The document provides details on the design process for a centrifugal pump given specific head, flow rate, and speed requirements provided by the client. Key steps include: 1) Calculating hydraulic parameters like flow rate, horsepower required, and shaft torque to size the shaft diameter. 2) Designing dimensions of the impeller like eye diameter, inlet and outlet angles, and widths to achieve the required flow while minimizing leakage losses. 3) Iteratively adjusting dimensions like impeller diameter until the calculated head matches the specified head within an acceptable tolerance.

1. CENTRIFUGAL PUMP DESIGN1
The client will usually specify the desired head and pump capacity. The type and speed of the
driver may also be specified. Speed is governed by considerations of cost and efficiency as well as drivers
available to the client. Given these parameters, the task of the engineer is to minimize cost.
Which cost to minimize, first cost or life-cycle cost, however, is an important consideration.
From a life cycle viewpoint, we must take into account power consumption and operation and
maintenance costs. These considerations call for optimizing efficiency, reliability (the mean time between
failure) and maintainability (the mean time to repair). In general, designing to optimize these categories
results in increased costs. Often, these considerations are not very important and we can design for
minimum first cost. In appropriate cases, the engineer should initiate a dialog with the client concerning
available options. For example, designing a boiler feed pump that operates continuously would probably
call for maximizing efficiency. Efficiency considerations would not be so important, however, for a
drainage pump that is only required to operate occasionally.
PIPE CONNECTIONS AND VELOCITIES
The diameter of the suction pipe is usually made larger that the pump suction flange and both are
made larger than the discharge flange and pipe. Church recommends keeping the velocity at the suction
flange about 9 or 10 ft/s and that at the discharge flange between 18 and 25 ft/s.
LEAKAGE LOSSES
To design the impeller, account must be taken of leakage from the discharge side back to the
suction side. To reduce the leakage, wearing rings are fitted to the impeller and casing. These rings are
designed with specified clearances. The leakage across each ring can be calculated from the following
formula:
2L LQ CA gH=
where: C = flow coefficient2
A = leakage area = / 2Dsπ
D = mean clearance diameter
s = diametrical clearance 3
0.010 ( 6)(0.001)D in= + −
For small wearing rings with precise machining and ball bearings, the minimum clearance may
be reduced to 0.008 in.
( )[ ]H U U gL = −
3
4
22
2
1
2
/ 3
IMPELLER INLET DIMENSIONS AND VANE ANGLE
The diameter of the impeller eye, Do, is dependent on the shaft
diameter, Ds, which must initially be approximated. The hub diameter,
1
This section is based on Church, A.H., Centrifugal Pumps and Blowers,
Ch. 6, John Wiley & Sons, 1950.
2
Id. Fig. 6-1, p. 92.
3
Attributed by Church to Stepanoff, A.J., Trans. A.S.M.E., HYD-54-5, 1932.
1

2. DH , is made 5/16 to ½ inch larger than Ds. After estimating Ds and DH ,
Do is based on the known flowrate. The inlet vane edge diameter, D1, is
made about the same as Do to ensure smooth flow.
EXAMPLE OF IMPELLER DESIGN4
Specified conditions: Required head: hP = 150ft
Required flowrate: Q = 2500 gpm
Required speed N = 1760 rpm
1. Quantity flowrate:
( )
( ) ( )
3
32500 min
5.57 /
min 60 7.48
gal ft
Q ft s
s gal
= =
2. Mass flowrate:
( ) ( )3
3
5.57 62.4
348 /m
ft lbm
m lb s
s ft
= =&
3. Specific speed: Assume a double suction impeller; then, Q = 2500/2 = 1250gpm, and:
[ ]
N
rpm Q gpm
h ft
rpmsd
p
= = =
ω( ) ( )
( )
( )
( )/ /3 4 3 4
1760 1250
150
1450
For this specific speed, a radial flow pump is indicated.5
4. Water horsepower.
2
2
(32.2) (150)(348)
550 (550) 32.2
ft ftmgh lbm s hp s
WHP
s s ft lbf ft
⋅
= =
⋅
&
94.6WHP hp=
5. Shaft diameter. Calculate shaft diameter based on torque. Increase the calculated value
somewhat to allow for bending moment which is unknown at this point and to ensure that
the critical speed exceeds the operational speed by a reasonable margin. The bending
moment will depend on the weight distribution of the shaft and any unbalanced radial thrust
acting on the impeller. From the figure shown below, with the given flow of 2500 gpm and
calculated value of specific speed of 1450, we select a tentative value of efficiency of 80%.
4
See, Church, p. 107-117.
5
See,Munson, Fig. 12.18, p. 812.
2

