A good thing? At first glance the answer is obvious, of course! closer examination of the facts reveals a different story. In order to reduce the NPSHr, the pressure in the eye of the impeller will have to be increased. As NPSHr is a pump condition, none of the external factors associated with NPSH available come into play, enter one Mr Bernoulli. This gentleman stated that as the velocity of a liquid increases, its pressure will decrease and conversely, as the speed of the liquid decreases, then the pressure will increase. Problem solved, because all that needs to be done for a given flow rate is to increase the cross sectional area that the fluid has to pass through and this will ensure a drop in velocity and a concomitant increase in the pressure. From all of this, if the impeller eye area in a centrifugal pump is decreased, then the pressure will increase, this means a reduction in the pump's NPSHr. Take a look at the two diagrams below.

 Figures 1 & 2 show two impeller designs which deliver the same head and capacity. Impeller 1 has a smaller eye than 2 and from bernoulli's law, impeller 1 will have a lower NPSHr due the the increased eye pressure. The lower NPSHr comes at a cost however as figure 2 now shows. By opening the impeller eye,the designer has increased the risk of recirculation occurring at the vanes as shown in Figure 2. This is fertile ground for the formation of cavitation with all the vibration, metal removal and other problems associated with this condition.

By solving one problem, a potentially far larger one has been created. The further left of the Best Efficiency Point (bep) the pump is operated, the more exaggerated the problem becomes. So how does the calculation of suction specific speed help in avoiding the problem? As was shown in part one of this series, Nss numbers higher that 175 indicate a pump which is prone to generating suction recirculation cavitation. Nss numbers of 175+ mean that the designer and the end-user have to closely examine the TDH range that the pump will "see”. If there is a possibility that the duty point could move less than about 80% of the flow rate at bep, then caution is advised.

It has been said that due to advances made in the Computational Fluid Dynamics field, the problem of high Nss pumps failing due to recirculation cavitation has been largely done away with. This might have a measure of truth in it but how many pump ranges that are currently on the market were designed in the seventies and early eighties? If NPSHr is a major design criterion in your next pump system, beware of bearers of “good” (NPSHr) news!

During a recent visit to a large chemical plant, I was intrigued by the unusually loud knocking noise which was emanating from three reasonably large (350kW) cooling water end suction pumps. A non-technical colleague who was with me at the time even remarked that the noise was unusual.

Closer inspection revealed an intermittent but severe knocking noise in both the the suction and discharge, to the degree that the rubber compensators were showing evidence of rapid internal pressure transients. Classic cavitation symptoms caused by operation too far left of the Best Efficiency Point (BEP) combined with a high pump suction specific speed number(NSS)!

The plant supervisor confirmed that it was non-standard practice to run all three pumps at one time as the system was designed for a two operating one standby basis. That at least explained the operation on the left hand side of the pump's curve (see last posting on system head curves).

The pumps in question were designed for a high capacity/low head type of operation where NPSH required becomes something of an issue. In order to meet the challenge, the pump designers enlarged the impeller inlet area to reduce velocity and use Bernoulli's theorem to increase the pressure in the eye of the impeller. This solved the problem of having an NPSH required which was too high, but created another problem in the form of recirculation of flow from the higher pressure areas of the pump into the impeller eye. The resultant drop in pressure often reaches a point which is less than the liquid's vapour pressure. This allows the formation of vapour pockets which implode as they reach areas in the impeller eye where the pressure has increased.

Getting back to the NSS, the Hydraulic institute recommends a number less than 175 (metric) with design values higher than this increasing the chance of pump failure exponentially. The formula used for calculating NSS is:

Where: Q = flow rate in cubic metres per sec at BEP

NPSHr = Net Positive suction head at the same impeller diameter and flow rate

Part 2 of this blog will cover the relationship between impeller eye area and NPSHr.

What is a system head curve? Simply put, it is a graph which shows the total dynamic head (TDH) in a pipework system at various flow rates. Often this is superimposed on the performance curve of an appropriate pump and the duty point will be at the intersection of the system curve and the pump's performance curve. If the static head varies for any reason, a second system curve can be drawn which shows the increased TDH. The two intersection points show the range of flows and heads the pump will "see" as the static head changes.

Where does this all originate from? Thanks must go to a colleague who took the time to take this trainer through the process (too) many years ago. At that time the light went on and, yes, somebody was found to be at home. The lesson stuck and whenever the suspicion arose that the system was the culprit, out came the graph paper (this was before Excel and other dedicated software) and the friction loss charts. Once the plotting process began the story unfolded before some eager eyes. Yes the pipe was too small or in some cases too big!

Once all the plotting was complete, there it was, in black and white or in colours, the whole sad story. No more arguing or debate. The solution(s) became pretty self evident and the feeling of achievement and satisfaction were palpable.