SECTION 12

Pumps & Hydraulic Turbines

Pumps
The most common types of pumps used in gas processing Modern practice is to use centrifugal rather than positive dis-
plants are centrifugal and positive displacement. Occasionally placement pumps where possible because they are usually less
regenerative turbine pumps, axial-flow pumps, and ejectors are costly, require less maintenance, and less space. Conventional
used. centrifugal pumps operate at speeds between 1200 and 8000
rpm. Very high speed centrifugal pumps, which can operate up

FIG. 12-1
Nomenclature

A = cross-sectional area of plunger, piston, or pipe, mm2 S = suction specific speed (units per Equation 12-7)
a = cross-sectional area of piston rod, mm2 sp gr = specific gravity at average flowing conditions.
AC = alternating current Equal to RD
bbl = barrel (42 U.S. gallons or 0.158987 m3) T = torque, N m (Newton meters)
bkW = brake kilowatt tr = temperature rise, C
C = constant (Fig.12-19) u = impeller peripheral velocity, m/s
Cp = specific heat at average temperature, J/(kg C) VE = volumetric efficiency, fraction
D = displacement of reciprocating pump, m3/h VEo = overall volumetric efficiency
DC = direct current VE = volumetric efficiency due to density change
d = impeller diameter, mm VEl = volumetric efficiency due to leakage
e = pump efficiency, fraction Vpd = pulsation dampener volume, mm3
g = 9.80665 m/s2 (acceleration of gravity) v = liquid mean velocity at a system point, m/s
H = total equipment head, m of fluid z = elevation of a point of the system above (+) or
h = head, m of fluid pumped below () datum of the pump. For piping, the
hyd kW = hydraulic kilowatts elevation is from the datum to the piping
k = factor related to fluid compressibility (Fig.12-19) centerline; for vessels and tanks, the elevation is
K = type of pump factor (Equation 12-18) from the datum to the liquid level, m.
kPa = kilopascal Greek:
kPa (abs) = kilopascal, absolute
= density at average flowing conditions, kg/m3
kPa (ga) = kilopascal, gage
i = inlet density, kg/m3
L = length of suction pipe, m
o = outlet density, kg/m3
Ls = stroke length, mm
= allowable pressure fluctuations as a percentage
m = number of plungers or pistons
of mean pressure
NPPP = net positive pipe pressure, kPa (abs)
(NPPP = Px Pvp > 0) Subscripts:
NPSH = net positive suction head of fluid pumped, m a = acceleration
NPSHA = NPSH available, m ave = with P, average pressure in pulsating flow
NPSHR = NPSH required, m bep = best efficiency point, for maximum impeller
n = speed of rotation, revolutions/minute (rpm) diameter
ns = specific speed (See Fig. 12-2 for units) c = compression
N = Polytropic exponent of charge gas. d = discharge of pump
(For nitrogen, N = 1.4) dv = discharge vessel
P = differential pressure, kPa D = displacement
P = pressure, kPa (abs) or kPa (ga) f = friction
Pvp = liquid vapor pressure at pumping temperature, i = inlet of equipment
kPa (abs) l = leakage
Q = rate of liquid flow, m3/h max = with P, maximum acceptable peak pressure
r = ratio of internal volume of fluid between valves,
in pulsating flow
when the piston or plunger is at the end of
the suction stroke, to the piston or plunger min = with P, minimum acceptable valley pressure
displacement. in pulsating flow
RD = relative density of pumped fluid at average flowing o = outlet of equipment
conditions to water density at standard conditions ov = overall
(15C, 1 atm) p = pressure
s = slip or leakage factor for reciprocating and rotary pd = pulsation dampener
pumps

12-1
 FIG. 12-1 (Contd)
Nomenclature

r = rise w = water
s = static, suction of pump, specific, or stroke x = point x in the inlet subsystem
sv = suction vessel y = point y in the outlet subsystem
v = velocity 1 = impeller diameter or speed 1
vp = vapor pressure 2 = impeller diameter or speed 2

DEFINITIONS OF WORDS AND the datum, whose sum is the total head. Also used to ex-
press changes of energy such as the friction losses, the
PHRASES USED IN PUMPS AND equipment total head and the acceleration head.
HYDRAULIC TURBINES
Head, acceleration: The head equivalent to the pressure
Alignment: The straight line relation between the pump change due to changes in velocity in the piping system.
shaft and the driver shaft.
HPRT: Hydraulic power recovery turbine.
Casing, axially split: A pump case split parallel to the pump
shaft. Impeller: The bladed member of the rotating assembly of a
centrifugal pump which imparts the force to the liquid.
Casing, radially split: A pump case split transverse to the
pump shaft. NPSHA: The total suction absolute head, at the suction nozzle,
referred to the standard datum, minus the liquid vapor ab-
Cavitation: A phenomenon that may occur along the flow solute pressure head, at flowing temperature available for a
path in a pump when the absolute pressure equals the liq- specific application. For reciprocating pumps it includes the
uid vapor pressure at flowing temperature. Bubbles then acceleration head. NPSHA depends on the system character-
form which later implode when the pressure rises above the istics, liquid properties and operating conditions.
liquid vapor pressure.
NPSHR: The minimum total suction absolute head, at the
Coupling: A device for connecting the pump shaft to the driv- suction nozzle, referred to the standard datum, minus the
er shaft consisting of the pump shaft hub and driver shaft liquid vapor absolute pressure head, at flowing tempera-
hub, usually bolted together. ture, required to avoid cavitation. For positive displace-
Coupling, spacer: A cylindrical piece installed between the ment pumps it includes internal acceleration head and
pump shaft coupling hub and driver shaft coupling hub, to losses caused by suction valves and effect of springs. It does
provide space for removal of the mechanical seal without not include system acceleration head. NPSHR depends on
moving the driver. the pump characteristics and speed, liquid properties and
flow rate and is determined by vendor testing, usually with
Cutwater: The point of minimum volute cross-sectional area, water.
also called the volute tongue.
Pelton wheel: A turbine runner which turns in reaction to
Datum elevation: The reference horizontal plane from which the impulse imparted by a liquid stream striking a series of
all elevations and heads are measured. The pumps stan- buckets mounted around a wheel.
dards normally specify the datum position relative to a
pump part, e.g. the impeller shaft centerline for centrifugal Recirculation control: Controlling the quantity of flow
horizontal pumps. through a pump by recirculating discharge liquid back to
suction.
Diffuser: Pump design in which the impeller is surrounded
by diffuser vanes where the gradually enlarging passages Rotor: The pump or power recovery turbine shaft with the
change the liquid velocity head into pressure head. impeller(s) mounted on it.

Displacement: The calculated volume displacement of a posi- Rotor, Francis-type: A reverse running centrifugal pump
tive displacement pump with no slip losses. impeller, used in a hydraulic power recovery turbine, to
convert pressure energy into rotational energy.
Double acting: Reciprocating pump in which liquid is dis-
charged during both the forward and return stroke of the Run-out: The point at the end of the head-capacity perfor-
piston. mance curve, indicating maximum flow quantity and usu-
ally maximum brake power.
Duplex: Pump with two plungers or pistons.
Runner: The shaft mounted device in a power recovery tur-
Efficiency, mechanical: The ratio of the pump hydraulic bine which converts liquid pressure energy into shaft
power output to pump power input. power.
Efficiency, volumetric: The ratio of a positive displacement Shut-off: The point on the pump curve where flow is zero,
pump suction or discharge capacity to pump displacement. usually the point of highest total dynamic head.
Head: The flowing liquid column height equivalent to the Simplex: Pump with one plunger or piston.
flowing liquid energy, of pressure, velocity or height above

12-2
Single acting: Reciprocating pump in which liquid is dis- Triplex: Pump with three plungers or pistons.
charged only during the forward stroke of the piston.
Vanes, guide: A series of angled plates (fixed or variable) set
Slip: The quantity of fluid that leaks through the internal around the circumference of a turbine runner to control the
clearances of a positive displacement pump per unit of fluid flow.
time. Sometimes expressed on a percentage basis.
Volute, double: Spiral type pump case with two cutwaters
Surging: A sudden, strong flow change often causing exces- 180 apart, dividing the flow into two equal streams.
sive vibration.
Volute, single: Spiral type pump case with a single cutwater
Suction, double: Liquid enters on both sides of the impeller. to direct the liquid flow.
Suction, single: Liquid enters one side of the impeller. Vortex breaker: A device used to avoid vortex formation in
the suction vessel or tank which, if allowed, would cause
Throttling: Controlling the quantity of flow by reducing the
vapor entrainment in the equipment inlet piping.
cross-sectional flow area, usually by partially closing a
valve.

