Eureka Marine AS Norway

Your partner in Eureka® marine pumps

EUREKA® trademark is registered with the Norwegian Patent Office

PUMP SELECT

This handbook is prepared as a support for customers that have the task to specify- and purchase pumps for a new or existing installation

About Eureka Marine

Eureka Marine is a company which is a specialized supplier of EUREKA® centrifugal pumps and spare parts to EUREKA® pumps on delivery just in time.

We can ensure full customer satisfaction with our technical know how and a detailed understanding. More than 20 years of experience with Eureka pumps from:Thune-Eureka marine division, Tranby NorwayKværner Eureka Tranby NorwayKværner Ships Equipment Tranby Norway

PUMP SELECT

This handbook is prepared to backup customer support according to their needs tospecify- and purchase pumps for a new or existing installation.

It is prepared by Truls H. Paulsen, former head of Development/research and technical department for engine room pumps with Kværner Eureka in Norway.

EUREKA® trademark is registered with the Norwegian Patent Office

General

2 Hydraulic data

2.1 Calculation example2.2 Units2.3 Inlet pressure2.4 Cavitation2.5 Discharge pressure and pump head.2.6 Power and efficiency2.7 Adjustment of pumps

3 Liquid properties

3.1 Density3.2 Viscosity3.3 Vapour pressure

4 Material

4.1 Selection4.2 Corrosion4.3 Erosion-corrosion4.4 Material combinations-galvanic corrosion4.4 Crevice- corrosion and pitting4.5 Strength properties4.6 Fatigue

5 Speed selection

5.1 Synchronous and asynchronous speed5.2 Change of rpm5.3 Vibrations5.4 Balancing5.5 Assembly5.6 Ball bearing lifetime5.7 Mechanical seal

1 General

This handbook is established as a support for customers that have the task of specifying and order

pumps for a new or existing installation. The starting point for the task will be the specified hydraulic

performance in the inquiry.

Large user organizations as shipyards and shipowning companies use standardized specification forms

that are very helpful in selecting of equipment, but sometimes inquiries are wanting in completeness

and may be the reason for unsatisfactory pump selection.

We shall go through some of the main points in specification and selection of centrifugal pumps for use

on ships.

2 Hydraulic data

2.1 Calculation example

We have a specification that calls for a pump for 5 bar pressure and flowrate of 100 m3/h and seawater

pumping. And here comes the first task. User- and supplier specifications often use different units and

the need for unit conversion rises in order to make a pump selection. To convert pressure into m head

you must divide by *g.

p 5bar 5 105Pa=:= 1033

kg

m3

:= g 9.807m

s2

=

Insertion in expressions H and Q gives:

Hp

g49.357m=:=

Q 100m3

hr0.028

m3

s=:=

On unit conversion it is important to calculate with corresponding units so that the digits and the units

are correct.

2.2 Units

There are 7 base units in the SI system and of these 4 are important in pump calculations

Length meter m

Mass kilogram kg

Time second s

Current ampere A

The base units may be grouped to form units important to pump calculations

Force newton1 N 1

kg m

s2

:=1

Energy joule 1J 1 N m:=1

Power watt1 W 1

J

s:=1

Pressure pascal1Pa 1

N

m2

:=1

Some common units which are not part of the SI coherent system is written below

Pressure bar

Specific weight1

gm

cm2

1000kg

m3

:=1

Density1kg

m3

Head m Pa

g

Energitet Y N m

kg

Power Hk 1 Hk 0.735 kW:=1

Force kp 1 kp 9.81N:=1

2.3 Inlet pressure

To make a good selection of a pump the inlet pressure must be a part of the specification. This pressure

determines the available NPSH which again is deciding for cavitation risk and pump speed selection..

Available NPSH is calculated by the expression below.

