159
CHAPTER
4
PumPs and PumP stations
4.1 INTRODUCTION
The main objective of liquid pipeline operations is to transport liquid petroleum prod-ucts from the producers to the customers. In order to achieve this objective, energy is added to the products to increase the pressure at the pump stations for offsetting the pressure loss in the pipeline. In addition, measurements of pressure and flow are required for facility control and custody transfer. There are other tasks required to oper-ate pipeline systems. This chapter discusses such key subjects as pump selection and sizing, pump operating points, pump station design, and station control.
A pump transforms energy to increase pressure of a liquid and is used extensively to transport liquid through a pipeline system. The pressure of a liquid has to be in-creased either to overcome frictional losses or to raise the liquid from one elevation to a higher elevation. As the flow rate increases, more pumps are required to produce the required pressure along the pipeline (Figure 4-1).
Depending on the method of adding energy to the liquid, pumps are classified into two types; centrifugal pumps and positive displacement (PD) pumps. Centrifugal pumps add kinetic energy to the liquid by increasing the liquid flow velocity, while PD pumps add energy periodically to the liquid by the direct application of a force to movable volumes of liquid. The two types of pumps can be compared in general terms, as listed in Table 4-1.
As shown in this table, centrifugal pumps are most suitable for transmission pipe-lines transporting most petroleum products. Therefore, they are extensively used in liquid transmission pipelines and thus this book only discusses the design and opera-tion of centrifugal pumps and pump staopera-tions.
Suction system Discharge system Pump - Pressure - Temperature - Specific gravity - Viscosity - Pressure - Flow
4.2 CENTRIFUGAL PUMPS
Centrifugal pumps have prevalent application in liquid hydrocarbon pipeline trans-mission systems as they are capable of handling variable heads and flow rates [1]. These pumps can handle multiple products over a wide range of viscosities and other properties [2].
Figure 4-2 shows the cross-sectional and axial views of a centrifugal pump with an end suction impeller. The main components of a centrifugal pump include an impel-ler, casing, housing or frame, and shaft and stuffing box/mechanical seal. A centrifugal pump uses the centrifugal action through the impeller within the pump to transfer energy. As the pump shaft is rotated by a pump driver, the impeller rotates inside the pump casing. Liquid flows from the suction piping into the impeller through its eye, and the rotating impeller imparts kinetic energy to the liquid. As the liquid slows down while passing through the volute, the liquid kinetic energy is converted into potential energy or pressure to conserve the total energy.
Figure 4-2. Centrifugal pump with impeller
TAbLE 4-1. Centrifugal pumps vs. positive displacement pumps
Centrifugal Pump PD Pump
Operating flow range Flexible Limited
Operating pressure range Low–medium High
Control of pressure and flow Flexible Limited
Pumping efficiency Low–medium High
Viscosity range Low–medium High
Capital cost Low–medium High
Maintenance requirement Low High
Pumps and Pump stations n 161
4.3 CENTRIFUGAL PUMP TYPES
Mainline centrifugal pumps are usually designed to ANSI/API 610 or ISO 13709 (Identical Standards) – Centrifugal pumps for petroleum, petrochemical and natural gas industries.
Pump types typically used in liquid hydrocarbon process, refining and pipeline transportation include:
4.3.1 End Suction Single Stage Pumps
These pumps have limited application in pipeline systems but are mostly used in proc-ess industries (Figure 4-3).
4.3.2 Vertical In-Line Single Stage Pumps
These pumps are typically used for product transfer within terminals or as booster pumps for smaller mainline pumps at initiating stations. API 610/ISO 13709 versions are used extensively in petrochemical and refinery service (Figure 4-4).
4.3.3 Horizontal Axially Split between-bearing Single-Stage Pumps
These pumps are typically used as mainline pumps. These high volume pumps are often piped in series configuration to provide the high pressures required for pipeline transmission lines. These double suction, double volute pumps provide high efficiency over a large range of flow. They are inherently balanced with minimal axial thrust issues at flows well off the best efficiency point due to the double volute design (Figure 4-5).
4.3.4 Horizontal Axially Split between-bearing Multi-Stage Pumps
These pumps have application where higher pressures and lower volume capacity than the single state pumps are required. This pump design, along with the single stage Figure 4-3. API 610 end suction pump — Courtesy of Flowserve Corporation, all rights
version, allows for ease of maintenance as the pump rotating element can be serviced or removed for impeller modifications without disturbing the station pressure piping. These pumps are robust in design and provide long service life. API 610/ISO 13709 is the guiding standard for design, manufacturing and performance testing requirements of these pumps. API 610/ISO 13709 designates these pumps as BB1 — axially split single stage bearing pumps and BB3 — axially split multistage between-bearing pumps (Figure 4-6).
4.3.5 Double–Case (Can) Vertically Suspended Volute Pumps
These pumps are used where there is limited suction pressure available such as in tankage terminals and where higher viscosity product is transported. Tank farm
de-Figure 4-5. Horizontal axially split BB single stage pump. Courtesy of Flowserve Corpora-tion, all rights reserve d.
Pumps and Pump stations n 163
signs usually provide manifolds located near the tanks and incorporate booster pumps with low NPSH requirements. These pumps are typically can-type vertical multi-stage centrifugal pumps that accommodate the low suction head available from the tankage. These booster pumps are designed to provide sufficient pressure to overcome frictional losses of valves, piping, and fittings throughout the station and to meet the NPSH requirements of the mainline pumps (Figure 4-7).
Impeller hydraulics can be optimized for individual system requirements with these pumps. Head rise from design point to shutoff can be as low as 15% over design Figure 4-6. Multi-stage horizontal axially split pump. Courtesy of Flowserve Corporation, all
rights reserve d.
head. This can result in considerable energy savings particularly under partial loads or off-peak operating conditions.
The can length will determine the NPSH requirements for the pump and is specific to the particular application. API 610/ISO 13709 designates this type of pump as VS-7 — Double casing volute vertically suspended pump.
4.4 PUMP SELECTION AND SIZING
The pump performance is the basis of pump selection. To select the pump, it is most critical to determine the operating range of the pump and the system curve in the pipe-line system. The actual procedure to be followed is:
Determine the required flow range of pump: It is important to define the ·
maximum, normal and minimum flow requirements. The normal flow is the flow at which the pump or pumps will usually operate. The rated flow is likely to be the maximum flow under current or near term conditions, whereas the minimum flow limit should be established to accommodate temperature rise and low efficiency or to install recirculation facility or multiple pumps. It is also important to clarify the number of days per year of service at which the pump will operate at maximum, normal and mini-mum flow rates.
Determine the system curves for the maximum, normal, and minimum flow ·
rates so that the operating envelope of one or more pumps is established. It is preferred to express the system curves in terms of head because the pump curve is expressed in head.
Select pump type, size, and arrangements, which are interrelated. For example, ·
two small pumps are arranged in parallel for wide range of flows between the minimum and maximum flows.
Prepare pump data sheets showing pump performance requirements and ser-·
vice conditions.