to 23 000 rpm and higher, are used for low-capacity, high-head When the elevation and size of inlet and outlet nozzles are the
applications. Most centrifugal pumps will operate with an ap- same, the equipment total head (H) equals the difference of
proximately constant head over a wide range of capacity. pressure heads.
Positive displacement pumps are either reciprocating or When using any suction-and-discharge-system points, the fol-
rotary. Reciprocating pumps include piston, plunger, and dia- lowing general equation applies.
phragm types. Rotary pumps are: single lobe, multiple lobe, ro-
tary vane, progressing cavity, and gear types. Positive displace- 1000 Px vx2 1000 Py vy2
zx + + hfx + H = zy + + + hfy
ment pumps operate with approximately constant capacities g 2g g 2g
over wide variations in head, hence they usually are installed Eq 12-4
for services which require high heads at moderate capacities. A
special application of small reciprocating pumps in gas process- When the points are located in tanks, vessels or low velocity
ing plants is for injection of fluids (e.g. methanol and corrosion points in the piping, the velocity head is normally negligible,
inhibitors) into process streams, where their constant-capacity but may not be negligible in equipment nozzles. Note that
characteristics are desirable. the subscripts i and o are used for variables at pumps and
HPRTs inlet and outlet nozzles, respectively, while the sub-
Axial-flow pumps are used for services requiring very high scripts s and d are used only for variables at pumps suction
capacities at low heads. Regenerative-turbine pumps are used and discharge nozzles. The subscripts x and y are used for
for services requiring small capacities at high heads. Ejectors variables at points in each inlet and outlet subsystem and usu-
are used to avoid the capital cost of installing a pump, when a ally are suction and discharge vessels. Also x and y are used
suitable motive fluid (frequently steam) is available, and are for friction head from point x to equipment inlet nozzle and
usually low-efficiency devices. These kinds of pumps are used from equipment outlet nozzle to point y.
infrequently in the gas processing industry.
The work done in compressing the liquid is negligible for
Fig. 12-1 provides a list of symbols and terms used in the text practically incompressible liquids and it is not included in the
and also a glossary of terms used in the pump industry. Fig. 12- above equations. To evaluate the total head more accurately
2 is a summary of some of the more useful pump equations. Fig. when handling a compressible liquid, the compression work
12-3 provides guidance in selecting the kinds of pumps suitable should be included. If a linear relationship between density and
for common services. pressure is assumed, the liquid compression head that substi-
tutes for the difference of pressure heads in above equations is:
EQUIPMENT AND SYSTEM EQUATIONS 500 (Po Pi) 1 1
o i
Hc = + Eq 12-5
The energy conservation equation for pump or hydraulic g
turbine systems comes from Bernoullis Theorem and relates When the differential pressure is sufficiently high to have
the total head in two points of the system, the friction losses a density change of more than 10%, or when the pressure is
between these points and the equipment total head. Elevations near the fluids critical pressure, the change in fluid density
are measured from the equipment datum. and other properties with pressure is not linear. In these cases
The total head at any system point is: Equations 12-3 to 12-5 may not be accurate. A specific fluid
properties relationship model is required in this case. For pure
1000 P v2 substances, a pressure-enthalpy-entropy chart may be used for
h = z + hp + hv = z + + Eq 12-1
g 2g estimating purposes by assuming an isentropic process. The
pump manufacturer should be consulted for the real process,
The system friction head is the inlet system friction head plus
including the equipment efficiency, heat transfer, etc. to deter-
the outlet system friction head:
mine the equipment performance.
hf = hfx + hfy Eq 12-2
The equipment total head is the outlet nozzle total head minus NET POSITIVE SUCTION HEAD
the inlet nozzle total head. H is positive for pumps and negative
for HPRTs: See NPSH definition in Fig. 12-1. There should be sufficient
net positive suction head available (NPSHA) for the pump to
1000 (Po Pi) v 2o v2i
H = ho hi = zo zi + + Eq 12-3 work properly, without cavitation, throughout its expected ca-
g 2g pacity range. Usually a safety margin of about 0.6 to 1 m of

12-3
 FIG. 12-2
Common Pump Equations

FLOW RATE

Given
multiply by to US gal/min UK gal/min ft3/sec bbl/day liters/s Kg/h
get
m3/h 0.227 125 0.272 766 101.941 6.624 47 103 3.600 00 1/(999.102 RD)
PRESSURE

Given ft water at * std atm 760

multiply by to lb/in2 m water at 0C m liquid bar kgf/cm2
39.2F mm Hg at 0C
get
kPa 6.894 76 2.988 99 9.805 11 g/1000 100 101.325 98.0665
DENSITY

Given
multiply by to lb/ft3 lb/US gal lb/UK gal kg/lt API gravity Baum gravity
get
kg/m3 16.0185 119.826 99.7763 1000 See Fig. 1-3

P 1000 P Q H RD QHg Q H RD Q H g
p=
h = hyd
kW = =
bkW = =
g RD 0.999 102 g 367.428 3 600 000 367.428 e 3 600 000 e
v2 Q P** Q P**
v =
h = =
2g 3600 3600 e (for pumps)
dn hyd kW (9549.30) (bkW)
u = bkW
= (for pumps) =
T
60 000 e n
(Q) (277.778) bkW = hyd kW e (for turbines) 14
v
= n Qbep n Hbep Qbep
A RD = relative density =
n =
s 34
Hbep Hbep
1 HP = 0.745 700 kW Water density at 15C = 999.102 kg/m3
= 550 ft lbf/s Standard gravity acceleration: See Fig. 1-7 for viscosity relationships
= 33,000 ft lbf/min g = 9.806 65 m/s2 = 32.1740 ft/s2 * Standard atmospheric pressure:
1 atm = 760 mm Hg = 101.325 kPa =
14.6959 psi
** See Equation 12-3 and 12-4.

CENTRIFUGAL PUMPS AFFINITY LAWS

1: Values at initial conditions
2: Values at new conditions
CHANGE SPEED DIAMETER SPEED AND DIAMETER
Q2 = Q1 (n2/n1) Q1 (d2/d1) Q1 (d2/d1) (n2/n1)
h2 = h1 (n2/n1) 2
h1 (d2/d1)
2
h1 [(d2/d1) (n2/n1)]2
bkW2 = bkW1 (n2/n1)3 bkW1 (d2/d1)3 bkW1 [(d2/d1) (n2/n1)]3
NPSHR2 = NPSHR1 (n2/n1) 2
NPSHR1 (n2/n1)2

NPSHA above NPSHR is adequate. Cavitation causes noise, When the pump suction nozzle pressure is not known, but
impeller damage, and impaired pump performance. Consider- the pressure at any point (x) of the suction system is known,
ation must also be given to any dissolved gases which may af- NPSHA may be calculated with the following equation, where
fect vapor pressure. For a given pump, NPSHR increases with hfx is the head friction loss from the point x to the suction noz-
increasing flow rate. If the pump suction nozzle pressure is zle. With commonly used suction pipe diameters, the velocity
known head may be negligible.
1000 (Pi Pvp) v 2i 1000 (Px Pvp) v x2
NPSHA = + zi + Eq 12-6a NPSHA = + zx + hfx Eq 12-6b
g 2g g 2g