3. Thus:
94.6
118
0.8
WHP
BHP hp
η
= = =
The required shaft torque then is:
(550)(118) min (60) (12)
4230
(1760) min (2)( )
ft lbfW hp s rev in
T lbf in
s hp rev rad ftω π
⋅
= = = ⋅
⋅
&
Assuming a shear stress of 4000 psi:
2
33
16 (16)(4230)
1.75
( )(4000)
s
s
T lbf in in
D in
s lbfπ π
⋅ ⋅
= = =
To account for the unknown bending moment and critical speed, increase the shaft diameter to
2 1/8 in. Church states that the hub diameter, DH , is made from 5/16 to ½ in. larger than Ds:
Let DH = 2 ½ in.
6. Suction line velocity and diameter of suction flange.
Assume a velocity of 10 ft/s at the suction flange; thus:
Q
V DSU SU
=
( )π 2
4
; or,
3 2
2
(4)(5.57) (144)4
10.1, ,10
( ) ( )(10)
SU
SU
ft s inQ
D say in
V ft s ftπ π
⋅
= = =
⋅ ⋅
;thus,
3 2
2 2 2
(4)(5.57) (144)
10.2 /
( ) (10)
SU
ft in
V ft s
s in ftπ
= =
⋅
3

4. Assume the velocity at the eye of the impeller is 11 ft/s.
For a double suction pump, assume that the leakage will not exceed 2%. Dividing the total flow by 2
gives:
Q V A V
D DH
= = −0 0 0
0
2 2
4 4
( )
π π
2 2
0
0
(4)(1.02) (4)(1.02)(5.57)(144) 5
(2.5) 7.33 , ,7
( )(2) ( )(2)(11) 16
H
Q
D D in say in
Vπ π
= + = + =
7. Wheel inlet dimensions and angle.
Assume an inlet diameter, D1, of 7 5/16 in.
U r ft s1
1760 2 7 315
60 2 12
56 2= = =ω
π( )( )( )( . )
( )( )( )
. /
The radial velocity should be slightly higher than V0 because a converging shape is more efficient than a
divergent one. Let Vr be 12 ft/s.
The inlet area will be decreased by the vane thickness. Assume a contraction factor, ε1
6
, of 0.85; the
entering width then is:
b
Q
D V
in
r
1
1 1 1
102 557 144
2 7 31 12 0 85
175= = =
π ε π
( . )( . )( )
( )( )( . )( )( . )
.
Inlet angle: Assume that water enters vanes radially.
β1
1 1
1
1 012
56 2
121= = =− −
tan tan
.
.
V
U
r
β1 is usually increased slightly to account for contraction of the stream as it passes the inlet edges as well
as prerotation. The inlet angle is usually between 10 and 25 degrees7
. Let β1 be 130
.
8. Impeller diameter, D2.
The theoretical head can be found from integrating the force on a differential mass:
dF dmr= ω2
and dP
dF
A
= ; dm d brd dr= ∀ =ρ ρ φ
dP
brd dr r
brd
rdr r r=
⋅
= = −∫∫ ∫
ρ φ ω
φ
ρω
ρω2
1
2
1
2
2
1
2 2
2
2
1
2
2
( )
6
ε1 is generally between 0.8 and 0.9, Church, p. 95.
7
Church, p. 95.
4
β1
Vr1 W1
U1
U
dφ

5. but U r= ω and H
P
g
=
ρ
;hence, H H
P P
g
U U
g2 1
2 1 2
2
1
2
2
− =
−
=
−
ρ
For a closed rotating cylinder containing a fluid, the pressure head developed at the outer rim is:
H
U
g2
2
2
2
=
Substituting D2 /2(ω) for U2 and solving for D2 :
( )22 2
2
2 (2)(32.2) (60) 122 2 1840
(2)( )
HgH H
D
N Nω π
= = = (12)
Where: H2 is in feet; N is in rpm; D2 is in inches.
Tests have shown that the required impeller diameter can be calculated from this expression by
substituting the head corresponding to the best efficiency point for H2 and then multiplying the right side
by an experimentally determined coefficient Φ:
D
H
N2
1840
=
Φ (13)
Church8
gives several charts for Φ which have been based on a large number of tests. Most of the plotted
points fall within a range of 0.9 to 1.1. Noting that if the head on test is found to be too high, the impeller
diameter can be machined to an appropriate diameter, select 1.05 for Φ; then:
D in2
1840 105 150
1760
13 4= =
( )( . )
( )
. ; say, 13 ½ in.
9. Outlet vane angle, β2, and impeller width.
The normal range for discharge angles is between 20 and 25 degrees9
. Furthermore, β2 is usually made
larger than the inlet angle. Assume β2 = 200
.
The radial outlet velocity, Vr2 , is made the same as, or slightly less than, the radial inlet velocity, Vr1.
Assume Vr2 = 11 ft/s10
.
Outlet area (based on required flow plus leakage).
3 2
2
2 2
2
(1.02)(5.57) (144)
74.4
(11)r
ft s inQ
A in
V s ft ft
⋅
= = =
⋅
Assume a contraction width, ε2 , (based on experience) of 0.925:
8
Church, pp. 199-104.
9
Id., p. 35.
10
Id., p. 110.
5

