The table below can be used as a guide for selecting the applicable pump standards.

1.1 Main design standards for centrifugal pumps are:

Standard Industrial Application
1 API 610 (American Petroleum Institute)/ISO 13709 Petroleum Industry Heavy Duty Utilities (Boiler Feed) Europe & USA
2 ISO 2858/5199 (International Organization for Standardization) Chemical (Mainly Europe)
3 ANSI (American National Standards Institute) General Industrial Service (Mainly USA)
4 ANSI B 73.1 Chemical (Mainly USA)
5 ASME Section VIII Div 1 (American Society Mechanical Engineers) American code for pressure vessels. generally applied in the USA for pump case wall thickness calculations.
6 NFPA 20 & 25 (National Fire Protection Association) Fire Water System (World wide application, originally USA)
7 DIN (Deutsches Institut Fϋr Normung) 24255 General purpose (water) application (Mainly Europe)

1.2 Operating and service criteria.

Operating conditions should address the following items as a minimum:

  • Flow/liquid medium; Density; Specific gravity
  • Viscosity
  • Vapor pressure
  • Capacity, minimum, maximum and normal flow
  • Operating temperature; minimum, maximum and normal
  • Ambient temperature
  • Inlet/Suction pressure
  • Discharge pressure
  • Differential head
  • Maximum operating pressure
  • NPSHA (Available)
  • Continuous or intermittent
  • Critical operating conditions


Applications and operating conditions will determine the pump design configuration. Three main categories of centrifugal pumps can be distinguished as:

  1. Radial flow
  2. Mixed flow
  3. Axial flow

Typical impeller lay out are show above in figure 1

Pump configuration:

Within the radial and mixed flow categories there two main configuration with regard to the hydraulic design

  1. Volute type pump configuration – Fig. 2 (showing clockwise rotation)
  2. Diffuser type pump configuration – Fig. 3 (showing counter clockwise rotation)

US manufacturers traditionally favor a volute design; while European manufactures traditionally provide centrifugal pumps with diffusers. A volute design configuration imposes radial loads on the pump shaft, whereas in a diffuser design the radial loads are neutralized as they balance each other out. However in double volute designs, the radial loads are also hydraulically balanced and thus neutralized.


The specific speed is dimensionless, and is universally used as indicator to show the relationship between speed (rpm\N), flow capacity (Q) and differential head (H) (wc = water column). Ns = (N x √Q) / H 3/4

General European practice
Q = Capacity in mᶾ/.sec.
H = Differential head in m (wc)

N = Pump rotating speed in rpm

General US practice
Q = Capacity in gpm
H = Differential head in feet (wc)

N = Pump rotating speed in rpm

Typical Ns Values are

European US
Radial flow 9.7 to 500 500 to 4000
Mixed flow 38,7 to 155 2000 to 8000
Axial flow 139,7 to 387,3 7000 to 20000
Most generally used Ns range for industrial applications is: European 15 to 40; US 775 to 2065


Design requirements are related to operating conditions, criticality of service, investment costs and maintenance costs. API 610 pumps typically are mounted on pedestals extending to the center line of the pump. Multi stage API 610 style pumps in general have horizontally split cases. For high pressure multi stage pumps (> 150 bar or 2200 psi) a barrel design is often applied, in essence this is placing the pump casing in a high pressure vessel, i.e.; a barrel. 

Criteria to be considered are:

  • Horizontal base plate mounted, vertical inline, vertically suspended line shaft pumps;
  • Single stage or multi stage;
  • Single or double suction;
  • Impeller wear ring and case wear ring requirements;
  • Mechanical seal or packing;
  • Mechanical requirements e.g. thrust bearings, friction bearings or anti friction bearings;
  • Hydraulic balancing requirements, opposite impeller design versus balancing discs;
  • Impeller balancing holes/ports (Note: holes in the back shroud located near the impeller eye area, are not recommended;
  • Lubrication;
  • Hydraulic balancing;
  • Dynamic (mechanical) balancing.

Axial forces/loads caused by the internal flow circulation in centrifugal pumps must be hydraulically balanced. A thrust bearings, when installed, only counteracts the residual thrust (axial) forces/loads. Impeller wear rings in combination with case wear rings shall be considered for all critical services and medium to high capital cost pumps.

Most commonly applied design option for balancing the axial thrust loads of multi stage pumps, volute as well as diffuser type designs, are :

  1. Application of a balancing drum
  2. Opposite impeller design

It should be noted that designers of multi stage diffuser type pumps usually apply the balancing disc design. The pump casings are vertically split per stage. The stages the are sandwiched, held together by heavy studs mounted externally alongside the casings or the pump casings are placed in a barrel; a balancing line for providing pressure in the balancing chamber is adequate for neutralizing axial thrust loads (Fig 5). This design configuration makes it difficult to provide a full flow crossover line (Fig 6), because the cross over flow line has to be mounted externally. With a multi stage horizontally slit volute design configuration, the crossover line can be integrated in the casting of the pump case. See - Workshops - Attachment 5 for a assembly details.