12-4
 Moreover, when the suction system point is the specific case rotary pumps may be considered, even though their capacity is
of the suction vessel, the equation is the following, where hfv affected by entrained and dissolved air or gases. See Hydraulic
is the head friction loss from the suction vessel to the suction Institute Standards.5
nozzle.
Datum
1000 (Psv Pvp)
NPSHA = + zsv hfsv Eq 12-6c The pump datum elevation is a very important factor to
g
consider and should be verified with the manufacturer. Some
The pressures in the above equations must be both absolute common references are shown in Fig. 12-4. Some manufactur-
or gage; when using gage pressure both must be relative to the ers provide two NPSHR curves for vertical can pumps, one for
same atmospheric pressure. To convert a system pressure gage the first stage impeller suction eye and the other for the suction
reading to absolute pressure add the existing local atmospheric nozzle.
pressure. The fluid vapor pressure must be at operating tem-
perature. If the fluid vapor pressure is given in gage pressure, NPSH Correction Factors
check which atmospheric pressure is reported. The use of the
true local atmospheric pressure is very important in the cases NPSHR is determined from tests by the pump manufac-
of high altitude locations, and of a close margin of NPSHA over turer using water near room temperature and is expressed in
the NPSHR. height of water. When hydrocarbons or high-temperature water
are pumped, less NPSH is required than when cold water is
The pressure shall be measured at the pipe or nozzle center- pumped. Hydraulic Institute correction factors for various liq-
line height; otherwise, adequate correction shall be made. Pay uids are reproduced in Fig. 12-5. Some users prefer not to use
special attention to large pipe or nozzle diameters and the el- correction factors to assure a greater design margin of safety.
evation of gage attached to them, pole or panel mounted instru-
ment elevation, and different density fluid in the instrument NPSH and Suction Specific Speed
line, see Hydraulic Institute Standards.5
Suction specific speed is an index describing the suction ca-
To avoid vapor formation in the suction system, there must pabilities of a first stage impeller and can be calculated using
also be a Net Positive Pipe Pressure (NPPP) along it. There- Equation 12-7. Use half of the flow for double suction impel-
fore, for every suction line point and operating condition the lers.
line pressure, at the top of the pipe must be higher than the

fluid vapor pressure, being the pressure determined taking into nQbep Eq 12-7
account the pipe elevation. S=
NPSHR bep 3/4
The entrained and dissolved air or gases in the pump suc-
Pumps with high suction-specific speed tend to be suscep-
tion affects the pump performance, both mechanically and hy-
tible to vibration (which may cause seal and bearing problems)
draulically, especially when the suction nozzle pressure is lower
when they are operated at other than design flow rates. As a re-
than the suction vessel pressure. In centrifugal pumps it causes
sult, some users restrict suction specific speed, and a widely ac-
the reduction of capacity and discharge pressure, because of the
cepted maximum is 11,000. For more details on the significance
reduced overall density; and also, at low flow, the impeller cen-
trifugal action separates the gas from the liquid resulting in the
cessation of the liquid flow. For these cases, specially designed
centrifugal pumps with higher tolerance to gas entrainment or FIG. 12-4
Datum Elevation

FIG. 12-3 Pump type Standard Datum elevation

Pump Selection Guide Centrifugal, hori- API 610 1
Shaft centerline
zontal Hydraulic Institute5
Centrifugal, vertical API 6101 Suction nozzle
in-line centerline
Centrifugal, other API 6101 Top of the founda-
vertical tion
Centrifugal, vertical Hydraulic Institute5 Entrance eye to the
single suction, volute first stage impeller
and diffused vane
type
Centrifugal, vertical Hydraulic Institute5 Impeller discharge
double suction horizontal centerline
Vertical turbine. AWWA E10118 Underside of the
Line shaft and sub- discharge head or
mersible types head baseplate
Reciprocating Hydraulic Institute5 Suction nozzle
centerline
Rotary Hydraulic Institute5 Reference line or
suction nozzle
centerline

12-5
 FIG. 12-5 The pump nozzles sizes and elevations.
NPSHR Reduction for Centrifugal Pumps Handling The minimum elevation (referred to the datum) of
Hydrocarbon Liquids and High Temperature Water liquid expected in the suction vessel.
The maximum elevation (referred to the datum)
to which the liquid is to be pumped.
The head loss expected to result from each compo-
nent which creates a frictional pressure drop at
design capacity.
3. Use appropriate equations (Equations 12-112-4).
4. Convert all the pressures, frictional head losses, and
static heads to consistent units (usually kPa or meters of
head). In 5 and 6 below, any elevation head is negative if
the liquid level is below the datum. Also, the vessel pres-
sures are the pressures acting on the liquid surfaces.
This is very important for tall towers. In the case of par-
titioned vessels, be sure to use the corresponding cham-
ber pressure and liquid level elevation. And when the
liquid is not a continuous phase, or it is not clear where
the liquid level is, as in the case of packed fractionating
towers, consider only the piping and exclude such vessels
from the system.
5. Add the static head to the suction vessel pressure, then
subtract the frictional head losses in the suction piping.
This gives the total pressure (or head) of liquid at the
pump suction flange.
6. Add the discharge vessel pressure, the frictional head
losses in the discharge piping system, and the discharge
static head. This gives the total pressure (or head) of liq-
uid at the pump discharge. According to the type of ca-
pacity and head controls, pump type and energy conser-
vation, required for the particular situation, provide a
of suction specific speed, consult pump vendors or references head and/or a flow additional margin to provide a good
listed in the References section. control. A control valve to throttle the discharge or to re-
circulate the flow, or a variable speed motor, etc. may be
Submergence the options to provide good control.
The suction system inlet or the pump suction bell should 7. Calculate the required pump total head by subtracting
have sufficient height of liquid to avoid vortex formation, which the calculated pump suction total pressure from the cal-
may entrain air or vapor into the system and cause loss of ca- culated pump discharge total pressure and converting to
pacity and efficiency as well as other problems such as vibra- head.
tion, noise, and air or vapor pockets. Inadequate reservoir ge-
ometry can also cause vortex formation, primarily in vertical 8. It is prudent to add a safety factor to the calculated pump
submerged pumps. Refer to the Hydraulic Institute Standards5 head to allow for inaccuracies in the estimates of heads
for more information. and pressure losses, and pump design. Frequently a safe-
ty factor of 10% is used, but the size of the factor used for
CALCULATING THE REQUIRED each pump should be chosen with consideration of:
DIFFERENTIAL HEAD The accuracy of the data used to calculate the re-
quired head
The following procedure is recommended to calculate the
head of most pump services encountered in the gas processing The cost of the safety factor
industry. See Example 12-1.
The problems which might be caused by install-
1. Prepare a sketch of the system in which the pump is to be ing a pump with inadequate head.
installed, including the upstream and downstream ves-
Example 12-1 Liquid propane, at its bubble point, is to be
sels (or some other point at which the pressure will not be
pumped from a reflux drum to a depropanizer. The maximum
affected by the operation of the pump). Include all compo-
flow rate is expected to be 82 m3/h. The pressures in the vessels
nents which might create frictional head loss (both suc-
are 1380 and 1520 kPa(abs) respectively. The relative density
tion and discharge) such as valves, orifices, filters, and
of propane at the pumping temperature (38C) is 0.485. The el-
heat exchangers.
evations and estimated frictional pressure losses are shown on
2. Show on the sketch: Fig. 12-6. The pump curves are shown in Fig. 12-7. The pump
nozzles elevations are zero and the velocity head at nozzles is
The datum position (zero elevation line) according negligible.
to the proper standard. See Fig. 12-4.

12-6
Required differential head is determined as follows: FIG. 12-6
Absolute Total Pressure at Pump Suction Example 12-1 Depropanizer
Reflux drum 1380kPa(abs)
Elevation 6 m0.9990.4859.807 = +28.5kPa
Friction piping 3.5kPa
valves 1.4kPa

1403.6kPa(abs)
= 1302.3kPa(ga)
Absolute Total Pressure at Pump Discharge
Tower 1520kPa(abs)
Elevation 22.5 m0.9990.4859.807 = +106.9kPa
Friction piping +20.7kPa
valves +13.8kPa
orifice +8.3kPa
filter +89.6kPa
check valve +6.9kPa
control valve +62.1kPa

1828.3kPa(abs)
= 1727.0kPa(ga)
Differential pressure = 1727 1302.3 = 424.7 kPa
(424.7)
Differential head = H = = 89.4 m
(0.485) (0.999)(9.807)