Solicit for bids and make final selection ·
4.4.1 Pump Performance
Normally, centrifugal pumps are characterized by efficient performance over a wide range of pressures and flow rates. Their size is relatively small. They cost less than other types of pumps and operate reliably. In addition, centrifugal pumps are capable of pumping high throughput and various products with different liquid densities and vis-cosities. They can handle products with viscosities up to 350 cSt depending on pump size and speed before efficiency begins to fall off significantly.
Pump characteristics are represented graphically to describe pump performance. The pump performance curves are normally provided by the manufacturer. The pump manufacturer tests the performance with water and only guarantees the performance at the rated point. Therefore, the pump performance curve, including the shut-off pres-sure, can vary and this variation must be specified or tested.
Pump performance curves show graphically the relationship of flow rate with head, efficiency, net positive suction head (NPSH) required, and power requirement for several impeller diameters. If the driver connected to the pump can vary its speed, these pump curves are produced for different speeds. The information is used for pump selection.
Pumps and Pump stations n 165
4.4.1.1 Pump Performance Curves
Pump manufacturers supply performance curves for their pumps with information on pump performance over a range of flows. The following diagram (Figure 4-8) dem-onstrates typical pump curves depicting pump head delivered by the impeller diam-eter chosen over the recommended flow rate for that pump. The Best Efficiency Point (BEP) is shown and usually forms the basis of the design flow rate. These pump curves typically provide the following information:
Range of impeller diameters available ·
Head vs. flow rate ·
Pump efficiency vs. flow rate ·
Brake power to drive the pump vs. flow rate ·
Net Positive Suction Head Required (NPSHR) for the pump vs. flow rate ·
These curves are plotted for the rated speed and different impeller diameters. For variable speed drivers, the curves are shown at various pump speeds.
Some pumps are equipped with double suction impellers with the impeller eyes located on both sides. Pump shut-off is the head developed at zero flow, while hydrau-lic runout is the pump capacity above which the pump should not be operated due to instability and other operational problems. This point is usually defined at 120% of the best efficiency point (BEP).
4.4.2 Service Conditions
There are many factors that must be considered in the selection and sizing of pumps for a liquid hydrocarbon pipeline system. In order to select the proper pump, the following parameters should be known:
Liquids to be pumped — clear liquids or liquids containing solids or vapor ·
Liquid specific gravity/density ·
Liquid vapor pressure at the pumping temperature ·
Viscosity— pump performance drops more rapidly above 110 cSt ·
System throughput — desired pump capacity and expected future changes to ·
the desired capacity are functions of pump speed and size
Pump head requirements — pump head depends on speed and impeller ·
diameter
Pressure conditions — suction and discharge pressures, future pressure condi-·
tion, and series/parallel operation conditions
Suction requirement — NPSHA must be greater than NPSHR of the pump ·
Pump unit efficiency — energy usage is dictated by pumping efficiency ·
Type of service — number of operating hours and criticality of the service ·
Preferred pump and driver type ·
Operability and maintainability ·
Equipment life cycle cost including the initial purchasing and installation costs, ·
operating cost, and maintenance cost
Specific site conditions and space limitations: topography and elevation above ·
sea level and NPSHA
Other considerations such as codes and regulations ·
A rule of thumb of sizing and selecting centrifugal pumps is to choose the physi-cally smallest pump that will satisfy the service requirements. Centrifugal pumps are sized on the following basis:
Impeller diameter: The pumping head is proportional to the square of the im-·
peller diameter, while the flow rate varies linearly with the diameter. Therefore, the larger the impeller diameter, the higher the head and the throughput. Nor-mally, pump vendors provide a range of impeller diameter suitable for a pump. Impeller diameters are determined based on required head at design point. The pump manufacturer will then trim the impeller to the required diameter. Impeller speed: The head and flow varies in a similar manner to the impeller as ·
described above. However, because of dynamic forces on the impeller, speed limits impeller size. The speed ranges from 1200 RPM up to 5500 RPM. Refer to Section 4.6.3 for the Affinity Laws.
Suction pressure: The NPSHR of a pump is the limiting factor that affects size, ·
speed, and capacity. This topic is discussed in Section 4.4.3.
Suction and discharge nozzle sizes: Suction nozzles are usually larger than dis-·
charge nozzles. The larger the nozzle size, the higher the flow capacity of the pump. Nozzles sizes are determined by the pump manufacturer.
Pipeline system hydraulic requirements will determine the selection of pumps. If there are batching requirements with products of varying density and viscosity, it may be necessary to provide multiple pumps operating in series or parallel, or to use some combination of series/parallel configuration. Variable speed pumps can sim-plify the selection process. The pump configuration should be selected to maintain the maximum pump efficiency over the range of flows expected for the pipeline system. Initial and future flow rates should be evaluated so that the selected pumps have suffi-cient flexibility to handle any anticipated flow and product properties. For comparison purposes, pump efficiencies should be evaluated for both series and parallel pump configurations.
Once the pump configurations with head requirements at various flow rates have been determined, it will be necessary to select actual pumps from manufacturers’ pub-lished pump performance data. Pump manufacturers publish pump performance maps that depict pump performance for a specific family of pumps over a large range of
Pumps and Pump stations n 167
flows and differential heads. Pump manufacturers produce charts to show the capaci-ties of different sizes of pumps. Figure 4-9 shows a typical performance map for cen-trifugal pumps suitable for pipeline service.
4.4.3 Net Positive Suction Head (NPSH)
Net positive suction head (NPSH) is the total absolute suction pressure at the impel-ler eye less the absolute vapor pressure of the liquid pumped. NPSH must be of a magnitude to avoid vapor formation in the liquid and hence cavitation. Another way to describe NPSH is that it represents the amount of head required to push the fluid into the pump to suppress cavitation [3]. Available and required NPSH are calculated as follows:
4.4.3.1 Net Positive Suction Head Required (NPSHR)
The NPSHR specified in the pump performance curves is another important perfor-mance parameter. All centrifugal pumps accelerate fluid and possess a corresponding internal friction loss. Therefore, every pump requires a certain amount of positive suc-tion pressure in order to avoid cavitasuc-tion. The pump NPSHR is determined by actual tests conducted by the pump manufacturer using procedures established by the Hy-draulic Institute.
Cavitation is a critical problem in controlling pumps. When a liquid enters into the pump impeller in such a manner that the local pressure drops below the liquid’s vapor pressure, the liquid transforms into the vapor phase and bubbles are formed in and around the impeller. If the local pressure recovers due to centrifugal action, the vapor bubbles collapse into the liquid phase. Since the bubbles occupy much larger volume than the liquid, collapsing bubbles release a huge amount of energy. This energy will hit the surrounding metal, causing physical damage to the impeller and casing. The other effects of cavitation include vibration and noise and dramatic reduction in pump performance.
