Pump sizing basics

38 www.aiche.org/cep December 2016 CEP
Back to Basics
W
hen I left university, I found that I needed addi-
tional information to turn my theoretical knowl-
edge of fluid mechanics into the practical knowl-
edge required to specify a pump. Judging by the questions I
see asked nearly every week on LinkedIn and elsewhere, I
believe this is a problem shared by many engineers early in
their careers. This article gives practical insight on how to
specify a pump.
Pump types
Pumps can be used to move fluids, which flow from
regions of high pressure to regions of low pressure, by
increasing the pressure of the fluid. Before you purchase
a pump, you must specify the type of pump and make
sure it is capable of delivering a given flowrate at a given
pressure.
There are two main pump types: rotodynamic and
positive-displacement. In a rotodynamic pump, a rotating
impeller imparts energy to the fluid. The most common type
of rotodynamic pump is the centrifugal pump (Figure 1).
The amount of liquid that passes through the pump is
inversely proportional to the pressure at the pump outlet.
In other words, the outlet flowrate of a rotodynamic pump
varies nonlinearly with pressure.
In a positive-displacement (PD) pump, a discrete amount
of fluid is trapped, forced through the pump, and discharged.
A gear pump is an example of a PD pump (Figure 2). This
pumping principle produces a pulsating flow, rather than a
smooth flow. Its output flow tends to vary little with respect
to the pressure at the pump outlet, because the moving
displacement mechanism pushes the slug of liquid out at a
constant rate.
Most process pumps are rotodynamic pumps, so you
need to know the required outlet pressure to specify the
pump that will provide the required flow. Alhough cer-
tain system head parameters are calculated the same way
whether the driving force for flow is a pump or gravity, this
article mainly addresses sizing concerns for rotodynamic
pumps.
Pump sizing
Pump sizing involves matching the flow and pressure
rating of a pump with the flowrate and pressure required for
the process. The mass flowrate of the system is established
on the process flow diagram by the mass balance. Achiev-
ing this mass flowrate requires a pump that can generate a
pressure high enough to overcome the hydraulic resistance
of the system of pipes, valves, and so on that the liquid must
travel through. This hydraulic resistance is known as the
system head.
In other words, the system head is the amount of
pressure required to achieve a given flowrate in the system
downstream of the pump. The system head is not a fixed
quantity — the faster the liquid flows, the higher the system
head becomes (for reasons to be discussed later). However,
a curve, known as the system curve, can be drawn to show
This article explains some of the
core concepts behind pump sizing.
Seán Moran
Expertise Ltd.
Pump Sizing:
Bridging the Gap Between
Theory and Practice
Copyright © 2016 American Institute of Chemical Engineers (AIChE)

CEP December 2016 www.aiche.org/cep 39
the relationship between flow and hydraulic resistance for a
given system.
Pump sizing, then, is the specification of the required
outlet pressure of a rotodynamic pump (whose output flow
varies nonlinearly with pressure) with a given system head
(which varies nonlinearly with flow).
Understanding system head
The system head depends on properties of the system the
pump is connected to — these include the static head and the
dynamic head of the system.
The static head is created by any vertical columns of
liquid attached to the pump and any pressurized systems
attached to the pump outlet. The static head exists under
static conditions, with the pump switched off, and does
not change based on flow. The height of fluid above the
pump’s centerline can be determined from the plant layout
drawing.
The dynamic head varies dynamically with flowrate (and
also with the degree of opening of valves). The dynamic
head represents the inefficiency of the system — losses of
energy as a result of friction within pipes and fittings and
changes of direction. This ineffiency increases with the
square of the average velocity of the fluid.
Dynamic head can be further split into two parts. The
frictional loss as the liquid moves along lengths of straight
pipe is called the straight-run headloss, and the loss as a
result of fluid passing through pipe fittings such as bends,
valves, and so on is called the fittings headloss.
Fully characterizing a hydraulic system is incredibly
complex. Remember that in order to specify a pump, you
only need to characterize the system well enough to choose
a pump that will perform the job in question. How exact you
need to be depends on where in the design process you are.