FIG. 12-7
Depropanizer Reflux Pump for Example 12-1

12-7
 10% safety factor 9m Therefore a 25kW motor is selected for the pump driver to pro-
vide full curve protection.
Required differential head (H) 98.4m
Calculation of NPSHA CENTRIFUGAL PUMPS
Reflux drum pressure 1380kPa(abs)
Figs. 12-8a through 12-8e are cross-sectional drawings
Elevation 6 m0.9990.4859.807 = +28.5kPa
showing typical configurations for five types of centrifugal
Friction valves = 1.4kPa
pumps. A guide to selecting centrifugal pumps is shown in Fig.
piping = 3.5kPa
12-9. Horizontal centrifugal pumps are more common; however,
Fluid vapor pressure 1380kPa(abs)
vertical pumps are often used because they are more compact
23.6kPa and, in cold climates, may need less winterizing than horizontal
NPSHA 23.6/(0.9990.4859.807) = 5.0m pumps. The total installed cost of vertical pumps is frequently
lower than equivalent horizontal pumps because they require
This NPSHA result is adequate when compared to the 3 m smaller foundations and simpler piping systems.
of NPSHR in the curve shown in Fig. 12-7.
Vertical can pumps are often used for liquids at their bub-
Calculation of Hydraulic Power ble-point temperature because the first stage impeller is located
below ground level and therefore requires less net positive suc-
Q H RD tion head at the suction flange. The vertical distance from the
hyd kW = (from Fig. 12-2)
367 suction flange down to the inlet of the first stage impeller pro-
vides additional NPSHA.
(82) (98.4) (0.485)
hyd kW = = 10.67 kW
367 Centrifugal Pump Theory
Calculation of Actual Power Centrifugal pumps increase the pressure of the pumped
fluid by action of centrifugal force on the fluid. Since the total
hyd kW head produced by a centrifugal pump is independent of the den-
bkW = (from Fig. 12-2)
e sity of the pumped fluid, it is customary to express the pressure
increase produced by centrifugal pumps in feet of head of fluid
Fig. 12-7 is the performance curve of the selected pump. The pumped.
efficiency at rated capacity and required head is 62%, with a
brake kilowatt calculated as follows:
FIG. 12-8b
10.67 kW
bkW = = 17.2 bkW Vertical Inline Pump
0.62

Motor Sizing
The maximum flow is 115m3/h with a head of 75m for this
particular pump impeller size, which results in a brake kilowatt
requirement of 19.5bkW at run-out (i.e., end of head curve).

FIG. 12-8a
Horizontal Single Stage Process Pump

12-8
 FIG. 12-8c Operating characteristics of centrifugal pumps are expressed
in a pump curve similar to Fig. 12-7. Depending on impeller de-
Horizontal Multi-Stage Pump sign, pump curves may be drooping, flat, or steep. Fig. 12-
10 shows these curves graphically. Pumps with drooping curves
tend to have the highest efficiency but may be undesirable be-
cause it is possible for them to operate at either of two flow rates
at the same head. The influence of impeller design on pump
curves is discussed in detail in Hydraulic Institute Standards.5

Specific Speed
Specific speed gives an indication of the impeller shape and
pump characteristics, as can be seen in the Fig. 12-11, from the
Hydraulic Institute Standards. The ratios of major dimensions
vary uniformly with specific speed. Specific speed is given by
the equation in Fig. 12-2.

Affinity Laws
The relationships between rotational speeds, impeller di-
ameter, capacity, head, power, and NPSHR for any particular
pump are defined by the affinity laws (See Fig. 12-2 for affinity
laws). These equations are to predict new curves for changes in
impeller diameter and speed.
The capacity of a centrifugal pump is directly proportional
to its speed of rotation and its impeller diameter. The total
pump head developed is proportional to the square of its speed
and its impeller diameter. The power consumed is proportional

FIG. 12-8d FIG. 12-8e

Vertical Can Pump Vertical, High Pressure, Double Case, Multi-Stage Pump

12-9
to the cube of its speed and its impeller diameter. The NPSHR and six pole motors. These charts were developed with data pro-
is proportional to the square of its speed. vided courtesy of Flowserve Corporation.
These equations apply in any consistent set of units but only Viscosity
apply exactly if there is no change of efficiency when the rota-
tional speed is changed. This is usually a good approximation if Most liquids pumped in gas processing plants have viscosi-
the change in rotational speed is small. ties in the same range as water. Thus they are considered non-
viscous and no viscosity corrections are required. Occasionally
A different impeller may be installed or the existing modi- fluids with viscosities higher than 5 10-6m2/s are encountered
fied. The modified impeller may not be geometrically similar to (e.g. triethylene glycol, 40 10-6m2/s at 20C) and corrections to
the original. An approximation may be found if it is assumed head, capacity, and power consumption may be required.
that the change in diameter changes the discharge peripheral
velocity without affecting the efficiency. Therefore, at equal ef- Viscosity correction charts and the procedures for using
ficiencies and rotational speed, for small variations in impeller them are included in Hydraulic Institute Standards.5
diameter, changes may be calculated using the affinity laws.
Matching the Pump to the
These equations do not apply to geometrically similar but
different size pumps. In that case dimensional analysis should
System Requirements
be applied. A pump curve depicts the relationship between the head
and capacity of a pump. A system curve shows the relationship
The affinity equations apply to pumps with radial flow im-
between the total head difference across the system and the
pellers, that is, in the centrifugal range of specific speeds, below
flow rate through it. The total head difference consists of three
4200. For axial or mixed flow pumps, consult the manufacturer.
components: static (gravity) head, pressure head, and head-loss
See Fig. 12-2 for specific speed equation.
due to friction. Static and pressure heads do not change with
Efficiency flow. However, frictional losses usually increase approximately
as the square of the flow rate through the system. If the system
Fig. 12-13 provides centrifugal pump optimum generally at- curve is plotted with the same units as the pump curve, it can
tainable efficiency vs. flow for several pump types for two, four, be superimposed as shown in Fig. 12-12.
For pump selection, the shape and slope of the pump curve
shall be considered in its position with respect to the system
FIG. 12-9 curve. When the curves are approximately perpendicular to
each other, the change in the operating point position due to de-
Pump Selection Guide Centrifugal Pumps
viations in the curves will be minimum. In addition, the shape
and slope shall be considered when several pumps are used in
series and/or parallel operation to produce the desired range of
flow and/or operating pressure. Refer to Fig. 12-14 and Fig. 12-
15 for series and parallel operation.
Throttling Control If a centrifugal pump and a system
were matched as shown in Fig. 12-12, the flow rate through

FIG. 12-10
Example Centrifugal Pump Head Curves

12-10
 FIG. 12-11
Values of Specific Speeds (ns)

Section Speeds
US Units US Units

1500
1000
500

600

700
800
900

2000

3000

4000

5000

6000

7000
8000
9000

15000

20000
10000
Impeller shrouds

Impeller shrouds
Impeller shrouds
Impeller shrouds

Impeller hub
Hub Hub Vanes
Hub Vanes Hub Vanes Vanes
Vanes Axis of
Radial-vane area Francis-vane area Mixed-flow area Axial-flow area rotation

Metric Metric
1000

10000
600

2000

3000

4000

20000
Note:
Profiles of several pump impeller designs ranging from the Low Specific Speed Radial Flow on the left to a High Specific
Speed Axial Flow on the right, placed according to where each design fits on the Specific Speed Scale.
See Fig. 12-2 for units.

the suction. This control method is used more frequently for

FIG. 12-12 positive displacement pumps than for centrifugal pumps, since
Example Combined Pump-System Curves the discharge of most positive displacement pumps should not
be throttled. This control method should be used with caution
for centrifugal pumps, since a wide-open recirculation may re-
sult in a head so low that the pumped fluid will be circulated
back to the suction at an extremely high rate, causing high
power consumption, increase in fluid temperature, and possibly
cavitation, as well as possibly overloading the driver.
Speed Control Another way of regulating centrifugal
pump capacity is to adjust the rotational speed of the pump. This
is frequently not easily done because most pumps are driven by
fixed-speed motors. However, pumps controlled by adjusting the
rotational speed often consume substantially less energy than
those controlled in other ways. The changed power consumption
can be calculated by Equation 12-8, which assumes that the fric-
tional head is proportional to the square of the flow rate.
e1 hs (Q2/Q1) + hf1 (Q2/Q1)3
e2
bkW2 = bkW1 Eq 12-8
hs + hf1
subscript 1 refers to initial flow rate
subscript 2 refers to the changed flow rate
the system will be A unless some kind of flow control is pro- hs (static) is equivalent to the zero flow system total head
vided. Control usually is provided by throttling a valve in the
discharge piping of the pump, which creates extra frictional On-Off Control Pump capacity can be controlled by
losses so that pump capacity is reduced to that required. In Fig. starting and stopping the pump manually or by an automatic
12-12, the required flow rate is represented by B. Required control such as pressure, level or temperature switches.
amount of extra frictional losses to achieve a flow rate of B is
represented on Fig. 12-12 by the difference between HB-PUMP
Temperature Rise Due to Pumping
and HB-SYSTEM. Frequently the throttling valve is an automatic When a liquid is pumped, its temperature increases because
control valve which holds some plant condition constant (such the energy resulting from the inefficiency of the pump appears
as liquid level, flow rate, or fluid temperature). This control as heat.
method consumes energy since it artificially increases the sys-
1
e
tem resistance to flow. 9.8067 H 1

