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Impeller Trimming

Performance curves for centrifugal pumps normally detail the performance for a number of impellerdiameters between a maximum and minimum allowable diameter. Impeller trimming means thereduction of the impeller diameter from maximum usually to adjust the pump performance to a requiredduty point.

Several things can happen when the impeller vane diameter is reduced. The diagram below will be usedthroughout this paper.

Gap "A" describes the clearance between theimpeller shrouds to the volute or casing and

Gap "B" describes the clearance between theimpeller vanes and the casing or volute.

"D" describes the diameters of the vanes andshrouds

1. CHANGE IN PUMP PERFORMANCE

The change in pump performance with changes in impeller diameter can be predicted similarly to thatwith speed change utilising the Affinity laws

a) Pump flow rate (Q) varies directly with the diameter (D)i.e. Q1/Q2 = D1/D2

b) Pump head (H) varies with the square of the diameter (D)i.e. H1/H2 = (D1/D2)

c) Power absorbed varies with the cube of the diameter (D)i.e. P1/P2 = (D1/D2)

These relationships (not laws) allow adjustment of the H-Q curve but there is a detrimental impact on

efficiency especially for impeller reductions greater than 10% of maximum.

There are several reasons why this is true:

The affinity laws assume the impeller shrouds are parallel. This is true only in low specific speedpumps.

There is increased turbulence at the vane tips as the impeller is trimmed because the shroud tocasing clearance (Gap "A") is increasing. This is sometimes referred to as "slip".

The liquid exit angle is changed as the impeller is cut back, so the head/capacity curve becomessteeper.

Note that mixed flow impellers are more affected than low specific speed, radial vane impellers (high

head/low capacity). Simply put, the greater the impeller reduction and the higher the specific speed ofthe impeller, the more the pump efficiency will decrease with impeller trimming.

More accurate information can be obtained from a complete performance chart with different impellerdiameters detailed.

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2. NPSH IMPACTS

In general, impeller diameter reductions greater than 5% to 10% of the maximum will increase theNPSHR (net positive suction head required) dependent on impeller specific speed.

The interaction of the physical geometry of the pump inlet, inclusive of the casing, impeller, and allassociated wetted parts within the inlet field of flow determines the NPSHR characteristic of a pump. Thevalue of NPSHR for any centrifugal pump is determined through performance testing. From NPSH testdata, Suction Specific Speed (S) is calculated using the following equation, where Q represents flow atthe best efficiency point of the pump.

S = rpm x Q

NPSHR

In can be seen in the above equation that NPSHR should not change with changes in impeller diameter

as long as flow and RPM remain constant. There is no factor in the S equation that relates to impellerdiameter. Suction specific speed (S) remains constant, for any defined inlet geometry, as long as thefield of flow into the impeller eye is not disrupted by events taking place downstream of the impeller inlet.

When trimming impellers on pumps that are of a low specific speed (Ns < 30 SI, 1500 US), tests haveshown that there is little effect on NPSHR within the allowable impeller cut range. Beyond the allowableimpeller cut range, recirculation between impeller discharge and the impeller inlet start to disrupt the inletfield of flow, increasing the NPSHR.

For low Ns applications, full diameter NPSH values may be used for estimating NPSHR for cut impellerperformance. For applications with Ns values above 30 (1500 US), a NPSH test is recommended todetermine the NPSHR for any impeller trim.

3. MECHANICAL IMPACTS

i) Excessive shroud to casing clearance (Gap "A") and the resultant recirculation to the lowpressure side of the pump will produce "eddy flows" around the impeller causing low frequencyaxial vibrations that can translate to mechanical seal problems. This can be a real concern inlarge pumps with powers over 200 kW or pumps pumping heads in excess 200 metres.

ii) For many years pump people have been machining the vane tips to reduce the vane passingfrequency vibrations (Gap "B") while carefully maintaining Gap "A". The pulsating forces actingon the impeller can be reduced by 80% to 85% by increasing gap "B" from 1% to 6%.

iii) For impeller diameters up to 355 mm, gap "B" should be at least 4% of the impeller diameter toprevent "Vane passing syndrome cavitation" problems. For impeller diameters above 355 mm,

Gap "B" should be at least 6% of the impeller diameter to prevent this type of cavitation. Thistype of cavitation damage is caused when the outside diameter of the impeller passes too closeto the pump cutwater. The velocity of the liquid increases as it flows through this small passage,lowering the fluid pressure and causing local vaporisation. The bubbles then collapse at thehigher pressure just beyond the cutwater. This is where volute damage occurs. Unless damage

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has penetrated to the outside of the volute, a flashlight and mirror will most likely be required toinspect the damage. The damage is limited to the centre of the impeller vane. If it's a closedimpeller, the damage will not extend into the shrouds.

iv) Although both the vanes and shrouds are often cut in end-suction, volute-type centrifugal pumps,it is not a good idea to do this in double suction designs. With these types of pumps you canreduce the vane diameters, but the shrouds should remain untouched.

v) Structural strength is a consideration when deciding how much to reduce the vane diameter in

double ended pumps because you could leave too much unsupported shroud. Somemanufacturers recommend an oblique cut that will improve the vane exit flow and add somestrength to the shrouds.

vi) Machining a radius where the trimmed vane meets the shroud is another good idea to addstrength to the assembly. Square corners are never a good idea.

4. NOISE

When writing a pump specification, many practicing engineers limit the impeller diameter to 85% of itsmaximum diameter. Such a limitation is actually a misunderstanding of a design concept known as"quiet pump operation." This misunderstanding may force the selection of a larger pump for the

application. The idea here is not that the impeller diameter should be 85% of the maximum publisheddiameter, but 85% of cutwater diameter (0.85 cutwater ratio). To fully understand the quiet pumpoperation design concept, refer to figure below.

In designing a pump casing, a design engineer first determines the volute scroll (A) necessary to handlethe desired volume of water. This volute scroll terminates at the volute cut water (B) at the base of thedischarge nozzle (C). The volute scroll is drawn around a base circle (D), which is sufficiently largeenough to allow insertion of the impeller. The distance from the shaft centerline to the volute cut water iscalled the cutwater radius and twice this distance is the cutwater diameter

Hydraulic noise becomes a factor when the periphery of the impeller passes too close to the cutwater. Indesigning a pump, the distance between the impeller and the cutwater is a compromise between thepump efficiency and pump noise. Typically, cutwater ratios (D/F) of 0.9 and above produce higher noiseand cutwater ratios of 0.8 and below produce significantly lower pump noise. Cutwater ratio of 0.85 iscommonly specified by practicing engineers, thereby realising a minimum reduction in efficiency with amean reduction in noise level.

From the above, it may be understood that a specification should more properly read "impeller diameternot to exceed 85% of the volute cutwater diameter rather than "impeller diameter shall not exceed 85%of the maximum impeller diameter capable of being installed in the pump casing."

Specifying the later statement is safer since the impeller diameter would be even smaller than thedesired maximum. In some cases this may force selection a larger pump than necessary.