If you are at the conceptual stage, you may be able to avoid
specifying the pump at all, but experience suggests that
you should use rules of thumb to specify certain parameters
(such as superficial velocity) to prevent difficulties later. I
also recommend designing the process so that it does not
have two-phase flow. Two-phase flow is difficult to predict,
and should be avoided in your design if at all possible —
head losses can be one thousand times those for single-phase
flow. Installing knock-out drums in the system and arranging
pipework so that gases are not entrained in liquids can help
mitigate two-phase flow.
Superficial velocity is the same as average velocity and
is the volumetric flowrate (in m3/sec, for example) divided
by the pipe’s internal cross-sectional area (e.g., in m2). A
very quick way to start the hydraulic calculations is to use
the following superficial velocities:
• pumped water-like fluids: <1.5 m/sec
• gravity-fed water-like fluids: <1 m/sec
• water-like fluids with settleable solids: >1, <1.5 m/sec
• air-like gases: 20 m/sec
Keeping the system within these acceptable ranges of
superficial velocities, and avoiding two-phase flow, will
typically produce sensible headlosses for the pipe lengths
usually found in process plants.
Determining frictional losses through fittings
Dynamic, or friction, head is equal to the sum of the
straight-run headloss and the fittings headloss.
The fittings headloss is calculated by what is known as
the k-value method. Each type of valve, bend, and tee has
Discharge
Suction
Impeller
Eye
Impeller
p Figure 1. In a centrifugal pump, a rotating impeller imparts energy to
the liquid moving through the pump.
Outlet
Inlet
Low-Pressure
Fluid
High-Pressure
Fluid
Fluid Carried
Between Teeth
and Case
p Figure 2. A gear pump is a type of positive-displacement pump in which
a discrete volume of fluid is trapped and then discharged.

Back to Basics
a characteristic resistance coefficient, or k value, which
can be found in Perry’s Handbook (1) and other sources
(Table 1) (2).
To use this method, count the number of valves on the
piping and instrumentation diagram (P&ID), and the fittings,
bends, and tees on the plant layout drawing for the relevant
suction or delivery line. Multiply the number of each type of
fitting by the corresponding k value, and add the k values for
the various types of fittings to get the total k value. Use the
total k value to calculate the headloss due to fittings:
where hf is the fittings headloss in meters water gauge (mwg),
k is the total k value, v is the superficial velocity (m/sec), and
g is the acceleration due to gravity (9.81 m/sec2).
Calculating straight-run headloss
At a more-advanced stage of design, you might want
to know a pump’s physical size to try out on a plant layout
drawing. An easy way to determine the straight-run head-
loss — the most difficult part of a headloss calculation — is
to use a nomogram such as Figure 3 or a table. Pipe manu-
facturers (and others) produce tables and nomograms that
can be used to quickly look up headloss due to friction for
liquids.
To use the nomogram, use a ruler to draw a straight line
through any pair of known quantities to determine unknown
quantities. For example, for a 25-mm nominal-bore pipe
with a flow velocity of 1 m/sec, the straight-run headloss is
about 6 m per 100 m of pipe. So the headloss through 10 m
of this pipe is around 0.6 mwg.
At an early design stage, you often need to calculate the
straight-run headloss multiple times. Rather than referring to
a table or nomogram numerous times, it can be quicker to set
up an Excel spreadsheet and use a formula to calculate the
Darcy friction factor and headloss.
Chemical engineering students are usually taught to find
the Darcy friction factor using a Moody diagram, which is a
summary of a large number of empirical experiments. You
can use curve-
fitting equations and software such as Excel to
approximate the Moody diagram’s output.
Don’t confuse the Darcy friction factor with the Fanning
friction factor — the Darcy friction factor is by definition
four times the Fanning friction factor. If you do decide to
use a Moody diagram to find the friction factor, be aware of
which friction factor is on the y-axis.