Recirculation Control Pump capacity can also be con- tr = Eq 12-9
Cp
trolled by recirculating a portion of the pumped fluid back to

12-11
 FIG. 12-13
Optimum Generally Attainable Efficiency Chart

Pump Efficiency versus Flow and Head at 2900 RPM

100

90

80
Pump Efficiency (%)

70

60 m
0
30
m 50 m
50 l 0
ee 10 1
r wh
40 pe
m
15 m m
30 30 50

20

10

0
1 10 100 1 000 10 000

Pump Efficiency versus Flow and Head at 1450 RPM

100

90

80
Pump Efficiency (%)

70
m
60
0m 300
15
50 0m
m 10
50
40 el m
he 30
rw
30 pe
m
15
20

10

0
1 10 100 1 000 10 000

Pump Efficiency versus Flow and Head at 970 RPM

100

90

80
Pump Efficiency (%)

70

60
m
50
m 100
50
l m 3 0m
ee 15
40 r wh m
pe 10
30 5m

20

10

0
1 10 100 1 000 10 000

The above figures indicate expected pump efficiencies for pumps close to the design conditions. The charts shown cover two pole (2900 RPM),
four pole (1450 RPM) and six pole (970 RPM) 50 Hz induction motors with typical slip. Please note that charts are provided for 60 Hz systems
(in english units) in the FPS version of this Data Book. These charts were developed with data provided courtesy of Flowserve Corporation.

12-12
 Usually when the pump is running normally, the tempera- curves rise steadily to shut-off. A drooping curve gives two pos-
ture rise is negligible. However, if the pump discharge is shut sible points of operation, and the pump load may oscillate be-
off, all energy is converted to heat and since there is no fluid tween the two causing surging.
flow through the pump to carry the heat away, the liquid in
the pump will heat rapidly and eventually vaporize. This can Drivers
produce catastrophic failures, particularly in large multistage
Most pumps used in gas processing service are driven by
pumps.
electric motors, usually fixed speed induction motors.
Pump vendors should be requested to provide data on mini-
API Standard 610, Section 3.1.4. (Drivers), states:
mum flow.
Motors shall have power ratings, including the service fac-
Expensive pumps, such as large multistage units, can be
tor (if any), at least equal to the percentages of power at
protected by installing minimum flow recirculation which will
pump rated conditions given in. . . the next table. Howev-
ensure an adequate flow through the pump.
er, the power at rated conditions shall not exceed the motor
Series and Parallel Operation nameplate rating. Where it appears that this procedure will
lead to unnecessary oversizing of the motor, an alternate
Often pumps are installed in series or in parallel with other proposal shall be submitted for the purchasers approval.
pumps. In parallel, the capacities at any given head are added;
in series, the heads at any given capacity are added. A multi- Motor Nameplate Rating Percentage of Rated
stage pump is in effect a series of single stage units. Figs. 12- kW hp Pump Power
14 and 12-15 show series and parallel pumps curves, a system
curve, and the effect of operating one, two or three pumps in a <22 <30 125
system. In both figures, the operating points for both pumps A 22-55 30-75 115
and B are the same only when one pump is operating. For 2 or
>55 >75 110
3 pumps operating, the points are not the same because of the
pump curve shapes. Hence, due consideration should be given Alternatives to electric motor drivers are:
to the pump curve shape when selecting pumps for series or
parallel operation. internal combustion engines
Parallel operation is most effective with identical pumps; gas turbines
however, they do not have to be identical, nor have the same steam turbines
shut-off head or capacity to be paralleled. When pumps are
operating in parallel it is imperative that their performance hydraulic power-recovery turbines

FIG. 12-14
Series Pumps Selection

12-13
 FIG. 12-15
Parallel Pumps Selection

Usually the speed of rotation of these drivers can be varied tions before the pump is started. If the operating temperature is
to provide control. greatly different from the temperature at which the alignment
was performed, the alignment should be checked, and adjusted
Variable Speed Drives Fig. 12-17 lists various types if necessary, at the operating temperature.
of adjustable speed drives, their characteristics and their ap-
plication. Pump and piping supports should be designed and installed
so that forces exerted on the pump by the piping will not cause
Materials of Construction pump misalignment when operating temperature changes or
other conditions occur.
Pumps manufactured with cast-steel cases and cast-iron in-
ternals are most common in the gas processing industry. API The shaft coupling should be selected to match the power
Std 610 is a good reference for material selection. The material transmitted and the type of pump and driver. A spacer type
selections in this document can be over-ridden as required to coupling should be used if it is inconvenient to move either the
reflect experience. pump or the driver when the seal (or other component) requires
maintenance.
Experience is the best guide to selection of materials for
pumps. Process pump manufacturers can usually provide sug- Piping
gestions for materials, based on their experience and knowl-
edge of pumps. Pump requirements, nozzle size, type of fluid, temperature,
pressure and economics determine materials and size of piping.
Shaft Seals
Suction lines should be designed to keep friction losses to
Mechanical seals are the most common sealing devices for a minimum. This is accomplished by using an adequate line
centrifugal pumps in process service. The purpose of the seal size, long radius elbows, full bore valves, etc. Pockets where
is to retain the pumped liquid inside the pump at the point air or vapor can accumulate should be avoided. Suction lines
where the drive shaft penetrates the pump body. Mechanical should be sloped, where possible, toward the pump when it is
seals consist of a stationary and a rotating face, and the actual below the source, and toward the source when it is below the
sealing takes place across these very smooth, precision faces. pump. Vertical downward suction pipes require special care
Seal faces may require cooling and lubrication. API Std 610 de- to avoid pulsation and vibrations that can be caused by air
scribes seal flush systems used to cool the seal faces and remove or vapor entrainment. Elbows entering double suction pumps
foreign material. Seal manufacturers can provide application should be installed in a position parallel to the impeller. Suf-
and design information. ficient liquid height above the suction piping inlet, or a vortex
breaker, should be provided to avoid vortex formation which
Alignment, Supports, and Couplings may result in vapors entering the pump. Suction vessel tan-
The alignment of the pump and driver should be checked and gential inlets and centrifugal pumps may induce a vortex in
adjusted in accordance with the manufacturers recommenda- the vessel and pump suction line, opening a vapor core that

12-14
 FIG. 12-16
Check List for Centrifugal Pump Troubles and Causes

Trouble: Possible Causes: Trouble: Possible Causes:

1. Failure a. Wrong direction of rotation 5. Pump over- a. Speed too high
to deliver loads driver
liquid b. Pump not primed b. Developed head greater than rated head
c. Suction line not filled with liquid c. Excessive recirculation
d. Air or vapor pocket in suction line
d. Either or both the specific gravity and viscosity of
e. Inlet to suction pipe not sufficiently submerged liquid different from that for which pump is rated
f. Available NPSH not sufficient e. Mechanical defects:
g. Pump not up to rated speed (1) Misalignment
h. Total head required greater than head which (2) Shaft bent
pump is capable of delivering (3) Rotating element dragging
2. Pump does a. Wrong direction of rotation (4) Packing too tight
not deliver
rated b. Suction line not filled with liquid 6. Vibration a. Starved suction
capacity c. Air or vapor pocket in suction line (1) Gas or vapor in liquid
d. Air leaks in suction line or stuffing boxes (2) Available NPSH not sufficient
e. Inlet to suction pipe not sufficiently submerged. (3) Inlet to suction line not sufficiently
f. Available NPSH not sufficient submerged
g. Pump not up to rated speed (4) Gas or vapor pockets in suction line
h. Total head greater than head for which pump b. Misalignment
designed
c. Worn or loose bearings
j. Foot valve too small
d. Rotor out of balance
k. Foot valve clogged with trash
(1) Impeller plugged
m. Viscosity of liquid greater than that for which
pump designed (2) Impeller damaged
n. Mechanical defects: e. Shaft bent
(1) wearing rings worn f. Improper location of control valve in discharge
line
(2) Impeller damaged
g. Foundation not rigid
(3) Internal leakage resulting
from defective gaskets 7. Stuffing a. Packing too tight
boxes
o. Discharge valve not fully opened overheat b. Packing not lubricated
3. Pump does a. Gas or vapor in liquid c. Wrong grade of packing
not develop
rated b. Pump not up to rated speed d. Insufficient cooling water to jackets
discharge c. Discharge pressure greater than pressure for e. Box improperly packed.
pressure which pump designed
8. Bearings a. Oil level too low
d. Viscosity of liquid greater than that for which overheat
pump designed b. Improper or poor grade of oil

e. Wrong rotation c. Dirt in bearings

f. Mechanical defects: d. Dirt in oil

(1) Wearing rings worn e. Moisture in oil

(2) Impeller damaged f. Oil cooler clogged or scaled

(3) Internal leakage resulting g. Failure of oiling system

from defective gaskets h. Insufficient cooling water circulation
4. Pump loses a. Suction line not filled with liquid i. Insufficient cooling air
liquid after
starting b. Air leaks in suction line or stuffing boxes j. Bearings too tight
c. Gas or vapor in liquid k. Oil seals too close fit on shaft
d. Air or vapor pockets in suction line l. Misalignment
e. Inlet to suction pipe not sufficiently submerged 9. Bearings a. Misalignment
f. Available NPSH not sufficient wear rapidly
b. Shaft bent
g. Liquid seal piping to lantern ring plugged c. Vibration
h. Lantern ring not properly located in stuffing box d. Excessive thrust resulting from mechanical
failure inside the pump
e. Lack of lubrication
f. Bearings improperly installed
g. Dirt in bearings
h. Moisture in oil
j. Excessive or insufficient cooling of bearings

12-15
 FIG. 12-17
Adjustable Speed Drives3 and Power Transmissions

Type Characteristics Applications

Electric Drivers
Solid State AC drives high efficiency 35 to 2000 kW
good speed regulation l arger pumps where good speed regulation over not too
wide a range is required
low maintenance
complex controls hazardous areas
high cost
can be explosion proof
can retrofit
Solid State DC drives similar to AC except speed regulation good over a 35 to 350 kW
wider range non-hazardous areas
Electromechanical
Eddy Current Clutch efficient, proportional to slip 35 to 350 kW
poor speed regulation smaller centrifugal pumps where speed is usually near
design
require cooling non-hazardous areas
Wound-Rotor Motor poor speed regulation 35 to 350 kW
reasonable efficiency larger pumps non-hazardous areas
Mechanical
Rubber Belt wide range of speed regulation possible fractional to 75 kW
small centrifugal and positive displacement pumps
Metal Chain low to medium efficiency chemical feed pumps
non-hazardous areas
Hydraulic medium efficiency available hydraulic head
Power Recovery continuously variable speed
Turbines reversible use as pump

feeds into the pump suction. Whatever the cause, vortexes can Protection may be considered for the pump driver and may
be eliminated with a straightening cross, also called a vortex be combined with pump protections.
breaker, installed at the vessel outlet nozzle.
Installation, Operation, Maintenance
For discharge piping, sizing is determined by the available
head and economic considerations. Velocities range from 1 to 5 Installation, operation, and maintenance manuals should
m/s. A check valve should be installed between the discharge be provided by the pump manufacturer and are usually appli-
nozzle and the block valve to prevent backflow. cation specific. See Fig. 12-16 for a checklist of pump troubles
and causes.
Auxiliary piping (cooling, seal flushing and lubrication) is a
relatively inexpensive but extremely important item. API Stan- Driver rotation and alignment should be checked before the
dard 610, Centrifugal Pumps for General Refinery Service, or pump is operated.
applicable national standard should be followed. Provisions for
A typical starting sequence for a centrifugal pump is:
piping of stuffing box leakage and other drainage away from the
pump should be provided. Ensure that all valves in auxiliary sealing, cooling, and
flushing system piping are open, and that these systems
Pump Protection are functioning properly.
The following protection may be considered: Close discharge valve.
low suction pressure Open suction valve.
high discharge pressure
Vent gas from the pump and associated piping.
low suction vessel (or tank) level
Energize the driver.
high discharge vessel (or tank) level
Open discharge valve slowly so that the flow increases
low flow gradually.
flow reversal
Note that, on larger multistage pumps, it is very impor-
high temperature of bearings, case, etc. tant that flow through the pump is established in a mat-
vibrations ter of seconds. This is frequently accomplished by the
previously mentioned minimum flow recirculation.
lack of lubrication
overspeed

12-16
 RECIPROCATING PUMPS The relationship of overall suction and discharge volumetric
efficiency, displacement, and suction and discharge flow rate of
The most common reciprocating pump in gas plants is the a reciprocating pump is defined in Equation 12-12. When the
single-acting plunger pump which is generally employed in ser- leakage is not considered, the overall efficiencies may be substi-
vices with moderate capacity and high differential pressure. tuted by the density change efficiencies.
These pumps fill on the backstroke and exhaust on the forward
stroke. They are available with single (simplex) or multi-plung- Qs Qd
D= = Eq 12-12
ers (duplex, triplex, etc.), operating either horizontally or verti- VEsov VEdov
cally. Examples of plunger pump service in gas plants are: high
pressure chemical or water injection, glycol circulation, and low The following equations are based on the discharge flow
capacity, high pressure amine circulation, and pipeline product rate. Similar equations may be written for the suction side, and
pumps. conversions may be made by multiplying them by the discharge
to suction densities ratio.
Double-acting piston pumps which fill and exhaust on the
same stroke have the advantage of operating at low speeds and The overall discharge volumetric efficiency is a combination
can pump high viscosity liquids which are difficult to handle of volumetric efficiency due to leakage and discharge volumet-
with normal centrifugal or higher speed plunger pumps. ric efficiency due to fluid density change.
VEdov = VEl VEd Eq 12-13
Pump Calculations
The volumetric efficiency due to leakage is related to slip
Power requirement bkW: see equation in Fig. 12-2. as follows:
Displacement for single-acting pump VEl = 1 s Eq 12-14
D = 60 109 A m Ls n Eq 12-10 The effect of the difference in the leakage flow rate mea-
sured at suction pressure vs discharge pressure is neglected
Displacement for double-acting pump here, assuming that all leakages are internal.
D = 60 109 (2 A a) m Ls n Eq 12-11 The discharge volumetric efficiency due to density change is:
i
Notes: VEd = 1 r 1
Eq 12-15
o