NPSHa

p1

g

c12

2 g+

B

g+

pd

g:=

p1 Inlet pressure Pa

density kg

m3

g acceleration of gravity9.81

m

s2

c1 inlet velocity m/s

B barometric pressure Pa

pd vapour pressure Pa

Required NPSH is a machine characteristic and is given in the pump data sheet and in the performance

curves. For the pump to have acceptable running performance the

NPSHavailable – NPSHrequired > 1m

There are several terms form the pressure in the suction pipe to characterize the cavitation risk. The

diagram below shows the relation between the terms with relation to the absolute – and the atmospheric

pressure.

2.4 Cavitation

This phenomenon has to do with the boiling temperature of the pump liquid at the actual pressure.

When the pressure in the liquid reaches the vapour pressure the liquid starts pulsating boiling with noise

and vibration as a result. The immediate damage effect is a drop in pump head. This effect is what is

used in measuring the onset of cavitation and the basis for the NPSH curve. The value of

NPSHavailable that leads to a fall of head of 3% is what is defining the NPSH curve. At this condition

it is not possible to increase the flowrate even if the throttle valve is fully open. The pump impeller inlet

passages are blocked by vapour collection.

The figure shows the NPSH curve as a function of the flowrate and the figure to the right shows the

effect of cavitation on the pump curve.

Cavitation in the impeller inlet is a common reason for erosion of blades, wear rings and shroud.

Cavitation erosion going on for a long time will lead to visible damage of the impeller and need for

replacing the impeller.

2.5 Discharge pressure and pump head.

The sum of the friction head and static head is balanced by the pump head. The expression of head is

given below.

Hp2

g

c22

2 g+

p1

g

c12

2 g:=

The head is equal to the difference of the flow energy between inlet and outlet.

The static head is given by the elevation difference between the suction level and the head basin. The

friction head is given by the pipe friction, bend friction and outlet loss. We find these head components

in the pump diagram.

2.6 Power and efficiency

The power needed to lift Q m3/s with a density of =-1033 kg/m3 up a head of H m is given by the

expression below.

Pg Q H

1000kW:=

The power supply performs the ”useful” work of g Q H , but must also cover the losses that occur in

the pump as: hydrodynamic loss, volumetric loss and parasitic loss. The pump efficiency is the total

effect of these and is expressed by h v m:=

To measure these singular losses is difficult and involves costly test facilities in laboratories. For

everyday use it is the total efficiency that is of interest and that is what is shown in the pump diagram.

The total efficiency can be calculated by separating from the expression for P and solving as shown:

g Q H

P 1000:=

h v m:=

The energy use in pumping and the value of efficiency are important factors in pump operation while

the largest part of the operating costs are the energy cost.

2.6 The pump curve and selection of pumps

The pump curve presents the performance data as functions of the flowrate, H-Q, P-Q, NPSH-Q and

-Q. This presentation gives a clearly set out presentation of the important performance data of the

pump and is the basis for a factual evaluation of the installation. Down below are given some points for

an evaluation.

On the figure below is shown pump curves of head, efficiency and NPSH. Several operating points are

marked by triangles on the head curve together with a short name which is commented further down.

The names indicate the flowrate at different operating conditions and give a short characteristic of the

condition to be commented.

Closed valve:

With closed valve the pump power is used to internal circulation and all power is converted to heat. The

temperature increase is fast and the pump must be stopped in a short time, less than half a minute, in

common cases.

Qhalv

A flowrate of about half capacity, 50% of Qopt may lead to recirculation and a

phenomenon called rotating stall leading to vibrations and noise.

Qmin:

A flowrate down to 75% of Qopt gives safe pump operation, but often vibrations

when approaching 75%.

Qgod An operation between 100% and 75% of Qopt is a safe alternative with good

utilization of power, low NPSH and quiet running.

Qopt This is the design point of the pump and the flow regime is good and the

efficiency is at a maximum.