I prefer the Colebrook-White approximation to calculate
the Darcy friction factor. Although it is an approximation, it
Table 1. Each type of pipe fitting has a resistance
coefficient, or k value, that can be used to calculate
the fittings headloss for the pump system (2).
Fitting Type k Value
Short-radius bends, for every 22.5 deg. allow 0.2
Long-radius bends, for every 22.5 deg. allow 0.1
Open isolation valve 0.4
Open control valve 10.8
Tee (flow from side branch) 1.2
Tee (flow straight-through) 0.1
Swing check non-return valve 1
Sharp entry 0.5
Internal
Diameter,
mm
Flowrate,
L/sec L/min
Flow
Velocity,
m/sec
Pressure
Drop,
m/100m
15
20
25
30
35
40
50
60
70
80
90
100
150
200
250
300
350
400
500
0.01
0.02
0.05
0.1
0.2
0.3
0.5
1
2
3
4
5
20
30
40
50
100
200
300
400
500
1000
2000
3000
4000
5000
0.4
1
2
3
4
5
20
30
40
50
100
200
300
400
500
1000
2000
3000
4000
5000
10
10000
20000
30000
40000
50000
100
200
300
m3
/min
0.05
0.1
0.15
0.2
0.3
0.5
0.4
1
2
3
4
5
1.5
10
15
20
20
10
2
3
4
5
1
0.2
0.3
0.5
0.4
0.1
0.05
0.04
0.03
0.02
0.01
Approximate values only
Water at 10°C
p Figure 3. A piping nomogram, available from pipe manufacturers,
can be used to estimate the straight-run headloss for a pump system.
In the example shown by the red line, a 25-mm pipe with a flow velocity
of 1 m/sec has a straight-run headloss of about 6 m per 100 m of pipe.
Copyright image reproduced courtesy of Durapipe SuperFLO ABS
technical data.

might be closer to the true experimental value than what the
average person can read from a Moody diagram.
The Colebrook-White approximation can be used to esti-
mate the Darcy friction factor (fD) from Reynolds numbers
greater than 4,000:
where Dh is the hydraulic diameter of the pipe, ε is the sur-
face roughness of the pipe, and Re is the Reynolds number:
where ρ is the density of the fluid, D is the pipe internal
diameter, and μ is the fluid dynamic viscosity.
The Colebrook-White approximation can be used itera-
tively to solve for the Darcy friction factor. The Goal Seek
function in Excel does this quickly and easily.
The Darcy-Weisbach equation states that for a pipe
of uniform diameter, the pressure loss due to viscous
effects (Δp) is proportional to length (L) and can be charac-
terized by:
This iterative approach allows you to calculate straight-
run headloss to the degree of accuracy required for virtually
any practical application.
I recently came across a paper (3) that suggested there
are other equations that provide more accurate results
through curve-fitting than the Colebrook-White approxi-
mation. If you are producing your own spreadsheet for this
purpose, I suggest you look into the Zigrang and Sylvester
(4) or Haaland equations (5) (Table 2). These equations also
apply for Reynolds numbers greater than 4,000.
Adding together the static head, the fittings headloss,
and the straight-run headloss will give you the total head the
pump needs to generate to overcome resistance and deliver
the specified flowrate to the system.
Suction head and net positive suction head
Even at an early stage, I also recommend determining
the pump’s required net positive suction head and calculat-
ing the net positive suction head (NPSH), as they can affect
much more than pump specification. The pump’s required
net positive suction head takes into consideration the liquid’s
vapor pressure to avoid cavitation in the pump.
I recommend creating an Excel spreadsheet that uses the
Antoine equation to estimate the vapor pressure of the liquid
at the pump inlet and then calculate the NPSH at that vapor
pressure.
The Antoine equation may be expressed as:
where Pv is vapor pressure of the liquid at the pump inlet,
T is temperature, and A, B, and C are coefficients that can be
obtained from the NIST database (http://webbook.nist.gov)
among other places. Table 3 shows an example for water.
