Reducing costs in parallel pumping

Transcription

1 32 Feature Operating Reducing costs in parallel pumping Pumps operated in parallel can be a mixed blessing. A well thought-out design, correctly operated, can do wonders for handling variations in duty point but poor operational procedures can be a source of excessive energy consumption, significant maintenance overspends and poor system reliability. John Tonkin discusses the details of one such installation. Many pumping processes require an uninterruptable flow of fluid to ensure the safe and cost-effective production of an end product. A stoppage in flow can often lead to very serious quality and/or safety issues; the recent occurrences at the Fukushima nuclear power plant are a case in point. To this end, many installations make use of one or two duty pumps while one extra unit is held as a stand-by in case of an operating pump failure. This arrangement has a number of obvious benefits but, as with any other system, it has to be operated in accordance with the original design intent. In a surprising number of instances, the original plan is ignored and the stand-by pump is run continuously, in parallel, with the operational units. This means that the pipe work system is subjected to the flow of more pumps than it was originally designed for. For Figure 1. Three 350 kw WBZ 350/400 EN 733 cooling water pumps operating in parallel / Elsevier Ltd. All rights reserved

2 Feature 33 a one operating, one stand-by system the pipe work has to accommodate a 100% increase in capacity. Due to the system resistance in the pipe, the increase in flow is converted into friction losses, which lead to an increase in the total dynamic head (TDH). In many cases this forces the pumps onto the left-hand side of their operating curves, which can have serious implications for efficiency, power consumption and reliability. This article discusses the negative effect of operating stand-by pumps on a continuous basis. It is based on the findings of investigations carried out on two installations, one with three 45 kw end-suction pumps installed in a mining operation and the other with three 350 kw units used in a chemical plant (Figure 1). Using systems head and pump performance curves, and accepted practices, the potential for significant operating cost increases through lower efficiencies, and for lower mean times between failures as a result of re-circulation cavitation, will be examined. Case studies The use of parallel pumping systems has found favour in addressing two ever-present problems facing a system designer. Firstly, the issue of reliability can be addressed by using two or more pumps operating on a continuous basis and at least one unit held in reserve as a stand-by. In the event of a breakdown occurring with one of the operating pumps, the stand-by can be brought into operation. Secondly, variations in flow rate can, to a degree, be accommodated either through manual interventions or automatically through the use of pressure or flow sensors. However, if the system is designed around a specified number of operating pumps and the remaining pump(s) as stand-by, then significant problems could arise if all the units are operated all the time. This seems an obvious statement. However, two separate installations (designed as two operating, one stand-by) were recently observed to be operating all three units continuously. In both cases, the net result was a significant drift to the left of the head/capacity (h/q) curve. In both, the system efficiency declined dramatically and the energy cost increased substantially. It was also noted in the second and larger of the two installations that reliability reached unacceptably low levels. First installation The first of the two installations mentioned above made use of 3 x 45 kw pump sets Figure 2. Original pump test curve. (Reproduced with permission from Sterling SIHI GmbH.) running at 1,450 rpm. The system was designed to run in a two operating, one stand-by configuration. As part of an infield practical assignment for a pump training programme, one of the pumps was switched off and the flow and TDH were monitored. Flow rate was found to decrease by 9%, which moved the duty point significantly closer to the best efficiency point (BEP), while TDH declined a mere 6%. Further investigation of the performance data for the installation spanning the previous six months revealed that: All three pumps ran continuously irrespective of the level of demand. On average, the duty point for each pump fell within the last 14% of the performance curve for the majority of the time. System efficiency was severely compromised but little was done about this situation as the company in question had negotiated a very favourable electricity rate (2.7 US cents per kwh) with the service provider. Further analysis of the historical data revealed that one pump would be sufficient to supply the needs of the plant for 76% of the time. If a suitable control system were installed, the cost would be relatively low while a saving in electricity of at least 18% could be made. Second installation The second installation has 3 x 350 kw WBZ 350/400 EN 733 (previously DIN 24255) end-suction, back pull-out pumps operating in