Pumpedata

0

20

40

60

80

100

120

140

0 250 500 750 1000

Q m3/s

H,e

ta,P

,NP

SH

NPSH

H

eta

Qmin

Qopt

Qhalv

stengt vent

Qfull

Qgod

Qmax Operating point with flowrate approaching 110% of Qopt is often utilized, but

with increasing flowrate the NPSHrequired increases quickly and cavitation may

give head drop, vibrations and erosion with longer operating durations.

Qfull With throttle valve fully open the pump operates with heavy cavitation, vibration

and noise. Long operating time will lead to damaged bearings, impeller and seals.

2.7 Adjustment of pumps

Adjustment of standard pumps to meet a given specification is often necessary. As an example the

operating point is below the H-Q curve. If the impeller diameter is reduced from D2 to D21 we get a

lower H-Q curve and the operating point is on the curve.

On adjustment of pumps it is useful to be aware of the fact that the pump has the best operating

condition for the flowrate with highest efficiency with a given speed. All other operating points have

lower efficiency. The flow angles into the impeller and into the tongue in the spiral are decisive for the

pressure losses and by that the efficiency. The impeller blade inlet is optimum at one flowrate at a

given speed. For the flow angle into the casing the operating points lying on a straight line through

origo and the head at optimum efficiency gives the best flow conditions. These two relations are given

in the below diagram

3 Liquid properties

3.1 Density

The mass of liquid per unit volume is given by in units kg/m3. The density is a factor in the expression

for the pump power

Pg Q H

1000kW:=

The density is also a factor in the formula for calculation of the pump head in m

p g HN

m2

:=

Density of some liquids:

Water998 2

kg

m3

,20 ° C

Seawater1027

kg

m3

20 ° C

Lubrication oil900

kg

m3

20 ° C

Pump head expressed in meters is independent of the liquid density. Whether you are pumping water or

mercury the impeller will produce the same head, but the pump power and the pump pressure will of

course be very different according to the above expressions.

3.2 Viscosity

Viscosity is a measure of the liquid friction and is measured in centistokes or m2/s.

1cSt 106 m

2

s:=

Viscosity for some fluids:

Water

1 106 m

2

s

20 ° C

Lubrication oil

90 106 m

2

s

20 ° C

For systematic description of frictional losses in flow regimes the Reynolds number is applied.Re

c d:=

Where c is velocity, d is a characteristic dimension of the flow system and is kinematic viscosity.

Frictional loss in pipes are frequently seen in calculations of dimensioning and the Moody diagram shows

the relation between the Reynolds number, relative roughness and the friction coefficient for pipe flow.

The friction loss in a pipe of length L, diameter D and flow velocity c is given in m friction head.

hr fL

D

c2

2 g:=

On pumping of viscous liquids like lubrication oil, bunker oil and crude oil the frictional losses internally

in the pump lead to noticeable changes of pump performance. The head is lowered, the flowrate is reduced

and the input power increases. Technical handbooks gives more information of this subject.

3.3 Vapour pressure

It is common knowledge the water boils at 100 degrees centigrade. In technical language this fact is

expressed as water vapour pressure is at 1 bar with 100 degrees centigrade. The vapour pressure is

dependant of the liquid temperature. On pumping of liquids with temperature near the boiling point the

risk of cavitation is present. In the expression for NPSH a we find the value of the vapor pressure as an

ingoing factor.

NPSHtil

p1

g

c12

2 g+ B+

pd

g:=

With low inlet pressure and high liquid temperature the NPSHa value must be controled to check

whether the pump has sufficient margin against the pump required NPSH. Below is shown the vapour

pressure curve for water between 0 and 100 degrees centigrade

4 Material

4.1 Selection

There are two parties in selection of materials for pumps. One of them, the pump manufacturer, has

built the pump with a selection af materials that has favorable strength properties , manufacturing

properties, manufacturing cost and which covers broad application fields of pump media. The user

has the task of selecting the pump in a material combination that is suited to handle liquids with

corrosion properties, temperature and particle content

4.2 Corrosion

Corrosion is a complex chemical-physical process which represents a large sphere in itself and which

demands a large study effort to get hold of. Here will be commented three forms of corrosion that

often leads to damage and breakdown of pumps.