1. Actual capacity (Q) delivered by pump is calculated by When the change in fluid density is linear with the change
multiplying displacement by the volumetric efficiency. in pressure and is smaller than 10%, and the temperature
change is negligible, Equation 12-16 may be used to calculate
2. The combination of mechanical and volumetric efficiency
hydraulic power. Hc comes from Equation 12-5. Additionally,
for reciprocating pumps is normally 90% or higher for
approximately 2 to 5% of power may be required for the work
noncompressible fluids.
done during the piston cycle, in compressing and in decompress-
3. In double-acting pumps with guided piston (rod in both ing the fluid that is held in the pump chamber without flowing
sides), change a to 2a in Equation 12-11. through the pump.
Qd o g Hc
hyd kW = Eq 12-16
Example 12-2 Calculate the power required for a simplex 3 600 000
plunger pump delivering 2.3m3/h of liquid of any relative densi-
ty at 20000kPa differential pressure and mechanical efficiency When the differential pressure is sufficiently high to cause a
of 90%. density change of more than 10%, or when the pressure is near
the fluids critical pressure, or when temperature change is not
(2.3) (20 000) negligible, this equation may not be accurate. In such cases the
bkW = = 14.2 kW
(3600) (0.90) pump manufacturer should be consulted. See Equipment and
System Equations last paragraph.
Volumetric Efficiency, Compressible Fluids Unlike wa-
ter, lighter hydrocarbon liquids (e.g. ethane, propane, butane) Data on density change with pressure and temperature can
are sufficiently compressible to affect the performance of recip- be found in Section 23, Physical Properties.
rocating pumps.
Example 12-3 For a 75mm diameter and a 125 mm stroke
The theoretical flow capacity is never achieved in practice triplex plunger pump pumping propane with a suction density
because of leakage through piston packing, stuffing boxes, or 505kg/m3 and a discharge density 525kg/m3 and given that r =
valves and because of changes in fluid density when pumping 4.6 and s = 0.03, find the overall discharge volumetric efficiency.
compressible fluids such as light hydrocarbons.
Discharge volumetric efficiency due to density change:
The ratio of real flow rate to theoretical flow rate (pump
505
525
displacement) is the volumetric efficiency. The volumetric ef- VEd = 1 4.6 1 = 0.824
ficiency depends on the size, seals, valves and internal config-
uration of each pump, the fluid characteristics and operating Volumetric efficiency due to leakage
conditions.
VEl = 1 0.03 = 0.97
When pumping compressible liquids, the volumetric efficien-
cy should be stated with reference to the flow rate measured in Overall discharge volumetric efficiency:
a specific side of the pump (suction or discharge side).
VEdov = (0.824) (0.97) = 0.799

12-17
Suction System Considerations Acceleration Head in 4" Pipe
The suction piping is a critical part of any reciprocating (1.2) (0.56) (360) (0.066)
ha4 = = 1.085 m
pump installation. The suction line should be as short as pos- (1.5) (9.8067)
sible and sized to provide not more than three feet per second
fluid velocity, with a minimum of bends and fittings. A centrifu- Acceleration Head in 6" Pipe
gal booster pump is often used ahead of a reciprocating pump to (6.1) (0.25) (360) (0.066)
provide adequate NPSH which would also allow higher suction ha6 = = 2.463 m
line velocities. (1.5) (9.8067)

NPSH required for a reciprocating pump is calculated in the Total Acceleration Head
same manner as for a centrifugal pump, except that additional ha = 1.085 + 2.463 = 3.548 m
allowance must be made for the requirements of the reciprocat-
ing action of the pump. The additional requirement is termed Karassik et al9 recommend that the NPSHA exceed the
acceleration head. This is the head required to accelerate the NPSHR by 20 to 35kPa for reciprocating pumps.
fluid column on each suction stroke so that this column will, at
Pulsation A pulsation dampener (suction stabilizer) is a
a minimum, catch up with the receding face of the piston during
device installed in the suction piping as close as possible to the
its filling stroke.
pump to reduce pressure fluctuations at the pump. It consists of
Acceleration Head Acceleration head is the fluctua- a small pressure vessel containing a cushion of gas (sometimes
tion of the suction head above and below the average due to separated from the pumped fluid by a diaphragm or bladder).
the inertia effect of the fluid mass in the suction line. With the Pulsation dampeners should be considered for the suction side
higher speed of present-day pumps or with relatively long suc- of any reciprocating pump, but they may not be required if the
tion lines, this pressure fluctuation or acceleration head must suction piping is oversized and short, or if the pump operates
be taken into account if the pump is to fill properly without at less than 150 rpm. A properly installed and maintained pul-
forming vapor which will cause pounding or vibration of the sation dampener should absorb the cyclical flow variations so
suction line. that the pressure fluctuations are about the same as those that
occur when the suction piping is less than 4.5m long.
With the slider-crank drive of a reciprocating pump, maxi-
mum plunger acceleration occurs at the start and end of each Similar pressure fluctuations occur on the discharge side of
stroke. The head required to accelerate the fluid column (ha) is every reciprocating pump. Pulsation dampeners are also effec-
a function of the length of the suction line and average velocity tive in absorbing flow variations on the discharge side of the
in this line, the number of strokes per minute (rpm), the type of pump and should be considered if piping vibration caused by
pump and the relative elasticity of the fluid and the pipe, and pressure fluctuations appears to be a problem. Pulsation damp-
may be calculated as follows: ener manufacturers have computer programs to analyze this
phenomenon and should be consulted for reciprocating pump
L v n C applications over 35kW. Discharge pulsation dampeners mini-
ha = Eq 12-17
kg mize pressure peaks and contribute to longer pump and pump
valve life. The need for pulsation dampeners is increased if mul-
where C and k are given in Fig. 12-18.
tiple pump installations are involved.
Example 12-4 Calculate the acceleration head, given a 50
Ensure that gas-cushion type pulsation dampeners contain
mm diameter 125mm stroke triplex pump running at 360
the correct amount of gas. The following equation may be used
rpm and displacing 16.5m3/h of water with a suction pipe made
for sizing estimation of bladder and diaphragm-type pulsation
up of 1.2m of 4" and 6.1m of 6" standard wall pipe.
dampeners, where the volume, length and area must be in self-
Average Velocity in 4" Pipe = 0.56m/s consistent units.
A Ls K (100/(100 ) ) N
1
Average Velocity in 6" Pipe = 0.25m/s
Vpd = 1 =
1 (100/(100 + ) ) N

FIG. 12-18
Reciprocating Pump Acceleration Head Factors

C = 0.200 for simplex double-acting k = a factor related to the fluid compressibility

= 0.200 for duplex single-acting hot oil 2.5
= 0.115 for duplex double-acting most hydrocarbons 2.0
= 0.066 for triplex single or double-acting amine, glycol, water 1.5
= 0.040 for quintuplex single or double-acting deaerated water 1.4
= 0.028 for septuplex single or double-acting liquid with small amounts of entrained gas 1.0
= 0.022 for nonuplex single or double-acting
Note: C will vary from the listed values for unusual ratios of connecting rod length to crank radius over 6.

12-18
 A Ls K (Pave/Pmin) N
1
Eq 12-18 culated back to a lower point on the impeller vanes; thus there
are two fluid helical paths around the impeller and chamber,
1 (Pave/Pmax) N
1

recirculating the fluid from vane to vane, from the suction to
the discharge ports, on both sides of the impeller. The recircula-
Where K has a value of: tion increases the head developed in each stage, so the head is
Single Acting Double Acting
a function of the number of recirculation cycles. Capacity, head
and power, and speed follow fan laws.
Simplex 0.60 0.25
Typically, the performance curve is a downward slope
Duplex 0.25 0.15
straight line; therefore, a throttling valve in a regenerative
Triplex 0.13 0.06 pump will permit more precise changes in flow than in centrifu-
Quadruplex 0.10 0.06 gal pumps. The maximum shut-off head developed may be up
to 6 times the shut-off head of a single stage centrifugal pump
Quintuplex 0.06 0.02 running at the same speed.
Because of close clearances, regenerative pumps can not be
Capacity Control Manual or automatic capacity control
used to pump liquids containing solid particles. They can pump
for one pump or several parallel pumps can be achieved by one
liquids containing vapors and gases, if they contain sufficient
or a combination of the following methods:
liquid to seal the close clearances.
on-off control
recirculation DIAPHRAGM PUMPS
variable speed driver or transmission Diaphragm pumps are reciprocating, positive displacement
type pumps, utilizing a valving system similar to a plunger
variable displacement pump  pump. These pumps can deliver a small, precisely controlled
amount of liquid at a moderate to very high discharge pressure.
Drivers Two types of mechanisms are commonly used for Diaphragm pumps are commonly used as chemical injection
driving reciprocating pumps; one in which the power of a motor pumps because of their controllable metering capability, the
or engine is transmitted to a shaft and there is a mechanism to wide range of materials in which they can be fabricated, and
convert its rotative movement to alternating linear movement their inherent leakproof design.
to drive the pumping piston or plunger. In the other type, there
is a power fluid, such as steam, compressed air, or gas acting
on a piston, diaphragm or bellow linked to the pumping piston MULTIPHASE PUMPS
or plunger. Multiphase pumps can pump immiscible liquids such as oil
Piping Suction and discharge piping considerations are and water with gas. There are screw types and rotodynamic
similar to those for centrifugal pumps. In addition, acceleration types. A progressive cavity design is used along the flow path
head must be included for pipe sizing. For piping materials and to accommodate gas volume reduction caused by increased
thickness selection, pressure pulsations amplitude and fatigue pressure. A full range of gas/liquid ratios can be handled. This
life should be considered. class of pumps is of interest in applications where convention-
al pumps and separate compressors with or without separate
pipelines are not economically feasible.
ROTARY PUMPS
The rotary pump is a positive displacement type that de- LOW TEMPERATURE PUMPS
pends on the close clearance between both rotating and station-
ary surfaces to seal the discharge from the suction. The most Two types of centrifugal pumps have been developed for
common types of rotary pumps use gear or screw rotating ele- cryogenic applications: the external motor type and the sub-
ments. These types of positive displacement pumps are com- merged motor type.
monly used for viscous liquids for which centrifugal or recipro-
cating pumps are not suitable. Low viscosity liquids with poor External Motor Type
lubricating properties (such as water) are not a proper applica- These pumps are of conventional configuration with a cou-
tion for gear or screw pumps. pled driver and can be single or multi-stage. The pump assem-
bly is usually mounted in the vessel from which it pumps and
REGENERATIVE PUMPS the motor is mounted externally.
Regenerative pumps are also called peripheral pumps. The Submerged Motor Type
unit has a rotary wheel or impeller with vanes on both sides of
its periphery, which rotates in an annular shaped chamber in This type of pump is characterized by being directly coupled
the pump casing. The fluid moves outwards through the vanes, to its motor, with the complete unit being submerged in the
at the vanes tips the fluid passes to the chamber and is recir- fluid.