6. 3 2
2 2
2 2 2
(1.02)(5.57) (144)
1.896
(11) ( )(13.5) (0.925)r
ft s inQ
b in
V D s ft in ftπ ε π
⋅ ⋅
= = =
⋅ ⋅
10. Outlet velocity diagram.
The absolute outlet velocity, V2 , is used in the design of the volute. We proceed as follows:
2 2
(1760) min(2)( ) (13.5)
103.7 /
min(60) (2)(12)
rev rad in ft
U r ft s
s in rev
π
ω
⋅ ⋅
= = =
⋅
Theoretical tangential outlet velocity, Vθ2.
V U
V
ft s
r
θ
β2 2
2
2
01037
11
20
735= − = − =
tan
.
tan
. /
Actual tangential outlet velocity, Vθ2’.11
The inertia of the rotating fluid causes a circulatory flow opposite to the direction of rotation of the
impeller. This flow, superimposed on the outward flow, results in the fluid leaving the impeller at an
angle less than that calculated from angular momentum theory. Thus β2 must be decreased and ,
therefore, the absolute angle, α2 , increased. The effect of circulatory flow is to reduce V2 and the
theoretical head. Church defines a circulatory flow coefficient, ηθ , as:
'
2
2
V
V
θ
θ
θ
η =
Church assumes a value of η∞ of 0.7. This coefficient can be calculated from tests. Pump manufacturers
will maintain records from which a reasonable value might be estimated for a given design.
V ft sθ2 0 7 735 515'
( . )( . ) . /= =
The outlet vector diagram can now be drawn:
α2
1 011
515
121'
tan
.
.= =−
, say, 130
V V V ft sr2 2
2
2
2 2 2
11 515 52 7' '
. . /= + = + =θ
11. Cross-section of impeller.
11
See, Church, p. 28 for a discussion of circulatory flow.
6
V’θ2
Vθ2
U2
α'2
α2
Vr2 Vr2
β2’
β2V2’ V2

7. Wall and vane thicknesses are usually made a minimum consistent with good foundry practice. The
stresses due to centrifugal force and fluid pressure are relatively low for average applications; otherwise,
they need to be taken into account12
.
Table of Calculated or Assumed Dimensions
b1 = 1.75 in per side
b2 = 1.90 in
D2 = 13 ½ in
D0 = 7 5/16 in
Dr = 8 ½ in (to outside of impeller wearing ring)
Impeller shroud tip thickness - 3/16 in
Connect outlet to inlet by a straight line faired into entrance to provide a smooth transition. Make tip of
hub core 3/16 in and fair into hub diameter. The drawing is shown in the figure on the following page.
12. Check leakage loss.
From the figure on page 8, the mean diameter of the clearance is 8 ½ in. Let s be the diametral clearance.
Church states that the wearing ring clearance for good practice is 0.01 in for rings of 6 in diameter and
less. For rings greater than 6 in, increase the clearance by 0.001 in for every inch of ring diameter greater
than 6 in:
s D say in= + − = + − =0 010 6 0 001 0 010 85 6 0 001 0 0125 0 013. ( )( . ) . ( . )( . ) . , , .
The clearance area is:
2 2
/ 2 ( / 2)(8.5)(0.013) 0.174 0.00121A Ds in ftπ π= = = =
Head across the rings13
:
H
U U
g
ftL =
−
=
−
=
3
4 2
3 1037 56 2
4 2 32 2
8852
2
1
2 2 2
( )( . . )
( )( )( . )
.
From Figure 6-1, p 92, Church, the flow coefficient for 1760 rpm and a 0.013 in clearance is 0.410.
Thus, the leakage is:
Q CA gH ft sL L= = =2 0 410 0 00121 2 32 2 88 5 0 0375 3
( . )( . ) ( )( . )( . ) . /
The per cent leakage is
0 075
558
100
.
.
( ) ; or 1.35 %, which is, close enough to the assumed value of 2 %.
12
Id,, p. 152.
13
Church attributes this equation to A.J. Stepanoff: “Leakage Loss and Axial Thrust in Centrifugal
Pumps,” A.S.M.E. Trans., HYD-54-5, 1932.
7