3 34 Feature Figure 3. Screen shot of the system data. parallel (Figure 1). The original system design was for two pumps running with the remaining one as a stand-by. These units were manufactured in 1979 and it is unclear when the decision to operate all three pumps together was made. The decision has brought with it a number of consequences, the majority of which are very negative from both an energy consumption and a maintenance cost point of view. A more-detailed description of the challenges facing the end-user of these three pumps in the second installation will be the focus of this article. Methodology Background I was fortunate enough to obtain a copy of the test curves (Figure 2) for these three pumps after 32 years, despite the fact that this German company has merged with a UK-based pump manufacturer! Using these curves, a comparison was made between the existing performance of the system and that of the original design. This comparison brought to light the substantial saving that could be made by instituting a programme that would restore the system back to its original design criteria. The system comprises a cooling tower, the three pumps, a surface steam condenser and an adsorption column. Figure 3 shows a screen shot of the system and the applicable performance data. The control of the temperatures in the condenser and adsorption column are extremely critical and variations of temperatures outside a specified range will lead to immediate plant shutdown and/or serious injury or death. Ambient temperatures, which vary considerably between summer and winter, are a constant source of concern and a close watch has to be maintained on the discharge and return water temperatures. In discussions held with the operations staff, there was concern expressed about the efficiency of the cooling tower. The process of cleaning out the pipe work that

4 Feature 35 discharges the warm (return) water onto spreader nozzles has led to the damage or complete destruction of some of these spreaders. The water curtain contains a higher-than-normal percentage of unbroken water streams, which reduces the water surface area exposed to the updraft created by the fans. Potentially, this could explain why operations had brought the third pump into service. Pump/system efficiency The discharge pressures and flow rates were measured using calibrated pressure gauges and the flow meters fitted to the two branch lines that feed the surface condenser and the adsorption column. From these figures the water (output) power was calculated. Real motor power was ascertained with the aid of a fluke power analyser using a three-phase, four-wire system. From published motor efficiency data and discussions with a reliable source in the motor repair industry, the shaft power was calculated. It was then possible to calculate the pump efficiency. the current duty point was superimposed (Figure 4). From the flow meter and pressure gauge readings, the current duty point for all three pumps operating is 3,928 m 3 /h at a TDH of 44 m. Figure 4 shows that the performance of the three pumps has deteriorated significantly despite a number of overhauls being carried out over the years. The original design called for a flow of 3,800 m 3 /h from two pumps with a third held as a stand-by in case of failure of any of the operating units. Figure 4 also shows that this reserve is no longer available and a failure in any one of the three pumps will lead to the immediate shutdown of the plant. Pump efficiency Using the power analyser, the real motor power was found to be a total of 681 kw for all three motors. With reference to the test curves in Figure 2, the tested power absorbed at the duty point is 300 kw per pump, which gives a total of 600 kw for two pumps. This means that to deliver the required quantity of cooling water now requires all three pumps to be operated with an additional energy requirement of 81 kw. Cavitation & system reliability Two of the pumps show symptoms of suction recirculation cavitation, namely, intermittent, Acoustic symptoms The presence of cavitation in pumping equipment is characterized by sounds that can be likened to the movement of coarse gravel through the pump. The sounds emanating from each of the pumps were carefully monitored for the presence of cavitation. Of particular interest in this type of situation is suction recirculation cavitation, which occurs when high-capacity pumps with high suction specific speed numbers are operated to the left of the BEP. Because of the hydraulic characteristics of these pumps (high flow, moderate head), this installation was a prime candidate for this fault condition to occur. Results Availability of stand-by capacity Using the test curves supplied from Germany, a set of curves showing the performance of three new pumps was constructed and

5 36 Feature Figure 4. Head/capacity curves for three new pumps in parallel with current duty point superimposed. clearly audible knocking sounds in the area of the impeller eye. The WBZ 350/400 (350 mm discharge flange and 400 mm nominal diameter impeller) can be classified as a high-capacity, moderate-head pump. In order to achieve this high capacity while still achieving a relatively low NPSH required, the eye of the impeller has to be increased. This has implications for recirculation in the impeller eye at duty points to the left of the BEP. The effect that this has on the three pumps in this installation will be dealt with later in the Discussion section. Suction specific speed (Nss) for these pumps was calculated as follows: RPM x Q Nss = NPSHr 0.75 where RPM = pump rated shaft rotational speed (rpm); Q = flow rate at the BEP (m 3 /s); and NPSHr = pump net positive suction head at maximum impeller diameter at the BEP. 1,485 x 0.53 Nss = Nss = The American Hydraulic Institute (AHI) states that for pumps with Nss values greater than 175 (SI units) the probability of repeat installation failure increases exponentially. For values higher than 180, sustained flow rate should not be allowed to be reduced below 85% of BEP. The Nss value for these pumps is 40% higher than this recommended value, which indicates that the duty point should be held as close as possible to the BEP. Costs Discussions with management revealed that the average duty cycle for these units is in the region of 7,440 h per annum. The average cost of electricity for this company is 5.3 US cents per kwh. The additional 81 kw absorbed due to the reduced system efficiency means an additional annual expenditure of: Additional energy cost = 81 x 7,440 x Additional energy charge = US$31,940 Feedback obtained from the maintenance department reveals that the pumps needed replacement impellers at least every six years and in a number of cases every three years. While an acceptable impeller life span is difficult to quantify, Vlaming 1 bases calculations for an acceptable level of cavitation on an impeller life span of 40,000 h. This was related to the ratio between the NPSHa and the NPSHr. Henshaw 2 compared this approach to that taken by Grist 3 and showed that centrifugal pumps, particularly those with high Nss values, operating on the left-hand side of the curve are prone to severe erosion of the impeller material due to cavitation. If this operating time span is accepted it must be compared to this installation s record of 6 x 7,440 h or 44,640 h, which is acceptable, and 3 x 7,440 h or 22,320 h, which is close to half the expected life and, as such, is not acceptable. Discussion The sustained flow of cooling water from these three pumps is critical to the efficient and effective operation of a plant valued in the hundreds of millions of dollars. Any interruption of flow will lead to very significant expense in the form of lost production, injury or loss of life. The operation of pumps in parallel outside the original