4.4.3.2 Net Positive Suction Head Available (NPSHA)
Net Positive Suction Head Available (NPSHA) is calculated by the following expression: a VP st fs NPSHA h= -h + -h h Or 2.31 ( ) st a VP fs NPSHA h P P h G = + - -where
Pa = absolute pressure at the surface of the liquid supply level
PVP = vapor pressure of the liquid at the temperature being pumped
ha = Pa expressed in equivalent head
hVP= PVP expressed in equivalent head
hst = static elevation of the liquid supply above or below the pump inlet
centerline.
hfs = suction line losses including entrance losses and friction of the piping
G = specific gravity
Figure 4-10 shows that NPSHA must be greater than NPSHR for stable operation and that NPSHR increases and NPSHA decreases as the flow rate increases.
The NPSHR stated by the pump manufacturer is at a point where the pump is in full cavitation. Therefore, it is important to allow a margin between NPSHR and NPSHA. As a rule of thumb, the NPSHA should be at least 10% greater than the NPSH required by the pump [1].
If the requirement for stable operation cannot be satisfied, either NPSHR should be reduced, NPSHA increased, or both. The NPSHR reduction can be accomplished by using double suction impellers or by impeller design with a larger impeller eye area. Also, smaller
Pumps and Pump stations n 169 pumps can be installed in parallel or a larger suction pipe size can be used to reduce frictional pressure losses in suction piping. NPSHA can be increased by installing a booster pump in front of the pump or by reducing the friction pressure losses in suction piping.
4.4.4 Specific Speed
Specific speed (NS) is used to predict pump characteristics for the purpose of classifying
pump impellers according to type, proportions and performance. Specific speed is often used for comparison purposes on selection of pumps from different manufacturers [3].
Specific speed is expressed as:
3 4 s N Q N H = where
Ns = pump specific speed, dimensionless
N = pump speed, RPM
Q = capacity at best efficiency point, USGPM
H = total head per stage at the best efficiency point, feet
For double suction impellers, one half of the flow is used to calculate the specific speed. The specific speed (Ns) determines the general shape or class of the impeller. As
the specific speed increases, the ratio of the impeller outlet diameter, D2, to the inlet or
eye diameter, D1, decreases. This ratio becomes 1.0 for a true axial flow impeller.
Radial flow impellers develop head mainly through centrifugal force. Pumps of higher specific speeds develop head partially by centrifugal force and partially by axial force. A pump with a higher specific speed generates head more by axial forces and less by centrifugal forces. An axial flow or propeller pump with a specific speed of 10,000 or greater generates its head exclusively through axial forces.
Typical values for specific speed (Ns) for different designs in US units (gpm, ft)
radial flow — 500
· < Ns< 4000 — typical for centrifugal impeller pumps with
radial vanes — double and single suction. Francis vane impellers operate in the upper range
mixed flow — 2000
· < Ns< 8000 — more typical for mixed impeller single
suction pumps axial flow — 7000
· < Ns< 20,000 — typical for propellers and axial fans
To convert between US units (USgpm) and Metric units (m3/h)
Ns(US gpm, ft)= 0.861 Ns(m3/h, m)
The specific speed of an impeller can provide a wide variety of information about its performance:
Impellers with low specific speed are long and thin and are used for low-flow, high-head applications. Impellers with high specific speed are short and stubby and are used for high-flow, low-head applications (Figure 4-11).
Efficiency is determined by considering the losses through pump impeller fric-tion, ring leakage, and mechanical losses, as well as losses incurred by movement of the liquid within the pump, referred to as hydrodynamic losses. Specific speed affects pump efficiency. The lower the specific speed, the lower the efficiency. The reason is
that a higher percentage of energy is lost to overcome the impeller disk friction that is necessary to generate high heads (Figure 4-12).
Once an impeller is designed for a certain specific speed, it will produce a typical head capacity curve and efficiency curve shape. A low specific speed impeller has a flat curve with a wide efficiency range. A high specific speed impeller produces a steep curve with a narrow efficiency range.
The major use of the specific speed number is to help specify pumps to be as efficient as possible for the service intended. Maximum pump efficiency is obtained in the specific speed range of 2000 to 3000. Pumps for high head low capacity occupy the range 500 to 1000 while low head high capacity pumps may have a specific speed of 15,000 or larger.
4.4.5 Suction Specific Speed
Suction specific speed is an index that describes the characteristics of the suction side of the impeller. It is calculated at the pump’s best efficiency point and maximum impel-ler diameter. The equation for suction specific speed (designated Nss or S) is:
Figure 4-12. Pump efficiency vs. specific speed
Pumps and Pump stations n 171
( )
0.5 0.75 ( ) N Q Nss NPSHR = where N = rotating speed (rpm)Q = flow per impeller eye (m3/second)
NPSHR = Net Positive Suction Head Required (see Section 4.4.3)
For double suction impellers, Q is one half of total flow. Nss derived using SI units can be converted to US Customary Units by multiplying by a factor of 51.64 (Ref. API 610)
From the equation, we can see that the lower the NPSHR for a pump the higher the Nss. Nss values for many standard impellers typically range from 7000 to 9000, but some designs may have an Nss as high as 18,000 to 20,000.
It is important to consider that increasing the Nss of a pump has been shown to shift the onset of suction or discharge recirculation closer to the best efficiency point (BEP) flow of the pump. This effectively decreases the window of stable operation for the pump. Suction recirculation is a reversal of flow in the impeller eye that can lead to increased noise, surges, and cavitation-like damage to the impeller vane. Discharge re-circulation is a similar reversal of flow occurring at the discharge of the impeller vane. It is recommended that pumps should have an Nss of no more than 9000 for water and 11,000 for hydrocarbons [3].
4.4.6 Pump Performance Curve Characteristics
There are a number of pump head-capacity (H-Q) curve shapes that are shown in Figure 4-13 below. These characteristic curves are listed below and described thereafter:
Rising · Drooping · Steep · Flat ·
Stable ·
Unstable ·
Rising characteristic — the head rises continuously as the capacity is decreased to
shut-off.
Drooping characteristic — the head developed at shut-off is less than at some of the
other capacities.
Steep characteristic — the head developed at shut-off is significantly larger than that
developed at the design capacity.
Flat characteristic — the head developed at shut-off is approximately that developed
at the design capacity. The curve can be a slightly rising or drooping.
Stable curve — is a rising curve where only one capacity can be obtained at any one
head. A curve with a rising characteristic would be an example of a stable curve.
Unstable curve — is a drooping curve where more than one capacity can be
obtained at one head. A curve with a drooping char acteristic would be an example of an unstable curve.
Unstable curves where the maximum developed head is at some flow greater than zero are undesirable in applications where multiple pumps operate in parallel. In such applica-tions, zero flow head may be less than system head, making it impossible to bring a second pump on line. It is also possible for pumps in this configuration to deliver unequal flow with the discharge pressure from one pump determining the flow rate from another [1].
4.4.7 Centrifugal Pump Power and Efficiency
The ideal hydraulic power to drive a pump depends on the flow rate, the liquid density, and the differential head generated by the pump. This can be calculated as:
h 6 (3.6 10 ) q gh P = r ´ where Ph = hydraulic power (kW) q = flow capacity (m3/h) r = density of fluid (kg/m3)
g = acceleration due to gravity (9.81 m/s2)
h = differential head (m)
Power is more commonly expressed as kilowatts (kW) or horsepower (hp = kW ´ 0.746).
