- Impeller diameter
- Propeller degree
- Speed
- Number of stages
Pump Performance Chart
The performance curves differ in exact one parameter, such as
Author: bs
Curve Corrections for Different Fluids
In practice this will however become not applicable for fluids of kinematic viscosities above 20mm²/s ( 20 cst ) due to the influence of a higher Reynolds Number with increasing viscosity. Empirical conversion methods have been developed for curve corrections for medium- and high-viscous fluids, which in practical application in older versions meant the time-consuming evaluation of diagrams, but in the current versions were prepared by corresponding formula sets.The most widespread method worldwide is the method from the Hydraulic Institute (USA), which has been standardised as ANSI/HI 9.6.7 and ISO/TR 17766. In practice, the conversion is usually carried out by computer programmes such as the Spaix PumpSelector. The computer implementation of this method enables the conversion of performance curves, whereby the user only has to define the desired delivery data and the pumped medium. The best efficiency point of the pump is of vital importance with all the known curve correction methods.The following conditions can be stated for the validity of this method:
- Centrifugal pumps with closed or semi-open impellers
- Kinematic viscosity in the range between 1 and 3000 mm²/s
- Flow rate at best efficiency point between 3 and 410 m³/h
- Head per stage between 6 and 130 m
- Fluid displacement by means of a centrifugal pump under normal operating conditions
- Handling of Newtonian Fluids
Performance curve conversion for impeller trimming
The following applies approximately:
Q = flow rate
H = delivery head
D = impeller diameter
r = index for the reduced impeller diameter
t = index for the reference wheel diameterThe throttle curve H (Q) can be roughly determined from this relationship.A more precise calculation, however, requires the consideration of performance charts in which an impeller diameter is assigned to each performance curve. The new performance curve is determined by interpolating the conversion from the neighboring curves. In order to fully utilize the efficiency of the method, it is recommended to record an duty chart with at least three performance curves. If there is a large difference between the smallest and largest impeller diameter, some (2..4) intermediate curves are required.An alternative calculation method is described in ISO 9906. Knowledge of the mean impeller diameter at the leading edge D 1 is required here. According to the standard, this procedure is valid for
D 1 = Mean diameter at the impeller leading edgeFor pumps with a type number K ≤ 1.0 and a maximum impeller diameter reduction of 3%, the efficiency can be considered as unaltered.
Q = flow rate
H = delivery head
D = impeller diameter
r = index for the reduced impeller diameter
t = index for the reference wheel diameterThe throttle curve H (Q) can be roughly determined from this relationship.A more precise calculation, however, requires the consideration of performance charts in which an impeller diameter is assigned to each performance curve. The new performance curve is determined by interpolating the conversion from the neighboring curves. In order to fully utilize the efficiency of the method, it is recommended to record an duty chart with at least three performance curves. If there is a large difference between the smallest and largest impeller diameter, some (2..4) intermediate curves are required.An alternative calculation method is described in ISO 9906. Knowledge of the mean impeller diameter at the leading edge D 1 is required here. According to the standard, this procedure is valid for- Diameter reduction up to max. 5%
- Type number K ≤ 1.5
- Unchanged blade geometry (outlet angle, tapering, etc.) after cutting
D 1 = Mean diameter at the impeller leading edgeFor pumps with a type number K ≤ 1.0 and a maximum impeller diameter reduction of 3%, the efficiency can be considered as unaltered.Calculating the System H-Q Curve
The required pumping head in a branchless pipeline is determined from BERNOULLI’s equation for one-dimensional, stationary flow of incompressible fluids:
pin, pout = pressures on suction respectively discharge liquid levels
ρ = fluid density
g = gravity (9.81 m/s2)
Hgeo = static height difference between suction and discharge liquid levels
Hl,tot = total pipe friction loss between inlet and outlet areas
vin, vout = mean flow velocities at inlet and outlet liquid level areasThe mean flow velocities at the inlet and outlet areas are, based on the Continuity Law, mostly insignificantly small and can be neglected, if the tank areas being relatively large compared to those of the pipe work. In this case, above formula will be simplified to:
The static portion of the system H-Q curve, that part that is unrelated to the rate of flow, reads:
For closed circulating systems this value becomes zero.The total friction losses are the sum of the frictional losses of all components in the suction and delivery piping. They vary, at sufficiently large REYNOLDS numbers, as the square of the flow rate.
g = gravity (9.81 m/s2)
Hl,tot = total friction loss between inlet and outlet areas
vi = mean flow velocities trough pipe cross-section area
Ai = characteristic pipe cross-sectional area
ζi = friction loss coefficient for pipes, fittings, etc.