The corrosion properties of materials is usually presented in tabulations where the position in the

galvanic series is described. These presentations are based on measurements in still electrolytes as for

instance seawater. Corrosion and breakdown of materials is strongly affected by flow velocity and in

common pumps the flow velocities are quite large, 5m/s-30m/s

4.3 Erosion-.corrosion

Erosion-corrosion takes place in flow regimes involving cavitation and/or particles that tear down the

passive protection of the material surface and this makes it open for corrosion. The diagram shows

loss of material as a function of flow velocity for some common materials. The large difference

between the resistance to corrosion of bronze and mild steel is apparent

4.4 Material combinations-galvanic corrosion

Pumps are equipped with details in different materials by reason of strength, function and manufacturing

properties. Material combinations utilized in seawater must consider galvanic corrosion. This corrosion

form comes where we have materials with different electrode potential that are in electrical contact with

each other. It is the less “noble”,anodic, component that is corroding while the “noble”, cathodic,

component is protected. It is good practice to have the anodic material for parts with large surface in

order to reduce the intensity of corrosion attack

Below is shown the galvanic series for some materials

With the pump casing as anode and the impeller cathodic the part with the highest velocities is

protected by the much larger surface of the pump casing. An example of this can be pump casing in

bronze and impeller in Ni-Al bronze which is a common material selection for seawater pumps. Bolts

and nuts must be selected in nobler materials for pump flanges and cover bolting. Stainless steel bolts

must be selected.

4.4 Crevice- corrosion and pitting

Pump details like shaft, sleeves, seals, impeller and wear rings are installed in the pump with narrow

clearance and these details are often manufactured in stainless steel. In the clearances still seawater

stays and will be unaffected by the flowing of seawater in the hydraulic passages. The passivation of

the surfaces in the clearances will break down and the surfaces are the open for corrosion attack,

crevice corrosion and pitting will attack the internal surfaces of the clearing. For pumps with long

periods of still standing it is recommended to start a short pumping operation with fresh water in order

to prevent this form of corrosion

4.5 Strength properties

The shaft is the part of the pump with the highest stress level and also with corrosion attack from the

pump fluid, fatigue load from imbalance and hydraulic forces and torque moment, bending moment

and thrust forces during operation.

The cast pump casing in bronze or cast iron for common duties in ships are subject to internal

pressure, pipe forces and corrosion and corrosion-erosion. Material stress are low because the casting

of these parts require larger material sections than what is required for strength reasons.

Cast iron and bronze materials have lower modulus of elasticity than steel materials and internal

pressure and pipe forces may lead to deformations and make the running clearances in wear rings and

bushings to disappear with extraordinary wear rates as a result.

Below is shown some material data for common pump materials

EN and NS standard

Tensile strength

Rm N/mm2

0,2%proof strength

N/mm2

E-modul E

N/mm2

Tetthet ñ

Kg/dm3

Copper Alloys

EN 1982 CuAl10Fe5Ni5-B CB333G 600 250 1,27E+05 7,6

EN 1982 CuSn10-B 260 130 8,7

EN1982 CuSn7Zn2Pb3 CC492K 250 130

EN1982 Cu85Zn5PB5-B CC491K 200 100

Stainless steel

SS142343 450 210 2,00E+05 7,9

SS142343 450 220 2,00E+05 8

SS142324 600 450 2,00E+05 7,7

Cast Iron

NS 11370 700 420 1,08E+05 7,1

SjG-30 300 1,44E+05

4.6 Fatigue

The pump shaft is exposed to fatigue loading from imbalance, pulsating hydraulic forces and incorrectshaft alignment. These cyclic loads are in some ways functions of the shaft speed. The fatigue strength ofmaterials are determined by cyclic loading of test specimens and in most steel materials have the fatiguestrength at 10 million cycles. Down below is shown some data for material strength and fatigue strength