The shaft power is the power required transferred from the motor to the shaft of the pump, and it depends on the efficiency of the pump. Shaft power can be calculated as:
h s P P = h where Ps = shaft power (kW) h = pump efficiency
4.4.8 Performance Modifications For Varying Pipeline Applications
The performance of pipeline pumps often needs to be altered to accommodate varying liquid transmission conditions. These generally require that the pump be physically mod-ified to meet the new conditions. The following are considered by the pipeline industry:
Pumps and Pump stations n 173
Impeller change — In order to change the specific speed, impeller size may be changed to meet the new demand on performance. Pump manufacturers can usually offer sev-eral different impeller diameters and vanes that will fit the pump casing without any further internal modifications. The effect of changing the impeller characteristics is illustrated in Figures 4-14 and 4-15.
Restaging — Pumps with multi-staging capabilities can be restaged (up or down) to meet the change in pressure, head or flow requirements. For example, if an entire pressure range is not needed for a particular period of time, a number of impellers can be removed to meet the required conditions. Manufacturers provide de-staging kits to block off the unused pump impeller areas to maintain efficiency (Figure 4-16). Figure 4-14. Changing performance by impeller vane number
Figure 4-16. Changing performance by re-staging
Pumps and Pump stations n 175
Figure 4-18. Changing performance with volute chipping
Impeller underfiling and overfiling — This is undertaken to alter the performance of a pump. It involves modifying the flow area of the impeller by grinding metal off the impeller outlet vanes (Figure 4-17).
Impeller volute chipping — This is a technique that is used to alter the outlet flow area of the pump casing in order to modify performance (Figure 4-18).
Impeller volute inserts — This technique (Figure 4-19) involves inserting special removable volutes into the pump to allow for a wider performance range. It also allows for more accurate and close control over the performance.
4.4.9 Cavitation [6–10]
Cavitation is the rapid formation and collapse of vapor bubbles that form in a pump inlet whenever the local absolute pressure of the liquid falls below its vapor pres-sure. These bubbles collapse rapidly and violently when the local absolute pressure increases due to kinetic forces being imparted by the impeller (see Figure 4-20 below). Cavitation is the rapid formation and collapse of these vapor bubbles.
Figure 4-20. Cavitation bubble implosion, arrows indicate fluid pressure [16 modified]
Collapsing vapor bubbles cause noise, vibration, and erosion of material from the damaged impeller as shown below in Figure 4-21 [4].
Cavitation control is a very important consideration in any liquid system and thus any cavitation induced conditions must be avoided when operating centrifugal pumps. If a liquid is accelerated in such a manner that the local pressure falls below the liquid vapor pressure, the liquid will transform into the vapor phase, which results in the formation of bubbles. If the local pressure recovers, the vapor bubbles will transform themselves back into a liquid. There is a tremendous volume change during transformation, because collapsing bubbles release a large amount of energy. Because the bubbles are very small, the resulting impact loads on the surrounding metal can be significant. This can result in the creation of high noise levels and physical damage to the metal [6].
Some liquids (such as water) are more difficult to handle from a cavitation point of view. When the vapor pressure of a homogenous fluid such as water is reached, the entire fluid begins to change phase, resulting in the formation of a large number of damage-causing bubbles. For a non-homogeneous fluid such as a hydrocarbon, only the light ends (such as condensates) are affected and the impacts of cavitation are reduced (Figure 4-22).
In a centrifugal pump, the fluid is accelerated by the impeller. The area of lowest pressure in the pump suction system, as shown in Figure 4-23, is the eye of the impeller at cross section A-A. If the pressure falls below the vapor pressure of the liquid, vapor bubbles form. As the mixture of liquid and bubbles continue through the pump, the pressure increases and the bubbles return to the liquid state. Damage to the impeller
Pumps and Pump stations n 177
Figure 4-21. Pump impellers destroyed by cavitation [10,8]
Figure 4-23. Cavitation formation
Pumps and Pump stations n 179 occurs where the bubbles collapse as shown at cross section B-B. This location varies for different impellers and different suction conditions:
The effects of cavitation include: Noise and vibration
·
Pump damage (e.g., pitting of the impeller) ·
Fall off of pump performance and efficiency ·
Cavitation in centrifugal pumps can be recognized by a characteristic noise, which sounds just like it is trying to pump gravel. A typical break-off in the performance curve of a pump due to cavitation is shown in Figure 4-24.
Figure 4-25. Performance correction chart for viscous liquids — with permission of Hydrau-lics Institute
4.4.10 Viscous Hydrocarbon behavior in Pumps [11]
The performance of centrifugal pumps is affected by fluids with higher viscosities in-cluding fluids that behave in a non-Newtonian manner. An increase in power required a reduction in head generated, a loss in efficiency and in some cases a loss of capacity can be expected with the transmission of high viscosity fluids.
Performance correction factors, as depicted in Figure 4-25, can be applied to esti-mate the actual performance of a pump handling viscous fluids.
Figure 4-25 is to be used only within the scales shown and only for pumps of conventional design. The chart is applicable only for conventional centrifugal pumps operating close to their best efficiency point (BEP), having sufficient net positive suc-tion head available (NPSHA) and transferring Newtonian fluids (such as crude oil). For description on non-Newtonian fluids refer to Chapter 1.
The following equations are used for determining the viscous performance when the water performance of the pump is known:
Qvis= CQ´ QW Hvis= CH´ HW Evis= CE´ EW 3960 vis vis vis vis Q H sp gr BHP E ´ ´ = ´ where
Qvis= viscous capacity (USgpm)
Hvis= viscous head (ft)
Evis= viscous efficiency (%)
BHPvis= viscous brake horsepower
Pumps and Pump stations n 181 QW = water capacity (USgpm)
HW = water head (ft)
CQ = Capacity correction factor
CH = Head correction factor
CE = Efficiency correction factor
QW = Water Capacity at which maximum efficiency is obtained
Figure 4-26 illustrates graphically the consequences of viscosity increases on the head, capacity, and brake power requirements. As the liquid viscosity increases, the head generated by a centrifugal pump is reduced, the pumping efficiency drops, and as a result the brake power required for pumping increases. Note that the viscosity is greater than 500 SSU or about 100 cSt, the pump performance drops rapidly.
4.4.11 Temperature Rise
Fluid temperature rises across the pump for two reasons. First, there is an adiabatic temperature rise due to compression of the hydrocarbon fluid across the pump. As well, heat is generated due to frictional forces generated within the pump and by recircula-tion caused by leakage through clearances within the pump. This major component of the temperature rise is directly related to the efficiency of the pump at the operating flow rate. The amount of heat generated is the difference between the input energy to the pump and the delivery energy.