Q = flow rate
k = proportionality factorUnder the above stated premises the parabolic system H-Q curve can now be drawn:
The proportionality factor k is determined of the specified duty point. The intersection of the system H-Q and the pump H-Q curves defines the actual operating point.
pin, pout = pressures on suction respectively discharge liquid levels
ρ = fluid density
g = gravity (9.81 m/s2)
Hgeo = static height difference between suction and discharge liquid levels
Hl,tot = total pipe friction loss between inlet and outlet areas
vin, vout = mean flow velocities at inlet and outlet liquid level areasThe mean flow velocities at the inlet and outlet areas are, based on the Continuity Law, mostly insignificantly small and can be neglected, if the tank areas being relatively large compared to those of the pipe work. In this case, above formula will be simplified to:
The static portion of the system H-Q curve, that part that is unrelated to the rate of flow, reads:
For closed circulating systems this value becomes zero.The total friction losses are the sum of the frictional losses of all components in the suction and delivery piping. They vary, at sufficiently large REYNOLDS numbers, as the square of the flow rate.
g = gravity (9.81 m/s2)
Hl,tot = total friction loss between inlet and outlet areas
vi = mean flow velocities trough pipe cross-section area
Ai = characteristic pipe cross-sectional area
ζi = friction loss coefficient for pipes, fittings, etc.
Q = flow rate
k = proportionality factorUnder the above stated premises the parabolic system H-Q curve can now be drawn:
The proportionality factor k is determined of the specified duty point. The intersection of the system H-Q and the pump H-Q curves defines the actual operating point.
System Curve
The system performance curve is composed of a static and a dynamic component.Hsystem = Hstat + Hloss(Q)The static component Hstat is independent of the flow velocity (and thus of the flow rate). It contains the geodetic height difference as well as the pressure difference between suction and pressure vessel or inlet and outlet point of the system under consideration. In closed circuits (e.g. heating circulation) the static heed is always zero.The dynamic part of the performance curve describes the piping losses, which depend on the flow rate. In the case of turbulent flow of NEWTON liquids with constant loss coefficients of the system components, the characteristic curve results in a square parabola. If the static head and the duty point are known, the system curve can be shown with sufficient accuracy.
Suction operation with normal priming centrifugal pumps
This means that the local air pressure p<sub>b</sub> is higher than the product of the holding pressure head HH and the vapour pressure and makes a supply pressure at these temperatures unnecessary. This correlation is causally related to the drastic decrease in vapour pressure at cold water. In practice this means:Pumps with negative minimum head H<sub>req</sub> are able to operate in suction mode (not self-priming).The size of the suction capacity corresponds approximately to the value of the negative minimum suction head minus 1m safety range.Since the pumps normally used in building services engineering do not normally self-priming, the following conditions must be met to ensure suction operation:
- Filling and venting of the suction-side pipeline including the pump before commissioning.
- Prevention of air intake during pump operation (in case of air pockets, collapse of the suction function).
- Prevention of the suction line running empty when the pump is at a standstill by using a foot valve (danger of leakage in case of contamination).
Pump Selection
It is highly recommended to always select the smaller pump if the specified system duty point is between two possible pump curves. The resulting capacity reduction has, in heating systems, no appreciable effect on the effective heating performance. The positive effects are lower noise levels, lower investment costs and improved economy. For heating installations it is customary to undersize pump capacities up to 10% below the specified duty.To avoid Cavitation (vapour formation within the pump) it is necessary to maintain at the pump suction port an adequately high positive pressure (static head) in relation to the vapour pressure of the fluid being handled. The minimum required inlet heads for Glandless pumps are generally listed in pressure charts. Glanded pumps require calculations in accordance with the NPSH information.
System NPSHavailable
NPSHavailable = NPSH of system
pe = Pressure available system inlet liquid level
pb = Barometric pressure
pD = Vapour pressure of pumped fluid at the pump suction inlet
ρ = Density of the pumped fluid at the pump suction inlet
g = Local gravitational acceleration (9.81 m/s2)
ze = Static level difference between system inlet liquid level and a reference level, the negative sign becoming applicable if the reference level is above the system inlet liquid. The reference point is the centre of the impeller.
Hv = Friction loss in suction-side system.The reference point for the NPSH value is the centre of the impeller, i.e. the intersection of the pump shaft axis with the plane perpendicular to it through the outer points of the blade leading edge.The duty point of a centrifugal pump can only be a permanent duty point if complying with:NPSHavailable > NPSHrequired + safety margin.
NPSH required
The NPSHrequired is the lowest NPSH value at which definite cavitation criteria (i.e. wear due to cavitation, vapour formation, vibration, noise, head loss) can be contained.As a function of the volume flow Q, NPSHreq is a characteristic of the centrifugal pump and is specified for many types as pump characteristic curve NPSH(Q). At low volume flow rates, the NPSH value is almost constant, whereas it rises steeply at high volume flows.The NPSH value of the pump changes with the speed as well as the impeller diameter.For some pump types, the NPSH value can optionally be reduced by an additional construction. A typical example of this is the inducer, in which an axial impeller with a small number of blades is arranged directly in front of the actual impeller of the centrifugal pump.
NPSH
It is calculated from the absolute energy level minus the evaporation pressure level. The evaporating head shall be calculated with the evaporating pressure corresponding to the temperature prevailing in the inlet cross-section of the pump.The NPSHavailable value is the system-specific NPSH relating to the given flow rate and the fluid characteristics.The NPSHrequired is the lowest NPSH value at which definite cavitation criteria (i.e. wear due to cavitation, vapour formation, vibration, noise, head loss) can be contained.