By applying a balancing drum or disc, all impellers face the same direction. After the last stage a balancing chamber with a balancing drum or disc is to be installed, as is schematically shown in Fig 5. The balancing drum is connected to the suction of the first stage by a balancing tube. By selecting the diameter of the balancing drum in the balancing chamber, the axial thrust load can be balanced to 0, or directed to the left, or to the right depending on the preference of the designer. It should be taken into consideration that the clearance between the balancing drum and the stationary part of the balancing chamber also will increase over time, and may reverse the direction of axial thrust load.

With opposite impeller design arrangement half of the number of impellers are installed in the opposite direction as schematically show in Fig 6. In opposite impeller configuration there is a crossover line from the center (third) stage discharge to the suction of the next (fourth) stage at the other end of the shaft. In this configuration the axial thrust load from either direction are hydraulically balanced. At the center of the pump between the third and fourth stage, the stages are separated by an inner bushing. By selecting the diameter of the inner bushing, the direction of the residual loads can be determined. It also must be taken into consideration that over time, the clearance between the stationary and rotary parts will increase and may increase the residual thrust loads or even reverse the residual axial thrust load direction.

4.1 Dynamic balancing

Dynamic balancing of the rotating assembly is recommended for:

  • All multi stage pumps.
  • One or two stage pumps with a speed of 3600 rpm r above. (API 610 - Section states 3800 rpm)

4.2 Packing/Mechanical seal selection and Auxiliary piping.

(Reference API 610 - Sections 5.3, 5.8 and Annex B,

  • Packing type for packed pumps
  • Mechanical seal – vendor standard.
  • Special applications;
  • Seal leakage detection:
  • Seal flush requirements:
  • Seal flush medium;
  • Cooling piping:
  • Lubrication piping.


Power output of motor drivers shall be at least 15 to 20 % above the maximum shaft peak power requirements of the pump, including power required for all losses due to efficiency factors. (Reference API 610 - Section 6 - Table 11)

  • Electric Motor driver types:
  • Synchronous motors;
  • Induction motors;
  • Variable Speed motors.
  • Diesel drivers - Mainly used for firewater system, de-watering pumps and temporary water supply and distribution.


Solid couplings shall only be used when pump and motor are assembled as an integrated unit. In some designs, the pump thrust loads are taken up by the motor thrust bearing. Flexible couplings shall be used when pump and motor have their own independent bearing assembly. The pump motor assembly shall always be aligned. The flexibility of couplings is only to compensate for residual misalignment.

Spacer length of the coupling shall be such that sufficient space is provided to allow back pull out of the rotating assembly and the removal or mechanical seal (for maintenance purpose) without removing either the pump casing or driver. Coupling guards or screens shall be provided to prevent personnel from contact with the rotating parts during operation. Non sparking coupling guards shall be used for volatile service in oil and gas, refinery and chemical services. (Refer to API 610 - Section 6.2 for special requirements).


Horizontal pumps and drivers shall be mounted on a single base plate. (For oil and gas applications please see API 610 - Section 6.3.) Skid type base plates, platforms and sub frames shall have adequate stiffness for ensuring continuous alignment, and shall not be subjected to deflection or distortion due torsional forces or bending loads.

Grouted base plates are recommended for heavy duty service on permanent on shore installations.


To prevent distortion of the pump casing and subsequent misalignment the nozzle loading from the connecting process piping must be avoided. (For maximum allowable nozzle loading See API 610 section 5. table 4 and annex F).

To prevent pump damage due to sustained low or no flow, a minimum flow must be guaranteed at all time. A bypass line must be considered to ensure the minimum flow. Minimum flow requirements may vary from 10% to 35% of rated flow depending in the pump design. Minimum flow requirement data must be specified and provided by the pump manufacturer.


For any installation, start up and shut is always critical activities. Sequence of actions must be established, specified and adhered to. The sequence of actions shall be such, that the process is under control at all time without any adverse effect on safety and equipment.

When establishing the sequence of actions the following issues are to be considered

  • Establish start up and shut down process criteria;
  • Establish equipment start up and shut down condition and assess the off specification conditions that may be a cause for any potential damage
  • Ensure all potential off specification conditions have been identified and safeguarded;
  • Review and analyze the complete pump driver protection system;
  • Define process shut down criteria;
  • Define pump protection shut down criteria;
  • Define the startup permissive i.e.; momentarily shutdown bypass, for each action.
  • Unless variable speed motors or motors are provided with soft starters, startup of centrifugal pumps must always be against closed discharge.