Temperature rise in a pump with a closed discharge valve can occur quickly if the pump generates high head. Power losses are equal to the shutoff input power, and all this power goes into heating the small quantity of fluid contained within the pump casing. Care must be taken to prevent extended shut-in head conditions during pump startup.
Since the heat is generated as the liquid passes through the pump, the pump dis-charge temperature is increased and is calculated using the fundamental concept that the mechanical energy lost in the pump due to mechanical efficiency is converted to heat energy. d s p 1 P T T C æ ö D - h = + ç ÷ r è h ø where Td = discharge temperature (oC) Ts = suction temperature (oC)
Cp = liquid heat capacity (kJ/kg oC)
h = pump mechanical efficiency r = liquid density (kg/m3)
Δ P = discharge and suction pressure difference (kPa)
When the pump is running normally, the temperature increase is small, in the order of a few degrees Celsius. If the pump discharge is shut off or the flow is too slow, energy is converted to heat and the heat cannot be carried away quickly. The liquid in the pump will heat and eventually vaporize. This can result in dramatic failures, particularly for large multi-stage pumps. Such a situation can be avoided by auto-matically shutting down the pump as the flow rate drops below the pump’s Minimum Continuous Stable Flow (see Section 4.4.12) or by providing a recirculation system (see Section 4.9.4).
4.4.12 Minimum Flow
Minimum Continuous Stable Flow (MCSF) is usually provided by the pump manu-facturer and is the lowest flow rate at which a pump can operate continuously without exceeding the vibration and/or noise limits specified. These limits are usually refer-enced to industry standards such as ANSI, API, ISO, and ASME or, in some cases, to the customer’s own pump specifications (Figure 4-27).
There are many ways by which a manufacturer determines its recommended MCSF for a specific pump. This can be based on actual test results, historical experi-ence or the specifics of the pump design. MCSF can be influexperi-enced by the properties of the pumped liquid as well [5].
4.5 PUMP SPECIFICATION AND PURCHASE
A pump purchase requisition must be prepared and should consist of at least two parts:
Completed API 610/ISO 13709 Pump Data Sheets ·
Clarifications/Supplements to API 610/ISO 13709 Standards, if required ·
4.5.1 Pump Data Sheets
Appendix B of API 610/ISO 13709 contains sample data sheets that should be used when purchasing an API 610/ISO 13709 pump.
It is important to complete all sections of the pump data sheets. There are many options in the API 610/ISO 13709 standards and the data sheet will clarify which Figure 4-27. Performance curve showing MCSF
Pumps and Pump stations n 183 o ptions are required and ensure that all vendors are quoting to the same requirements. Important aspects that the data sheets clarify are:
Service conditions (see Section 4.4.2) ·
Liquid properties (see Chapter 2) ·
Site conditions ·
Preferred coupling type ·
Materials of construction ·
Mechanical seal type ·
Mechanical seal flush piping ·
Bearing lubrication system requirements ·
Instrumentation requirements ·
Painting requirements ·
Electrical power available and hazardous area classification ·
Pump testing requirements ·
Material inspection requirements ·
Spare parts requirements ·
4.6 RETROFITTING CENTRIFUGAL PUMPS FOR CHANGING
SERVICE CONDITIONS
4.6.1 Reduced Pipeline Throughput
When pipeline throughput is reduced, the pumps in effect become technically over-sized and are therefore operating inefficiently. These pumps then become subject to problems that are associated with low flows, such as vibration or seal and bearing failures. There are a number of retrofit solutions:
Change the pump speed to adapt to the desired flow conditions in the pump, reduce the pump impeller diameter to better suit the lower flow rate, ·
install volute inserts; or, ·
combination of the above solutions. ·
4.6.2 Increased Pipeline Throughput
When pipeline throughput above current system design is required, it will be neces-sary to add intermediate pump stations to the pipeline system. By adding stations, the system hydraulic performance curves are modified to allow increased flow rates while staying within system Maximum Allowable Operating Pressures.
These increased flow rates through existing pump stations may change the pump operating conditions to the point that pumps operate inefficiently or overload the pump drivers. Depending on the actual flow rate increases required, there are a number of modifications that can be adopted to have the existing pumps perform well under the new operating conditions. Retrofit solutions include:
Trimming the pump impeller to allow a constant speed pump driver to remain ·
within its power limits.
Reducing the speed of a variable speed driver to limit the power requirements ·
Adding a parallel pump unit capable of delivering the same pump head as the ·
existing pump/pumps that operate in series. This approach then reduces the flow to the existing pump/pumps to allow pump performance to be maintained as originally designed.
The following section on affinity laws outlines the method of calculating impact on capacities, pressure rise and pump speeds by changing certain parameters.
4.6.3 Affinity Laws [4]
The flow rate and head generated by a centrifugal pump may be changed by varying either the pump speed or changing the impeller diameter. This results in a change to the impeller tip speed or velocity of its vanes, which causes a change in the velocity at which the liquid leaves the impeller. Usually, impellers can be cut down to 80% of their original diameter without lowering their efficiency significantly.
For centrifugal pumps with radial impellers, the relationships are approximated as follows:
For diameter change only:
2 3 2 2 2 2 1 2 1 2 1 1 1 1 , , D D D Q Q H H BHP BHP D D D æ ö æ ö æ ö = ç ÷ = ç ÷ = ç ÷ è ø è ø è ø
For speed change only:
= = = 2 3 2 2 2 2 1 2 1 2 1 1 1 1 , , N N N Q Q H H BHP BHP N N N ,
Pumps and Pump stations n 185 For diameter and speed change:
2 3 2 2 2 2 2 2 2 1 2 1 2 1 1 1 1 1 1 1 , , D N D N D N Q Q H H BHP BHP D N D N D N æ ö æ ö æ ö = ç ´ ÷ = ç ´ ÷ = ç ´ ÷ è ø è ø è ø , where
D = impeller diameter (in.) H = head (ft)
Q = capacity (USgpm) N = speed in RPM BHP = brake horsepower
1= original conditions subscript
2= new design conditions subscript
The affinity laws can be presented graphically as shown in Figure 4-28. The pump-ing capacity and efficiency increases as the impeller diameter and/or speed increases. Since the head and flow capacity are higher, the power needed for higher speed and/ or larger diameter is greater. Figure 4-30 illustrates the capacity changes with speed change on the left and impeller diameter change on the right hand side.
4.7 PIPELINE HYDRAULIC REQUIREMENTS
4.7.1 System Head Curves and Pump Operating Points
A system curve or system head curve for a pipeline demonstrates the head required at that location for the flow rate range. As demonstrated in Chapter 3, the pressure or head required to overcome frictional losses increases with the flow rate. The system curve depends on the following variables:
Flow rate ·
Liquid density and ·
Liquid viscosity ·
Pipe diameter, wall thickness, and · Pipe roughness · Elevation difference · Pressure or head Flow Rate Friction
As detailed in Chapter 3, the friction pressure loss can be calculated for a given liquid with its known product properties and pipe parameters. Since pump curves and elevation differences are given in head, the friction pressure loss is usually expressed in head. Using the example given in Section 3.1.3.2, a complete system curve can be built for the entire operating flow range. Assuming that the operating flow ranges from 250 m3/hr to 835 m3/hr, the friction pressure loss is calculated in the same way as described in the section and the system head required for each flow rate is determined from the relationship of Dh = Dp/(r ´ g), where Dh is the head, Dp the friction pressure loss, r the fluid density, and g the gravitational constant (Figure 4-29).
