
v D ρRe = Re =14708K 1+3n88 4n
λ=0,015
n⎛ ⎞⎛ ⎞⋅⎜ ⎟⎜ ⎟⎝ ⎠⎝ ⎠
→
Now the pressure losses in the suction system can be calculated.
2
Δp =0,172MPa (=25psi)
Δp = g h=0,25MPa
1
1
L vΔp =λ ρd 2
Δp =0,034MPa
⋅ ⋅
Also the required pressure at the desander inlet and the hight difference has to be taken in account.
2 required pressure at hydracyclone inlet
3 ρ ⋅ ⋅
3
The sum equals the required TDH. In this case 20% for safety reasons are added.
ii=1
(Δp )+20%=0,842MPa∑
So we are looking for a pump with a TDH of 84m and a flow rate of 230m³/h. A possible pump would be a Goulds Pumps – model 3700 with a 4x6‐9 impeller operated at 3560RPM. The pumps caracteristic curve is shown in the diagram below.
30
31
9) Problems at centrifugal Pumps
A major problem at centrifugal pumps is, like at all fast moving parts in a fluid, cavitation. Other difficulties obtain solid handling, abrasives and corrosives as well as leakage. Most errors during pump operation can be avoided by selecting a quality pump designed for the application and adequate maintenance.
Cavitation Cavitation occurs when the static pressure in a fluid is lower than the fluids vapour pressure, mostly caused by high velocities. Due to Bernoulli’s law, static pressure decreases when velocity is increasing. If this happens, the fluid locally starts boiling and forms gas bubbles which need more space than the fluid would take. In a centrifugal pumps’s impeller, the bubbles are moving to an area of decreasing pressure. If the pressure now exceeds the vapour pressure, the gas condensates at the bubble’s inner surface and so collapse rapidly. This implosion of gas bubbles causes high, temporarily pressure fluctuations of up to a few 1000bar. As the fluid flows from higher to lower pressure, this flow causes a jet of the surrounding fluid, which may hit the surface. These high energy micro‐jets cause high compressive stress weakening the material. Finally, crater‐shaped deformations and holes known as cavitation pitting (fig.231) occur. Other reasons for cavitation can be a rise of fluid temperature, a low pressure at the suction side or an increase of delivery height. Cavitations in centrifugal pumps mainly occur at the impeller leading edges (fig.24) but also at the impeller vane, wear rings and thrust balance holes. To avoid cavitation, it is important to deliver sufficient NPSH and to keep fluid temperature low. High fluid temperatures can occur if the pump is on to keep the pressure up but no fluid is taken out
Figure 31 – regions of impeller cavitation2
The harm of cavitation to the impeller and other parts of the pump is significant.
1 www.motorlexikon.de (modified) 2 www.cheresources.com (modified)
Figure 30 – bubble collapse1
32
Figure 32 – typical impeller wear due to cavitation1
Solids and slurry handling (abrasive medias) When expecting solids in the fluid or dealing with slurries, it is important to select a pump that is designed for this application. On the other hand side, slurry pumps are much more expensive than a standard water pump, so the decision is not that easy. As there is a very wide range of slurries it is useful to divide them into three categories, light, medium and heavy slurries as shown in the table below.
property light slurry medium slurry heavy slurry
particle size <200µm 0,2mm – 5mm >5mm
settling / non settling non settling settling & non settling settling
specific gravity <1,05 1,05 – 1,15 >1,15
amount of solids <5% 5% ‐ 20% 20%
To provide a pump that can be used with slurries, special design features must be made. Slurry pumps can be equipped with e.g. thicker wear sections, larger impellers, special material and semi‐volute or concentric casing. All these features extend pump lifespan but also cause disadvantages like higher initial costs, higher weight or less efficiency. Slurry pumps can be separated into two main categories, rubber lined and hard metal pumps. At rubber lined pumps, the inner surface is covered by a layer of rubber, to absorb solid’s impact energy. Rubber lined pumps have a limited application range. This type of wear prevention is only suitable for light at least for medium slurries at low head applications. Also the fluid temperature should not exceed 150°C. Rubber lined pumps are not applicable for hydrocarbon based slurries. On the other hand side, hard metal pumps are suitable for high power applications used at even heavy slurries. Hard metal slurry pumps can also handle sharp, jagged solids even at fluid temperatures above 150°C. Standard hard metal slurry pumps can be designed of hardened steel but for high corrosive fluids high alloyed steels are used. When selecting a hard metal pump it is important that the pump material is harder than the solid particles. Cartable ceramics provide excellent resist to erosion but limit impeller tip velocity. The lifespan of a pump can be increased by selecting the correct materials of construction. Another important factor when handling slurries is speed. By decreasing the pumps RPM also the fluid speed is decreasing and therefore the solid’s speed is decreasing too. This leads to lower impact energy and less wear. Experiments by pump manufacturers have shown that a slurry pump’s wear rate is proportional to speed raised by the
1 www.korros.de
33
power of 2,5. Therefore, by decreasing the speed of a slurry pump by half, this will lead to approximately 6 times lifespan. For this reason most slurry pumps are operated at slowest speed possible equipped with impeller large in diameter to increase pump lifespan.