The fact that it is important to make sure that shaft and other parts exposed to cyclic loading are

dimensioned for fatigue loading is indicated by the simple calculation below giving the time in days to

reach the fatigue limit with different rpm

npump

1200

1800

3600

:= dager107

npump 60 5

28

19

9

=:=

We notice that it is only 9 days at 3600 rpm to reach 10 million cycles with 5 hours running per

day. The importance of shaft dimensioning, manufacturing with generous radii in section changes

and good surfaces can not be overestimated.

5 Speed selection

5.1 Synchronous and asynchronous speed

Short circuit motors have rpm directed by the frequency of the supply circuit. In Norway and most

countries the net frequency is 50 Hz, but for ships it is usually 60 Hz. The most common rpm for pumps

are: 50 Hz,

Net frequencyf50

50 Hz:=

Synchronous rpm: 750, 1000, 1500, 3000

Asynchronous rpm: 727, 970, 1455, 2910

Net frequencyf60

60 Hz:=

Synchronous rpm: 900, 1200, 1800, 3600

Asynchronous rpm: 873, 1164, 1740, 3492

5.2 Change of rpm

On changing of the speed a number of effects are also changed such as: Hydraulic performance, material

stress, wear and vibrations. Change of rpm up to a higher speed leads to higher performance , but also to

higher loads. Some off these changes will be commented below and also given numbers As an example

for numeric calculation rpm changes from 1740 rpm to 3492 for a pump with impeller diameter 250 mm.

n1

1746:= n2

3492:= D2

0.250:=

Peripheral speed outlet diameteru21

n1D2

6022.9=:=

Peripheral speed at outlet diameter at increased rpm u22

n2

n1

u21

45.7=:=

For the inlet diameter the peripheral speed is changed by the same amount

The flowrate is changed by the same

amountQ2 Q1

n2

n1

:=

He pump head is changed by the quotient of the rpm in power 2H2 H1

n2

n1

2

:=

The risk of cavitation expressed by NPSH increases by the quotient of rpm in

the second power NPSH

2NPSH

1

n2

n1

2

:=

From this it is clear that the increase of flowrate also increases the flow velocities in the same rate. The

danger of increased erosion-corrosion is strongly influenced by velocity increase.

The head and with that the pressure increases by:H2

H1

n2

n1

2

:=

He pump casing gets greater material stress and the shaft seal experiences higher pressures..

The power increases with:P2

P1

n2

n1

3

:=

And the shaft torque increases with the amount:

He torque stress increases in the same amount M2

P2

n2

:=

NPSH required increases by:NPSH

2NPSH

1

n2

n1

2

:=

And this leads to a need for higher pressure in the inlet section

5.3 vibrations

The forcing frequency is proportional to the rotational frequency multiplied by a factor depending on the

form of the excitation.

Slide bearing f1

0 5 n,:=

Imbalance f2

1 n:=

Assembly inaccuracy f3

2 n:=

Ball bearing damage F5

zkn:=

It is evident from this that all these excitation sources are proportional with the rpm and with that also with

the ratio of change of the rpm. The critical frequency of the shaft however is practically independent of the

speed. The information that we now have about critical speed, excitation frequencies and change of rpm

may be shown in a diagram that gives a good view of the possible resonance frequencies in the shaft

system.

In the diagram is shown 4 lines for pump speed and excitation frequencies 0,5*n, 1*n, 2*n, 6*n together

with first.- and second order harmonic frequencies f1 and f2 .It is evident from the diagram that for speed

1746 the excitation frequency 1n is lower than the harmonic frequency f1. This is designated an

undercritical shaft system which is considered as the preferred design solution for pumps. We also see that

if the rotational speed is changed to 3492 the excitation frequency is almost equal the second order

harmonic frequency f2. This may lead to higher vibration level, but is usually to live with.