In general, r the friction pressure loss approximately varies as the square of the flow in a pipeline. This approximation works well if the operating flow range is narrow or the friction factor does not vary with the flow rate significantly. Since the friction factor change is sensitive to flow rate change for high viscosity liquid, the friction fac-tor and friction head has to be calculated for each flow rate. Once the friction heads are calculated for various flow rates, the total head can be determined by adding the static head to the friction head. Assuming that the head increase due to elevation gain is 100 m, the total head requirements are summarized in Table 4-2. This table is used to generate the system head curve over the operating flow range.
In addition, the frictional components of the system curve can vary, sometimes over a long time period, with the following changes:
Density due to batch movement, ·
Temperature, ·
Wax buildup on the pipe wall, ·
Side stream delivery or injection, ·
Throttling of a control valve. ·
TAbLE 4-2. system head vs. flow rate
Flow Rate (m3/hr) Friction Head (m) Static Head (m) Total Head (m)
250 110 100 210 625 524 100 624 835 865 100 865 Pressure or HEad Flow Rate High Friction Low Friction Friction Friction
Pumps and Pump stations n 187
These changes could move the system curve and consequently pump operating point. Therefore, the range of these changes in operating point and their consequences in the power requirement, capacity and NPSH has to be considered in designing and selecting pumps.
A system curve is a graphic representation plotted on an x–y graph, where x-axis represents the flow rate and y-axis the pressure or head caused by the frictional pressure drop along the pipeline. Figure 4-31 shows a typical system curve for a flat pipeline system in which a single product is transported.
Figure 4-30 below shows system curves for a level pipeline system in which two products are transported, assuming that the two products have different densities and/ or viscosities. The higher specific gravity and viscosity product requires greater pres-sures compared to the liquid with lower gravity and viscosity, generating different sys-tem curves. Figure 4-30 shows two examples of syssys-tem curves; one for heavier product
Pressure or Head Flow Rate System Curve Friction Head Static Head Elevation Difference 0
Figure 4-31. System curve with elevation difference
Pressure or Head Flow Rate Q0 Pump Curve System Curve Operating Point PO
causing higher frictional pressure drop and the other for lighter product causing lower frictional pressure drop. It is assumed that the pipe size remains constant and the eleva-tion profile is flat. If the pipe size changes, so does the slope of the system curve where the pipe size changes. If the elevation changes at two different locations, the system curve simply moves up or down as shown in Figure 4-31.
If the next pump station or delivery point is higher in elevation, the system curve is shifted upward as shown in Figure 4-33.
The pump operating point is the point where the pump curve meets the system curve, as shown in Figure 4-32 below.
4.7.2 Hydraulic Performance in batched Pipeline Systems with
Constant Speed Pumps
If more than one product is batched in a pipeline, the operating points will change depending on their positions with respect to pipeline system and pump curves. As an example, refer to Figure 4-33, showing the system curves for the heavy and light oil. The operating points of these two liquids are located at H for the heavy oil and L for the light oil for the pump. Even though these two products have different densities, the pump curve in head is the same. With all heavy oil in the pipeline, the operating point is located at H, the flow rate being QH and the head
being HH. When batching products, the positions of the batched products shift
with respect to a pump and pipeline, filling the pipeline with the heavy oil and the rest with the light oil. Therefore, the flow rate will be at some point between QH
and QL, and the operating point is determined with a new system curve. In this
example, as the light oil moves into the pump and the pipeline, the operating point moves to L, eventually reaching equilibrium at an operating point where the flow rate is located at QL and the head at HL.
Figure 4-34 below shows the pump and system curves plotted in pressure instead of head. Since the pressure changes with density, two pump pressure curves are dis-played with four operating points. If all heavy oil is in the pipeline and passes through
Head Flow Rate Heavy Oil Light Oil H L HH HL Pump Curve QH QL
Pumps and Pump stations n 189
the pump, the operating point is located at HH in the figure. As the light oil enters the pump, while the heavy oil is still flowing in the pipeline, the operating point slowly moves from HH to HL, but with reduced flow rate. In reality, this movement does not occur abruptly because the batch interface enters first and the density gradually changes before the light oil completely replaces the interface. The flow reduction is caused by the low density of the light oil passing through the pump and thus reducing the differential pressure. As the light oil moves into the pipeline, the operating point gradually moves to the new operating point, LL, because the light oil requires lower system curve. As the light oil fills the pipeline, the flow rate increases but the pressure decreases. The new point represents the operating point with the light oil in both the pipeline and the pump.
In the same process as described in the previous paragraph, as the heavy oil moves into the pump and the pipeline is filled with light oil, the operating point moves to LH, and eventually the operating point moves back to HH as the heavy oil fills both the pump and the pipeline. In batch operations, the flow rates vary between QHL and
QLH, so does the pressure assuming the pump is not throttled. Note that the positions
of these four operating points may be altered depending on the viscosity of the heavy oil because pump performance curves are changed with higher viscosity liquids (refer to Section 4.4.10).
4.7.3 Hydraulic Performance in batched Pipeline Systems with
V ariable Speed Pumps
With the introduction of economically available variable speed drivers for centrifugal pumps (see later descriptions in this chapter); it is possible to provide more capacity to batched product pipelines very efficiently. By increasing pump speed in the example shown above (Figure 4-34) as light oil enters the pump with the pipeline full of heavy oil, it is possible to maintain pressures and thus flow rates at QHH with more rapidly
increasing flow rates to QLH than in the previous example. In this case, the design point
for the pumps would be at LH with heavy oil density and viscosity determining the power requirements of the variable speed pump.
Pressure Flow Rate Heavy oil Light oil HH LH PHH Pump curve for heavy oil
HL LL PHL PLH PLL Pump curve for light oil
QHLQHH QLL QLH
4.7.4 Pump Configurations
4.7.4.1 Parallel Operation
If a station has more than one pump unit, it is operated in one of two modes: parallel operation and series operation. Multiple pumps in a pump station can be arranged in parallel, series, or combination of parallel and series arrangement. Some combination of these arrangements is needed to increase both flow and pressure.
The primary purpose of operating units in parallel is to allow a wider range of flow than would be possible with a single unit for systems with widely vary-ing flows. Generally, pumps are arranged to operate in parallel to increase sys-tem flow rates by maintaining discharge head through a much greater pipeline throughput. The following alternatives should be considered in selecting pump configurations:
One operational pump with one spare unit in parallel; ·
Two operational pumps in parallel with one parallel pump as spare; ·
Two pumps in parallel without a 100% spare, if less capacity can be tolerated ·
at times.