corrosion Corrosion is breaking down of essential properties in a material due to chemical or electrochemical reactions with its surroundings. As there is a wide range of pump applications within the chemical industry, including the petroleum industry, handling oil and gas up to high aggressive acids it is important to provide pumps that can be operated under these difficult conditions. There are several types of corrosion and many factors it depends on, like fluid temperature, contained elements and pH‐value. Most common and dangerous corrosion in pumps is the so called uniform corrosion. This is the overall attack of a corrosive liquid on a metal. The chemical reactions between fluid and metal surface lead to uniform metal loss on the moistened surface, known as corrosive wear. To minimize corrosive wear it is important to select a resistant pump material.
10) Comparison centrifugal pumps vs. Piston pumps
Centrifugal and piston pumps base on two different physical principles to cause flow. While a centrifugal pump accelerates the fluid along impeller vanes, a piston pump causes flow by the principle of positive displacement. The pressure in a piston pump is directly increased by fluid displacement, due to a force applied on an enclosed fluid volume. At the first step, only the inlet check valve is open and the back moving piston sucks fluid from the suction side. After a half rotation of the cam, the piston reaches the back dead centre. Now the piston starts moving forward and applies a force on the fluid. Therefore, the inlet check valve closes and the outlet check valve opens. The fluid is pressed into the piping at the discharge side. After the piston reaches the forward dead centre, fluid is sucked in again (fig.271). Obviously, a piston pump causes a pulsating flow, what is the first major difference. To reduce these pulsations, piston pumps are mainly designed as duplex, triplex or multiplex pumps. Most applications require an additional pulsation damper to reduce pulsations in the piping system. General centrifugal pumps are unstable at low flow rates but are a good choice at medium up to high flow rates. Piston pumps could be manufactured for similar flow rates but would get extraordinary big and too expensive for most applications. Centrifugal pumps are most suitable for low to medium pressure application while piston pumps are generally used in high pressure service. Multistage centrifugal pumps can be designed for pressured up to 400bar but are most efficient at high flow rates. Piston pumps on the other hand are generally a better choice for applications exceeding 200bar at low to medium flow rates. A piston pump is continuously increasing the pressure, while working against an enclosed fluid volume. Therefore, a relief valve is needed to prevent pump and piping system of overpressure. Centrifugal pumps cannot increase pressure upon the pumps typical shut‐off pressure on the pump characteristic curve. The shut‐off pressure is always lower than the pump’s design pressure and in a well designed application also lower than the piping systems maximum pressure. So when using a centrifugal pump, no relief valve is needed. An exception is to prevent the pump of damage due to temperature rise at low flow rates or shut down the pump and ensure a minimum flow to keep it stable. As a centrifugal pump operates on a various‐flow, various‐head curve, the flow rate increases if the discharge pressure is reduced. A piston pump always delivers a constant flow rate at a given speed, independent of discharge pressure.
1 www.lcresources.com (modified)
Figure 33 – piston pump1
34
Generally, centrifugal pumps, apart from special designs of some manufactures, are not self priming. So most applications require an external priming source. In application where both, a centrifugal pump as well as a piston pump, may be suitable another factor is required space and costs. A centrifugal pump is in general cheaper in acquisition and maintenance and requires less space than a comparable piston pump. On the other hand side, a piston pump requires less power. Of course this is only a general guideline. A pump operated outside of its optimum operating parameters can turn this around by causing e.g. higher maintenance costs. Therefore a pump should be carefully selected to avoid extra costs. So it is important to know that centrifugal pumps are suitable for handling clear, non abrasive fluids up to abrasive fluids with a high amount of solids, but do not work well with high viscous fluids because efficiency would drop dramatically. There would also appear problems when handling fluids combined with gasses due to the required close clearances. Piston pumps also work well for clean, clear non abrasive fluids up to abrasive slurries. Due to the relatively low fluid velocities, piston pumps are unsusceptible to erosions and wear.