The net positive suction head is:
where Po is the absolute pressure at the suction reservoir,
ho is the reservoir liquid level relative to the pump center-
line, and hSf is the headloss due to friction on the suction side
of the pump.
Note that NPSH is calculated differently for centrifugal
and positive-displacement pumps, and that it varies with
pump speed for positive-displacement pumps rather than
with pressure as for centrifugal pumps. Equation 6 should
only be used with centrifugal pumps.
Article continues on next page
Table 3. Vapor pressure for water at 30°C, calculated using the Antoine equation.
Material A B C T, °C T, K Pv, bar Pv, Pa
Water 5.40221 1,838.675 –31.737 30 303.15 0.042438 4,243.81
Table 2. These alternative curve-fitting equations can be used
in lieu of the Colebrook-White equation to determine the Darcy friction factor.
Equation Range Source
e = 0.00004–0.05 (4)
e = 0.000001–0.05 (5)

Back to Basics
Determining pump power
After the system head has been calculated, it can be used
to calculate an approximate pump power rating for a centrif-
ugal pump:
where P is the pump power (kW), Q is the flowrate (m3/hr),
H is the total pump head (m of fluid), and η is the pump
efficiency (if you do not know the efficiency, use η = 0.7).
The pump manufacturer provides the precise power rat-
ings and motor size for the pump, but the electrical engineers
need an approximate value of this (and pump location) early
in the design process to allow them to size the power cables.
You should err on the side of caution in this rating calculation
(the electrical engineers will be much happier if you come
back later to ask for a lower power rating than a higher one).
In certain stages of design development, the preliminary
drawings are modified to match likely hydraulic conditions
across the design envelope. This may require you to do
many approximate hydraulic calculations before the design
has settled into a plausible form.
After you have performed the hydraulic calculations, the
pump and possibly the pipe sizes might need to be changed,
as might the minimum and maximum operating pressures at
certain points in the system. As the system design becomes
more refined, there might even be a requirement to change
from one pump type to another.
Hydraulic networks
The previous sections describe how to calculate the
headloss through a single line, but what about the common
situation where the process has branched lines, manifolds,
and so on? When each branch handles a flow proportional
to its headloss, and its headloss is proportional to the flow
passing through it, producing an accurate model can become
complex very quickly.
My approach to this is to first simplify and then improve
the design as much as possible with a few rules of thumb:
• Avoid manifold arrangements that provide a straight-
through path from the feed line to a branch. Entry perpendic-
ular to branch direction is preferred.
• Size manifolds such that the superficial velocity never
exceeds 1 m/sec at the highest anticipated flowrate.
• Specify progressively smaller manifold diameters to
accommodate lower flows to downstream branches.
• Include a small hydraulic restriction in the branch so
the branch headloss is 10–100 times the headloss across the
manifold.
• Design-in passive flow equalization throughout the pip-
ing system wherever possible by making branches hydrauli-
cally equivalent.
Perform headloss calculations for each section of the
simplified plant design at expected flows to find the flow
path with the highest headloss. Use the highest-headloss
path to determine the required pump duty — calculate the
pump duty at both the average flow with working flow
equalization, and at full flow through a single branch. Usu-
ally these do not differ much, and the more rigorous answer
lies between them. Only if the two results of this approach
are very different will I do a more rigorous (and time-
consuming) analysis.
If such a rigorous analysis is needed, I create an Excel
spreadsheet based on the Hardy Cross method — a method
for determining the flow in a pipe network when the flows
within the network are unknown but the inputs and outputs
are known — and solve for individual pipe flows. Excel’s
Solver function can be used to find the change in flow that
gives zero loop headloss. In the unlikely event that you have
to do this, an explanation of how to carry out the method
can be found in Ref. 6. There are many computer programs
available to do these calculations.
Pump curves
A pump curve is a plot of outlet pressure as a function
of flow and is characteristic of a certain pump. The most
frequent use of pump curves is in the selection of centrifugal
pumps, as the flowrate of these pumps varies dramatically
with system pressure. Pump curves are used far less fre-
quently for positive-displacement pumps. A basic pump
curve plots the relationship between head and flow for a
pump (Figure 4).