When pump units are arranged in parallel, loads are shared by more than one unit. The following load sharing strategies are available:
Base loading: one or more units may be operated at a constant load while other ·
units are operated to handle additional flow rate requirements;
Optimum load sharing: set points for each unit are determined based on know-·
ing the individual operating curves and allocating load to optimize pump e fficiency and minimize overall energy consumption;
Equal load sharing: each unit shares load equally. This strategy is used for ·
identical units which are arranged in parallel.
There are several areas of concern when operating centrifugal pumps in parallel. If adequate flow control is not provided, hydraulic imbalance could occur resulting in high flow rates in one of the parallel pump units and low flow rates in the other unit. Possible issues include:
Station Block valve Control valve Station Suction valve Station Discharge valve Isolation valve Isolation valve V-23 V-24 V-23 V-24 Pump 1 Pump 2 Check valve Filter Bypass check valve Bypass check valve
Pumps and Pump stations n 191
High flow rate induced cavitation; ·
Pump driver overloading; ·
Prolonged operation of one of the pumps at a flow below its minimum accept-·
able continuous flow rate.
Selection of appropriate control systems is essential with pumps operating in parallel configuration.
Figure 4-35 shows a parallel arrangement of two pumps, where more than one unit can be operated at the same time. Figure 4-36 demonstrates the operating points of one pump and two identical pumps when they are arranged in parallel. When two or more units operate in parallel, all units have common suction and discharge pressures. Figure 4-36. Operating points for parallel operation
Station Block valve Control valve Station Suction valve Station discharge valve Isolation valve Isolation valve Suction
valve Dischargevalve
Pump 1 Pump 2
Check valve Filter
Bypass
check valve check valveBypass Suction
valve Dischargevalve
4.7.4.2 Series Operation
The main reason for operating units in series is to increase the pumping head from what would be possible with a single unit. Pumps operating in series add head and capacity to the system output. In areas of large pipeline elevation rise, if two pumps are arranged in series and one shuts down, the remaining pump alone may not be able to provide the necessary head for the static lift necessary to maintain pipeline flow. In this case, appropriate station design would require that a spare pump unit be installed to be activated in the event of a loss of a pump unit operation.
Figure 4-37 shows a series arrangement of two pumps, and Figure 4-38 demonstrates the operating points of one pump and two identical pumps when they are arranged in series. In series operation, the flow through all of the units is equal and the discharge of one pump feeds the suction of the next unit.
4.8 PUMP DRIVERS
A mainline horizontal centrifugal pump can be driven by an electric motor, gas turbine or a diesel engine. Liquid hydrocarbon transmission pipelines are typically driven by electric motors where electrical power is available. This is primarily due to their lower initial capital cost and inherent reliability. As well, electric motor drivers have benefits over gas turbine and diesel drivers due to ever more stringent emission level limits.
This section will consider pump station operation using both constant speed trically driven pumps and variable speed electrically driven pumps. Fixed-speed elec-tric motors provide a cost-effective solution for base load applications where elecelec-trical power is available and reliable. They have the advantage of low-maintenance costs and are simple to operate. Variable speed drive (VSD) motors are becoming the standard Figure 4-38. Operating points for series operation
Pumps and Pump stations n 193
for pump stations that have varying flow or product density requirements such as on batched product pipelines. Despite their control systems being more complex than for a constant speed motor, variable speed motors are much more energy efficient. This is because pump capacity can be controlled without the disadvantage of pressure loss incurred by the throttling through a discharge control valve.
Variable speed pumps control the flow and pressures by varying the speed of the drivers with maximum power override. For a pump station that contains both fixed-speed and variable speed motors, the control strategy is to run the fixed speed units at a base load with minimal throttling and use the variable speed units to adjust for the required station set point. Figure 4-39 exhibits the performance curves of a variable speed pump.
In applications that require flow or pressure control, the most energy efficient option is an electronic VSD, referred to as a Variable Frequency Drive (VFD). The most common form of VFD is the voltage-source, pulse-width modulated frequency converter [12]. The converter develops a voltage directly proportional to the frequency which produces a constant magnetic flux in the motor. This type of speed control can be driven by set points of discharge pressure or flow rate.
As energy cost increases, VFDs and thus variable speed pumps are becoming more cost-effective. This is because they allow pressure and flow control without wast-ing energy incurred by throttlwast-ing a control valve to control the discharge pressure of the pipeline. Even though variable frequency drives are more expensive, they have the advantage of reduced energy consumption and are efficient over a wide range of flow.
Energy savings of between 30% and 50% have been achieved in many installa-tions by installing VFDs [13].
The following simplified graph, Figure 4-40 shows the significant reduction of power required by variable speed pumps as compared to fixed speed pumps through a range of flows.
Figure 4-40. Power requirements — constant speed drive vs. variable speed drive
Pumps and Pump stations n 195 Figure 4-41 below shows a pump operating at the operating point 1X, where the flow rate and head are Q1X and H1X, respectively. Here, for illustrative purposes, it can
be assumed that the operating point 1X is the best efficiency point (BEP), where the efficiency of the pump is highest. Then, the power required by the pump is PW1X =
Q1X ´ P1X/(dens ´ ηhBEP), where ηhBEP is the pump efficiency at the BEP.
To achieve a lower flow rate QY, the control valve is partially closed for the fixed
speed pump or the pump speed is reduced for the variable speed pump. These are il-lustrated in Figure 4-41. 1Y is the operating point of the fixed speed pump and 2Y is that of the variable speed pump for the lower flow rate. At 1Y, the flow rate is reduced to Q1Y but the pump head increases to H1Y for the fixed speed pump where the pump
efficiency, h1Y, is considerably lower than the pump efficiency at BEP. In the case of
the variable speed pump, the operating point 2Y for the same flow rate generates only the head required for the system curve, H2Y and the pump efficiency, h2Y, is only
mar-ginally lower than the best efficiency point.
In summary, energy requirements are directly proportional to head generated and are factored by changes in pump efficiency. As demonstrated by Figure 4-41 above, there is considerable energy savings by use of variable speed drives. The above example also does not take into account the continuous pressure loss that occurs in a fully open con-trol valve and therefore is conservative in its energy savings by use of variable speed motors and elimination of the control valve.
See Section 4.10.6.2 for further discussion on the energy savings of variable speed pumps.
4.9 PUMP STATION DESIGN
As described in Section 3.4, intermediate pump station locations on a pipeline system are determined by the hydraulic design of the transmission pipeline and the frictional pressure losses that occur with the product proposed for transportation. Adjustments to station location from its ideal hydraulic location are often required because of such
factors as land access issues or topography constraints. Final locations for intermedi-ate pump stations may result in some slight reduction in pipeline capacity because of some hydraulic imbalance in the final station sites. This section addresses several key equipment and features required for designing a pump station.
4.9.1 Pump Station Diagram
A simple pump station diagram is shown in Figure 4-42. In this diagram, the main com-ponents include two pump units, station piping connecting various comcom-ponents, station isolation valves, bypass valves, pressure control valve, and check valves. In addition, many intermediate pump stations are installed with a pig receiver and launcher set. Pump stations pumping heavy crudes may be equipped with heaters.