Centrifugal pump Piston pump
optimum flow and pressure application
medium/high capacitylow/medium pressure
low/medium capacity medium/high pressure
maximum flow rate 50000m³/h + 3000m³/h +
low flow capability no yes
maximum pressure 400bar+ 7000bar+
requires relief valve no yes
smooth or pulsating flow smooth pulsating
self priming no yes
variable or constant flow variable constant
space conditions requires less space requires more space
fluid handling Suitable for a wide range including clean, clear, non‐abrasive fluids to fluids with abrasive, high‐solid content.
Suitable for clean, clear, non‐abrasive fluids. Specially‐fitted pumps suitable for abrasive‐slurry service.
fluid viscosity Not suitable for high viscosity fluids
Suitable for high viscosity fluids
gases Lower tolerance for entrained gases
Higher tolerance for entrained gases
costs lower initiallower maintenance higher power
higher initial higher maintenance lower power
35
11) Standards1
There are different organisations dealing with standardisation. Also some standards in pump design are available. Standards of design and dimensional specifications are necessary to bring unity to centrifugal pumps. Standards are provided by organizations like
• ISO ‐ International Standards Organizations
• API ‐ American International Institute
• ANSI ‐ American National Standards Institute
• DIN ‐ Deutsches Institut für Normung
• NPFA ‐ National Fire Protection Agency
• BSi ‐ British Standards institute Some commonly used centrifugal pumps standards
ANSI/API 610‐1995 ‐ Centrifugal Pumps for General Refinery Service ‐ Covers the minimum requirements for
centrifugal pumps, including pumps running in reverse as hydraulic power recovery turbines, for use in petroleum, heavy duty chemicals, and gas industry services. The pump types covered by this standard can be broadly classified as overhung, between bearings, and vertically suspended. DIN EN ISO 5199 ‐ Technical specifications for centrifugal pumps ASME B73.1‐2001 ‐ Specification for Horizontal End Suction Centrifugal Pumps for Chemical Process ‐ This standard covers centrifugal pumps of horizontal, end suction single stage, centreline discharge design. This standard includes dimensional interchange ability requirements and certain design features to facilitate installation and maintenance. It is the intent of this standard that pumps of the same standard dimension designation from all sources of supply shall be interchangeable with respect to mounting dimensions, size and location of suction and discharge nozzles, input shafts, base plates, and foundation bolt holes ASME B73.2‐2003 ‐ Specifications for Vertical In‐Line Centrifugal Pumps for Chemical Process BS 5257:1975 ‐ Specification for horizontal end‐suction centrifugal pumps (16 bar) ‐ Principal dimensions and nominal duty point. Dimensions for seal cavities and base plate installations.
12) Conclusion
Due to the wide range of applications and millions of sold pumps, nowadays centrifugal pumps are technically mature machines. Reasons for high efficiencies are a lot of experience as well as modern finite element optimisation. These flow optimisation procedures are standard engineering methods and lead to well constructed casings and impellers. This leads to many different special designs, constructed for a specific range of applications. Equipped with well selected anti wear systems and materials in combination with reasonable maintenance, a long lifespan can be met.
1 www.engineeringtoolbox.com
36
13) References
Internet sources:
ITT – Goulds Pumps http://www.gouldspumps.com
Light my Pump http://www.lightmypump.com/pump_glossary.htm
Mc Nally Institute http://www.mcnallyinstitute.com
Yokota Manufacturing Co., Ltd. http://www.aquadevice.com
The engineering tool box http://www.engineeringtoolbox.com
Literature:
Radial‐ und Axialpumpen (by A.J. Stephanoff)
Fundamentals and Applications of Centrifugal pumps (by Alfred Benaroya)
Die Bohrspülung (by Gerd‐Ulrich Lotzwick)
Reinhütte Pumpen – Centrifugal pumps, technical design (by Stephan Näckel)
Lawrence Pumps – Run Times, sept.04, jan.05 & oct.05 issue (by Dale B. Andrews)
World Pumps, sept.07 issue (by Joseph R. Askew)
Pump User’s Handbook (by Heinz P. Bloch, Allan R. Budris)
Figure sources are mentioned at the end of each page
Contact:
Christian Allerstorfer Roseggerstraße 10/6 8700 Leoben Supervised by Univ.‐Prof. Dipl.‐Ing. Dr.mont. Kessler Franz Dep. of design and conveying technology MU Leoben