Calculation of harmonic frequencies are often complicated and must usually be tackled by specialist

engineers. Vibration problems in pump operation leads to calculation work or even measurement of

vibration levels and frequencies. A good rule of thumb is that the excitation frequency should be 15% -

20% under the first order harmonic frequency.

5.4 Balancing

Balancing is done according to ISO 1940 standard that gives the rest imbalance in 10 classes where G6.3

is the commonly used for pumps. It is evident from the diagram that the rest imbalance in the class is

depending on the pump speed. A pump for 3600rpm must have a rest imbalance lower than the 1800 rpm

pump has.

5.5 Assembly

Inaccuracies in the alignment of pump and motor shaft leads to excitation frequency of 2*n. Check of

alignment is done with micrometer and accepted eccentricity between pump and motor coupling is 0,05

l k kk

mm radially and axially. Inaccuracy in alignment leads to bending stress in the shaft and imbalance and a

speed increase makes these damages worse. Below is shown some shafting systems in principle which

indicates the CG pump has the most tolerant shaft system taking care of both eccentricity and angular

faults.

5.6 Ball bearing lifetime

The lifetime of ball bearings depending on dynamic load and speed may be had of this formula giving the

lifetime in hours.

L10h

C

P

3

60 h:=

C is the dynamic load number from the ball bearing catalog and P is the dynamic load and n is the rpm.

In the formula the rpm is a direct factor, but the rpm is also a factor in the calculation of the dynamic load

such as is shown above. Imbalance increases with the second power of the rate of change of the rpm and

this leads to the result that the lifetime is reduced by the speed in power 5.The lifetime of ball bearings are

influenced by a number of mechanisms as corrosion, wrong lubrication, vibration from the support

structure under still standing conditions

5.7 Mechanical seal

Mechanical seals work like a thrust bearing in operation. The wearing surfaces operate in mixed friction

and the parameter for wear is the PV factor which is going to be discussed a little more in detail..

Wear occurs when two surfaces are forced together by forces axially to the sliding movement.

Irregularities in the surfaces come in direct contact and the relative sliding between the surfaces tear of

particles from the high points. The mechanical seals do have a thin fluid film between the surfaces but this

is not enough to make a hydrodynamic lubricating film. The seal faces operate in a mixed friction mode

with partial materials contact in the surfaces. The wear rate is a function of the surface pressure and the

relative sliding velocity between the surfaces. The P*V parameter which is the product of the surface

pressure and the sliding speed is what determines the rate of wear and this is also used as a parameter for

seal selection. In order to make the calculation of the P*V a little simpler without recourse to the

geometrical details of the seal and shaft the expression is p*v with p being the stuffing box pressure and v

is the peripheral speed of the shaft diameter and the unit the is bar* m/s

vn

30

d

2:=

With this simple expression in mind it is quite quick to establish a preliminary selection of a seal. We also

see that here again the speed is a factor in wear rate. The wear is a function of the pv factor. Below is

shown a diagram with pv factor for two different seaL with pv = 50 for ordinary applications and one up

to pv = 500 for larger pumps with higher pressures.

Contact us

Please contact Eureka Marine if you want an estimated cost for a new Eureka centrifugal pump without any obligations.

Eureka Marine ASSkogliveien 4,N-3047 Drammen.Norway. Free service Tel.: +47 40616140 E-mail: spares@eurekamarine.noWeb: www.eurekamarine.com

Eureka Marine AS Norway

Your partner in Eureka® marine pumps

EUREKA® trademark is registered with the Norwegian Patent Office

Eureka®GUARANTEE

Eureka® marine pumps. Delivered from Eureka Marine is guarantied against faults in material and workmanship for 12months after delivery, based on “Liability for defects” as per ORGALIME S92 General Conditions