On a typical pump curve, flowrate (Q) is on the horizon-
tal axis and head (H) is on the vertical axis. The pump curve
shows the measured relationship between these variables, so
Pump Curve
S
y
s
t
e
m
C
u
r
v
e
Duty Point
Head,
H,
m
Pressure,
psi
Flowrate, Q, L/sec
Flowrate, Q, m3
/hr
p Figure 4. A basic pump curve plots pressure (or head) as a function of
flowrate.

it is sometimes called a Q/H curve. The intersection of this
curve with the vertical axis corresponds to the closed valve
head of the pump. These curves are generated by the pump
manufacturer under shop test conditions and ideally repre-
sent average values for a representative sample of pumps.
A plot of the system head over a range of flowrates, from
zero to some value above the maximum required flow, is
called the system curve. To generate a system curve, com-
plete the system head calculations for a range of expected
process flowrates. System head can be plotted on the same
axes as the pump curve. The point at which the system curve
and the pump curve intersect is the operating point, or duty
point, of the pump.
Remember that a system curve applies to a range of
flows at a given system configuration. Throttling a valve in
the system will produce a different system curve. If flow
through the system will be controlled by opening and closing
valves, you need to generate a set of curves that represent
expected operating conditions, with a corresponding set of
duty points.
It is common to have efficiency, power, and NPSH
plotted on the same graph (Figure 5). Each of these variables
requires its own vertical axis. To obtain the pump efficiency
at the duty point, draw a line vertically from the duty point
to the efficiency curve, and then draw a horizontal line from
there to the vertical axis that corresponds to efficiency. Sim-
ilarly, to obtain the motor power requirement, draw a line
down from the duty point to the motor duty curve.
More sophisticated curves may include nested curves
representing the flow/head relationship at different supply
frequencies (i.e., the AC electrical supply’s frequency in Hz)
or rotational speeds, with different impellers, or for different
fluid densities. Curves for larger impellers or faster rotation
lie above curves for smaller impellers or slower rotation, and
curves for lower-density fluids lie above curves for higher-
density fluids. A more-advanced pump curve might also
incorporate impeller diameters and NPSH. Figure 6 depicts
pump curves for four different impellers, ranging from
222 mm to 260 mm. Corresponding power curves for each
impeller are shown on the bottom of the figure. The dashed
lines in Figure 6 are efficiency curves.
These curves can start to look a bit confusing, but the
important point to keep in mind is that, just as in the simpler
examples, flowrate is always on a common horizontal axis,
and the corresponding value on any curve is vertically above
or below the duty point.
These more-advanced curves usually incorporate effi-
ciency curves, and these curves define a region of highest
efficiency. At the center of this region is the best efficiency
point (BEP).
Choose a pump that has an acceptable efficiency across
the range of expected operating conditions. Note that we are
not necessarily concerned with the entire design envelope —
Pump Curve
Efficiency
System
Curve
Power Consumption
NPSH
Head,
m
0
10
20
30
40
50
60
Efficiency,
%
0
10
20
30
40
50
60
70
80
Power,
kW
0
2
4
6
8
10
12
NPSH,
m
0
2
4
6
8
10
12
Flowrate, m3
/hr
10 20 30 40 50 60 70 80
p Figure 5. Efficiency, power, and net positive suction head can also be
plotted on a pump curve. Original image courtesy of Grundfos.
Head,
m
80
70
60
50
40
30
20
0 20 40 60 80 100 120 140 160 180
0 20 40 60 80 100 120 140 160 180
Flowrate, m3
/hr
0
4
8
12
16
20
24
28
32
36
40
Power,
kW
0
2
4
6
8
10
12
NPSH,
m
NPSH
222 mm
235 mm
247 mm
260 mm
222 mm
235 mm
247 mm
260 mm
67.5%
66.8%
70.2%
71.7%
64%
67%
70%
64%
67%
70%
p Figure 6. A complex pump curve integrates efficiency, NPSH, and
impeller diameters on one diagram. Copyright image reproduced courtesy
of Grundfos.