8. 8

9. DESIGN OF VANES
The entrance vane angle, β1 , has been found to be 130
; that at the exit, 200
. For smooth flow,
we must design the vane such that this angle increases smoothly from 130
to 200
. We note also that the
radial components of velocity to these two angles are 12 and 11 ft/s, respectively. We also see from the
vector diagram that W Vr= / sin β . The relative velocities corresponding to the entrance and outlet
stations are then: 12 13 53 30
/ sin . /= ft s and 11 20 32 20
/ sin . /= ft s . To obtain intermediate
values of radii corresponding to intermediate values of the position angle, θ , we proceed as follows (see
Fig. 3):
1) Plot β, Vr, and W against vane radius, r, for the entrance and outlet stations and connect by a
straight line (or a smooth curve).
2) The corresponding values for vane angle, β , are computed from sin /β = V Wr . These
values are also plotted against their radii.
Alternatively, write a computer program to perform the above functions. Referring to the figure below:
tan β
θ
=
dr
rd
or d
dr
r
θ
β
=
tan
θ
π β π β
0 180 180
1 1
= =
∇
∫ ∑
dr
r
r
rr
r
r
r
tan tan
Note: Use MATLAB or other computer system to perform the integration.
Use a sufficiently close spacing of r to obtain a smooth vane shape.
3) Plot the radii against θ to give the shape of the trailing edge
of the vane.
dθ r
rdθ
dr
9

10. Draw the front edge of the vane with the same curvature as the back edge with a thickness of about 1/8
in14
.
NUMBER OF VANES
The number of vanes is given by the Pfleiderer equation15
. First, calculate the average vane
angle:
β
β β
m =
+
=
+
=
1 2 0
2
13 20
2
165. ; then,
z no vanes
D D
D D
saym= =
+
−
=
+
−
=. . sin ( . )
( . . )
( . . )
sin . . , , .65 65
135 7 312
135 7 312
165 6 21 6
2 1
2 1
0
β
The circumferential pitch of the vanes is:
( )( . )
( )
.
π 7 312
6
383= in
Check the contraction factor:
ε
π
β
π π β
=
−
= −
D
zt
D
zt
D
sin
sin
1
ε
π1 01
6 0125
7 31 13
0855 085= − =
( )( . )
( . ) sin
. ( . )assumed
ε
π2 01
6 0125
1350 20
0 948 0 925= − =
( )( . )
( . ) sin
. ( . )assumed
The assumed values agree reasonably with those calculated.
SUMMARY
Diameter of suction flange, Dsu ------------------------------- 10in
Velocity in suction flange, Vsu ---------------------------------10.22 ft/s
Shaft diameter, Ds -----------------------------------------------------------------------2 1/8 in
Impeller hub diameter, DH --------------------------------------2 ½ in
Impeller eye diameter, D0 --------------------------------------7 5/16 in
Velocity through impeller eye,V0-------------------------------11 ft/s
Diameter of inlet vane edge, D1--------------------------------7 5/16 in
Velocity at inlet vane edge, V1 = Vr1 ---------------------------12 ft/s
14
Church, p. 115.
15
Id.
10

11. Passage width at inlet, b1 -----------------------------------------1.75 in per side
Tangential velocity of inlet vane edge, U1 -------------------56.2 ft/s
Vane angle at inlet, β1 --------------------------------------------130
Impeller outlet diameter, D2 -------------------------------------13 ½ in
Radial component of outlet velocity, Vr2 ----------------------11 ft/s
Vane angle at outlet, β2 -------------------------------------------200
Total passage width at outlet, b2 ---------------------------------1.98 in
Tangential velocity of outlet vane edge, U2 -------------------103.7 ft/s
Absolute velocity leaving impeller, V2
'
-------------------------52.5 ft/s
Tangential component of absolute leaving velocity, Vθ2
'
---51.5 ft/s
Angle of water leaving impeller, α2
'
----------------------------130
Number of impeller vanes, z ------------------------------------6
11