Figure 4-42 shows a typical pump station, which is composed of the following equipment and instruments:
Station isolation valves are MOV-103 and MOV-104. These valves are used to ·
isolate the pump station for safety, maintenance, or other operating purposes. Check valves are shown as CKV-101, CKV-102, and CKV-103.
·
Bypass valves include the station bypass valve along the main line, MOV-103 ·
on the suction side, and MOV-104 on the discharge end. The station bypass valve is open and the other two valves are closed when the pump station is shut down, so that the fluid flows through the station bypass valve.
Instrumentation for pressure measurement (
· Ps, Pc, and Pd) is essential for
sta-tion control. Certain pump stasta-tions are equipped with a flow meter for control purposes, not for custody transfer.
A pressure control valve (PCV-101) is installed on the station discharge piping ·
to control the discharge pressure and flow rate.
Pumps are driven by pump drivers which can be either fixed speed type or variable speed type. In some cases, combined fixed and variable speed drivers have been used to take advantage of low cost fixed speed drivers for base load and of low energy cost variable speed drivers for extra load.
4.9.2 Pump Station Piping
Station suction piping design is important to ensure that NPSHA be maintained above the NPSHR for the pump units. Piping must be designed with sufficient pressure to withstand potential surge pressure changes during pump start-up and shut-down opera-tions. Piping sizing should be determined based on minimizing pressure losses to pump suction. As a rule of thumb, flow velocity on the suction side should be in the 1.5 m/sec to 2.5 m/sec range. Higher velocities increase the frictional pressure loss and potential surge pressure. Due to potential surge when a pump shuts down, the suction pressure can suddenly increase significantly. Therefore, the suction line pipe must be able to withstand the increased pressure.
Since centrifugal pumps generate a performance curve that rises as the flow de-creases, the discharge pressure should be determined at shut-off.
If the liquid is fed from a tank, the amount of entrained vapor or air must be kept to an absolute minimum. Entrained vapor causes not only vibration and possibly cavita-tion but also reduced capacity and efficiency. Normally, a booster pump is installed to avoid this condition.
Station piping in pipeline systems that are intended for batched products should be designed to minimize any areas of potential product contamination. All fittings
Pumps and Pump stations n 197 should be close-coupled and piping should be designed to eliminate any connections that could include isolated pockets of product that would contaminate a batch of dif-ferent products.
For pipelines transporting batches of different product, pump units that are not in operation should have their suction and discharge valves left in the open position so that all products flowing through the station piping are continuously being flushed. When there is a need to put the pump unit into operation, the control logic for the unit start should first close the discharge valve before the pump unit is to be started. Once the pump is on-line, the discharge valve is opened and the flow through the pump is re-established with its added head.
In addition, a control valve is required for fixed speed pumps and some variable speed pumps. Some initiating, intermediate pump and pressure reducing stations are equipped with a pig launcher and receiver. In certain situations, heaters are installed to heat heavy crude oil. In cases where the flow rate through a pump can drop below the minimum continuous flow limit, recirculation piping and valves may need to be installed (See Section 4.9.4 Station Flow Recirculation).
4.9.3 Control Valve and Sizing
The proper selection of a valve for petroleum liquid pipelines depends on factors such as liquid properties, the system curve, the pressure drops to be controlled, and cost. The liquid properties to be specified are its phase (whether or not it contains solids or vapor) and its corrosion/erosion property. Temperature may not be a critical factor for most liquid pipeline applications. The most suitable type of valve required for throttling the flow or controlling the pressure includes globe, control ball, butterfly, and rotary plug valves.
Valve size is obtained from a basic liquid sizing equation, which can be written as follows: v P Q C= D g where Q = flow rate
Cv = valve sizing coefficient determined experimentally for each type and size of
valve
DP = pressure differential g = specific gravity of liquid
To calculate the expected Cv for a valve controlling liquid flows, the above equation is
re-arranged for Cv in terms of the flow rate and pressure differential.
v C Q P g = D
We can obtain the valve’s Cv requirements by inputting into this equation the minimum
and maximum flow rates together with valve upstream and downstream pressures. This value is used to select from the manufacturer’s data the size of the valve required. The valve size should be compared to the pipe size in which the valve is to be installed. As a rule of thumb, the control valve should not be smaller than two nominal pipe sizes below the nominal pipe size.
4.9.4 Station Flow Recirculation
Pumps can be operated at or near zero flow for a short time at start-up without adverse con-sequences to the pumping system. To avoid recirculation problems, at least 20% of BEP flow is required for small pumps, and 50% or even 60% of BEP flow for large pumps.
However, if centrifugal pumps operate at low or no flow conditions frequently and the duration of such operation is long, most pumps will need to be provided with a flow recirculation system to protect them from potential damage or unstable flow conditions in a low flow operation. The following conditions can develop at low flow rates:
Temperature in the pump rises significantly due to low pump efficiency; ·
Unstable flow conditions occur, resulting in surging pulsations and pipe vibration. ·
Pumps and Pump stations n 199 If the low flow conditions persist for a prolonged period or occur frequently, pumps and other equipments can be damaged. If a pump station has to pump low flows fre-quently, there are several options to address this problem:
Install two or more small pumps in parallel; ·
Install variable speed pumps; ·
Install a recirculation system with flow measurement, recirculation valve and ·
piping.
The overall economics will dictate the choice of these options.
Several alternative recirculation systems are available. Figure 4-43 shows one such alternative recirculation system, which includes a flow meter, recirculation con-trol valve and piping. When the flow rate gets closer to a set minimum flow, a bypass control valve is activated and the flow downstream of the pump is recycled back to the pump suction through the recirculation piping. As shown in this figure, the recircula-tion system consists of a flow sensing device, a minimum flow control valve, a check valve, and a bypass valve. Note that a pressure safety valve (PSV) is installed down-stream of each pump to prevent the discharge piping from over-pressuring.
Recycling flow rate is controlled by a control valve with the flow meter located in the suction line. As the flow rate decreases, the flow meter sends a signal to open the recirculation control valve to keep the combined flow rate at or above the minimum required flow rate. In general, this system works well in maintaining the minimum flow rate required by the pump. Note that the take-off for the recirculation line is upstream of the check valve.
4.9.5 Pig Launcher and Receiver
Pipeline pigs are extensively used in petroleum pipelines. They are intended to clean and/or inspect the inside of pipelines; or, in some cases, to separate batch interfaces in a multiproduct pipeline to reduce interface mixing. So-called smart pigs can detect pipe corrosion by measuring pipe wall thickness and cracks in the pipe wall by means of ultrasonic responses.
The pig is inserted into the pipeline through a pig launcher. After a pig is loaded into the pig launcher, the MOV-107 valve connected to the pig launcher (refer to Figure 4-42) is opened. When MOV-108 is closed, the flow pushes the pig out of the launcher and the pig travels along the pipeline until it reaches the next pig receiving station or pig trap where the pig is retrieved. The valves along the