Back to Basics
it is not crucial to have high efficiency across all conceivable
conditions, just the normal operating range.
The optimal pump for your application will have a BEP
close to the duty point. If the duty point is far to the right of
a pump curve, well away from the BEP, it is not the right
pump for the job.
Even with the most cooperative pump supplier, some-
times the curves that you need to make a pump selection
may not be available. This is commonly the case if you want
to use an inverter to control pump output based on speed.
However, you can often generate acceptable pump
curves using the curves you have and the following approxi-
mate pump affinity relationships:
where the subscript 1 designates an initial condition on a
known pump curve and subscript 2 is some new condition.
The NPSH relationship in Eq. 11 is more of an approx-
imation than the others. The value of x lies in the range of
–2.5 to +1.5, and y in the range of +1.5 to +2.5.
Closing thoughts
These are the basics of pump selection. A final word of
advice: If you don’t understand what is presented here, or
need to know more, I suggest that you talk to a pump sup-
plier in private. Think twice before you post on social media
to ask for advice on the basics of pump selection — the
advice you receive may not be correct, and your post may
reflect badly on you and your employer.
SEÁN MORAN has had 25 years of experience in process plant design,
troubleshooting, and commissioning. He was an associate professor
and Coordinator of Design Teaching at the Univ. of Nottingham for four
years, and is presently a visiting professor at the Univ. of Chester. He
has written three books on process plant design for the Institution of
Chemical Engineers. His professional practice now centers on acting
as an expert witness in commercial disputes regarding process plant
design issues, although he still has cause to put on a hardhat from time
to time. He holds a master’s degree in biochemical engineering from
Univ. College London.
Literature Cited
1. Perry, R. H., and Green, D. W., “Perry’s Chemical Engineers’
Handbook,” 8th Ed., McGraw-Hill, New York, NY, p. 6-18
(2007).
2. Moran, S., “An Applied Guide to Process and Plant Design,”
Butterworth-Heinemann Oxford, U.K. (2015).
3. Genić, S., et al., “A Review of Explicit Approximations of
Colebrook’s Equation,” FME Transactions, 39, pp. 67–71
(June 2011).
4. Zigrang, D. J., and N. D. Sylvester, “Explicit Approximations
to the Solution of Colebrook’s Friction Factor Equation,” AIChE
Journal, 28 (3), pp. 514–515 (May 1982).
5. Haaland, S. E., “Simple and Explicit Formulas for the Friction
Factor in Turbulent Flow,” Journal of Fluids Engineering,
105 (1), pp. 89–90 (1983).
6. Huddleston, D., et al., “A Spreadsheet Replacement for Hardy-
Cross Piping System Analysis in Undergraduate Hydraulics,”
Critical Transitions in Water and Environmental Resources
Management, pp. 1–8 (2004).
Nomenclature
A, B, C = Antoine coefficients
D = pipe internal diameter
Dh = hydraulic diameter of the pipe
fD = Darcy friction factor
g = acceleration due to gravity (9.81 m/sec2)
H = total system head
hf = headloss due to fittings in meters water gauge
(mwg)
ho = reservoir liquid level relative to the pump
centerline
hSf = headloss due to friction on the suction side of
the pump
k = resistance coefficient of valves, fittings, bends,
tees, etc.
L = length of pipe
NPSH = net positive suction head
P = power (kW)
Pv = vapor pressure of the liquid at the pump inlet
Po = absolute pressure at the suction reservoir
Q = flowrate
Re = Reynolds number
T = temperature
v = superficial velocity
Greek letters
Δp = pressure loss due to viscous effects
ε = surface roughness of the pipe
η = pump efficiency
μ = fluid dynamic viscosity (kg/(m-sec))
ρ = density of fluid (kg/m3)
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