12-19
 Hydraulic Turbines
Many industrial processes involve liquid streams which flow Applications
from higher to lower pressures. Usually the flow is controlled
with a throttling valve, hence the hydraulic energy is wasted. HPRTs may be used to drive any kind of rotating equipment
Up to 80% of this energy can be recovered by passing the liquid (e.g. pumps, compressors, fans, electrical generators). The main
through a hydraulic power recovery turbine (HPRT). To justify problems are matching the power required by the driven load
the installation of an HPRT, an economic analysis of the power to that available from the HPRT and speed control. Both the
savings versus added equipment and installation costs should power produced and the speed can be controlled by:
be performed. throttling the liquid flow, either downstream or upstream
from the HPRT
TYPES OF HPRTs
allowing a portion of the liquid to bypass the HPRT
Two major types of centrifugal hydraulic power recovery
turbines are used. adjusting inlet guide vanes installed in the HPRT

1. Reaction Single or multistage Francis-type rotor with Sometimes HPRTs are installed with a helper driver. If
fixed or variable guide vanes. this is an electric motor, the speed will be controlled by the mo-
tor speed.
2. Impulse Pelton Wheel, usually specified for relatively
high differential pressures. Typical gas-processing streams for which HPRTs should be
considered are:
HPRTs with Francis-type rotors (inward-flow reaction tur-
bine) are similar to centrifugal pumps. In fact, a good centrifu- Rich sweetening solvents (e.g. amines, etc.)
gal pump can be expected to operate with high efficiency as an Rich absorption oil
HPRT when the direction of flow is reversed.
High-pressure crude oil.
The Pelton Wheel (impulse runner type HPRT) is used in
high head applications. The impulse type turbine has a nozzle Condensed high pressure natual gas liquids.
which directs the high pressure fluid against bowl-shaped buck- Liquid refrigerant letdown, in mechanical refrigeration
ets on the impulse wheel. This type of turbines performance is cycles.
dependent upon back pressure, while the reaction type is less
dependent upon back pressure. High pressure LNG letdown, in natural gas liquefaction.

Power Recovered by HPRTs The lower limit of the power recovery which can be economi-
cally justified with single-stage HPRTs is about 20kW and with
The theoretical energy which can be extracted from a high multistage, about 75kW. HPRTs usually pay out their capital
pressure liquid stream by dropping it to a lower pressure cost in from one to three years.
through an HPRT can be calculated using the hydraulic power.
See Fig. 12-2 for bkW equation. Since some of the energy will be Frequently, when an HPRT is to be used to drive a pump,
lost because of friction, the hydraulic power must be multiplied both devices are purchased from one manufacturer. This has
by the efficiency of the HPRT. the advantage of ensuring that the responsibility for the entire
installation is assumed by a single supplier.
In applications where the fluid that enters the HPRT has
large dissolved gas content, the available power is larger than The available pressure differential across the HPRT is cal-
the power that may be calculated using the liquid equations, culated using a technique similar to that used to calculate the
so, the power shall be calculated using an adequate two-phase differential head of centrifugal pumps.
calculation method. Example 12-5 Specify an HPRT driven pump for a gas sweet-
The amount of power recovered by an HPRT is directly pro- ening process.
portional to the efficiency rather than inversely proportional Given:
as is the case when calculating the power required by a pump.
Thus, if a fluid is pumped to a high pressure and then reduced lean DEA flow 227m3/h
to its original pressure using an HPRT, the proportion of the
lean DEA temperature 43C
pumping energy which can be supplied by the HPRT is equal
to the efficiencies of the pump and turbine multiplied together. lean DEA relative density 1.00
Typically, good centrifugal pumps and good HPRTs have effi-
ciencies of between 70% and 80%. Thus, the HPRT can be ex- lean DEA vapor pressure at 49C 11.7kPa(abs)
pected to provide between 50% and 60% of the energy required rich DEA flow 227m3/h
for pumping.
rich DEA temperature 71C
Usually the high-pressure liquid contains a substantial
amount of dissolved gas. The gas comes out of solution as the rich DEA relative density 1.01
liquid pressure drops. This does not cause damage to the HPRT,
pump suction total pressure 517kPa(ga)
presumably because the fluid velocity through the HPRT is high
enough to maintain a froth-flow regime. The term NPSHR does pump discharge total pressure 6791kPa(ga)
not apply to HPRTs.

12-20
 HPRT inlet total pressure 6619kPa(ga) (227) (640.4) (1.00)
bkW for pump = = 504.6 kW
HPRT outlet total pressure 586kPa(ga) (367) (0.785)

Solution: Available head

(6619 586)
for HPRT = = 609.7 m
For this example, the suction and discharge pressures have (9.807) (1.01) (0.999)
already been calculated using a technique similar to that sug-
gested for centrifugal pumps. The HPRT selected is a 3-stage unit. From the performance
curve (Fig. 12-19), the expected efficiency of the HPRT is 76%.
(517 + 101.3 11.7) Hence, the available power will be:
NPSHA for pump = = 61.9 m
(9.807) (1.00) (0.999) (227) (609.7) (1.01) (0.76)
bkW from HPRT = = 289.5 kW
367
(6791 517)
Required head for pump = = 640.4 m Another driver, such as an electric motor, would be required
(9.807) (1.00) (0.999) for the pump to make up the difference in bkW between the
The pump selected is a 5-stage unit. From the pump curve pump and HPRT. The other driver would have to be capable
(Fig. 12-20, the expected efficiency of the pump is 78.5%. Hence, of providing at least 215kW. It is good practice to provide an
the required power will be: electric motor driver large enough to drive the pump by itself to
facilitate startups. The pump, HPRT, and electric motor driver

FIG. 12-19
Rich DEA Pressure Letdown

FIG. 12-20
Lean Amine Charge Pump

12-21
(helper or full size) would usually be direct connected. In some REFERENCES
cases, a clutch is used between the pump and HPRT, so the unit
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leum Institute, International Standards Organization.
The pump and HPRT are similar in hydraulic design except
2. Bingham-Willamette Ltd., Sales Manual, Burnaby, B.C., Cana-
that the pump has five stages and the HPRT, three stages. In da.
this case, the HPRT is a centrifugal pump running backwards.
3. Doll, T. R., Making the Proper Choice of Adjustable-speed
Drives. Chem. Eng., v. 89, no. 16, August 9, 1982.
CODES & ORGANIZATIONS
4. Evans, F. L., Jr., Equipment Design Handbook for Refineries
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Displacement Pumps. Chem. Engr., v. 92, no. 15, July 22, 1985,
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p. 38-52.
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ULCUnderwriters Laboratory of Canada Denver, 1988.

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