6 Feature 37 design parameters will inevitably lead to increased costs and reduced plant effectiveness. The dangers of operating centrifugal pumps on the far left-hand side of a performance curve have been well documented in publications such as Europump s System Efficiency guide 4. By operating all three pumps together, increased friction losses ensured that the duty point for each pump was forced significantly further to the left of the BEP than was originally designed for. This gives rise to the onset of cavitation, inevitably leading to impeller damage and vibration and radial reaction forces, which combine to shorten bearing life. The remediation of this installation could take the form of replacing the pumps with new units and reverting back to a two operating, one stand-by format. This presents some problems as the replacement units would have to be EN 733 type pumps as there is insufficient space in the pit for the installation of other kinds of pumps such as horizontal split case units. There are not many manufacturers capable of offering pumps of this size so lead times and pricing could be a significant problem. Research has shown that spares for these pumps are no longer available. If this installation is to be bought back to its original level of performance, the pumps will have to be replaced. It is imperative that the operations team is fully informed of the dangers inherent in the sustained operation of the three pumps at one time. This will require the drafting/redrafting of operating procedures in order to ensure that this situation does not continue into the future. Consideration can also be given to the installation of appropriate controls that would prevent the operation of more than two pumps at any one time. This would be a straightforward procedure that could be completed at minimal cost. During the investigation into the workings of this system, it became apparent that variations in flow rate are small and are unlikely to require the starting of the stand-by unit. To this end, operational staff should be given the necessary knowledge and skills to be able to operate these pumps to the original design. Conclusions The operation of more pumps in parallel than was designed for in a given system will inevitably lead to increased energy, maintenance and downtime costs. The positive aspect of operating pumps in parallel is that significant cost savings can be made by operating the system strictly in accordance with the original design specification. n References [1] D.J. Vlaming, Analysis of Cavitation Provides Advanced NPSH Estimates for Centrifugal Pumps, Oil & Gas Journal, November 19, (1984). [2] T. Henshaw, What is a safe NPSH for a centrifugal pump? Can you provide too much NPSH?, Pumps & Systems, June, (2010). [3] E. Grist, Net Positive Suction Head Requirements for Avoidance of Unacceptable Cavitation Erosion in Centrifugal Pumps, Paper No. C163/74, IMechE conference publication CP , (1974). [Available from the UK Institution of Mechanical Engineers; library@imeche.org] [4] Europump, System Efficiency A guide for energy efficient rotodynamic pumping systems, Europump, (2006). This paper was first presented at the 2011 International Pump User Conference (IPUC), held in Johannesburg, South Africa, September 2011, and is reproduced with permission from the organizers. Contact John Tonkin John Tonkin and Associates PO Box 5081, Weltevreden Park 1715, South Africa Tel: johntonkin@mweb.co.za ROTEX with T-PUR The new standard HAN N OVE R FAI R April 23 27, 2012 Hall 17 / booth B40 Hall 27 / booth H54 T-PUR the new heart of our ROTEX offers more and extends the service life. Particularly in critical or highly loaded working environments. The T-PUR spider material resists peak temperatures of ice-cold -50 C up to a red-hot +150 C. And it still keeps it s elasticity and hardness at no extra cost. Sustained saving. Without compromise.