Centrifugal pumps by nature of design are not self-priming. Some special pump models provided with high inlet nozzles location above the eye center line have some priming ability but are rarely used for continuous process or utility duty and are usually applied as drainage or sump pumps with flexible inlet/suction piping or hoses. A non return foot-valves must be installed in the suction piping of a self-priming centrifugal pump to prevent back flow. To achieve the self-priming effect the suction line and pump housing must be filled with the pumping liquid (water) before start up. Provisions for filling the inlet line and the pump housing must be provided. 


Material selection reference API 610 - Section 5.12 and API 610 - Appendix G shall be used for high pressure heavy duty utility and oil and gas services. For hydrocarbon services, pressure casing shall be made from carbon steel or alloy steel. Cast iron may be considered for other services e.g.; low pressure water service. Cast iron is also acceptable for non- pressure containing parts, such as; bearing housings or bearing brackets. Inspection certificates as per ISO 10474 – are to be provided as applicable and required. For heavy duty applications: B.3.1 b, B.3.1.c material certificates should be considered at all times. Material certificates 3.2 may be accepted for less critical duties.

Cast iron is less expansive and easier to machine that carbon steel, hence some pump manufactures use cast iron for bearing housings to reduce cost and naturally to be more competitive.


11.1 Power requirements

Driving power requirement.

For estimating the power (P) requirements, the equation for P can be used for establishing the indicative power requirements. The efficiency factor (η) is not a constant but is a performance function of flow (Q; in m³/sec) and differential head (H; in m) and therefor is a variable. For sizing purpose it can be considered as a constant, η = 0.70 (70%) is a common efficiency value for general purpose pumps, axial and mixed flow.

P in kW: P = (ρ x g x Q x H) / (36180 x η)

11.2 Cavitation and Net Positive Suction Head (NPSH)

Cavitation occurs just after the point of entry of the fluid in the vane channel area. Due to the acceleration of the flow, a temporary pressure drop may occur just after the fluid has entered the vane channels, and can reach a pressure level below its vapor pressure at the operating temperature, and may create a vapor bubble. The bubble will implode when the pressure recovers before or at the point of exit of the impeller; the energy being released will cause impingement on the impeller vane channels or in the transit area to the diffuser vanes or volute. Frequent re-occurrence of this phenomenon will cause severe damage to the impeller and pump housing. Hence the suction pressure must be such that the phenomenon of cavitation can be avoided by keeping the pressure inside the impeller above the vapor pressure of the fluid.The net positive suction head/pressure is universally referred to as NPSH. The NPSHR (Required) is the required NPSH to keep the pump running in a balanced mode.The available NPSH is referred to as NPSHA (Available). The NPSHR data must be provided by the pump manufacturer. The NPSHA is a function of the suction piping configuration and must be established by the designer/engineer of the piping system. 

Hence, cavitation may occur when: NPSHA (Available) < NPSHR (Required)

NPSHA can be fairly accurately determined by: NPSHA = Ha + Hs – (Hl + Hv)

The NPSHR (Required) is a function of the pump configuration and must be provided by the pump manufacturer.

  • Ha = Atmospheric pressure
  • Hs = Static head
  • Hl = Line losses i.e. pressure drop in the suction line.
  • Hv = Vapor pressure at operating temperature.

Pressure can be indicated in meter head is or any other measuring unit as long as it has been consistently applied.

11.3 Modifying existing pumps.

Performance of existing pumps can be increased or reduced by changing the pump speed or changing the impeller diameter. The power requirements will increase or decrease proportionally. It should be noted that losses due friction will increase when increasing the flow (Q), hence the efficiency will decrease.  

11.3.1 Changing the impeller OD

When changing the impeller diameter OD the differential head will change as the peripheral velocity of the impeller will change. There may be a change in efficiency due to the change of velocity. This effect however can be discarded for estimating purposes. The original situation is indicated as (0) the changed situation is indicated as (1). The flow capacity Q will not change. 

The required peripheral speed U2 of the impeller OD  U2 = √(2g*H)/η; not considering the efficiency η the head H = U2²/2g.  U2 is a linear function of the OD. The differential head - H - will change quadratic.

  • ξd = modification factor for changing impeller OD. ξd = OD1/OD0
  • The differential head H1 = ξd² x H0
  • The specific speed Ns1 = (N0 x √Q0) / H1¾

11.3.2 Changing the speed (rpm)

When changing rotating speed the peripheral velocity U2 of the impeller will change, so the general equation for the differential head is the same as for changing the impeller OD; H = U2²/2g.

  • The modification factor ξs = N1 / N0
  • The differential head will change will change: H1 = ξs² x H0
  • The flow Q1 will change linearly and will be: Q1 = ξs x Q0
  • The specific speed Ns1 = (N1 x √Q1) / H1¾