Before pump design or selection can begin, specifications need to be established which express several requirements. Only requirements pertaining to the hydraulic performance of the pump are considered here. Other considerations are reliability, durability, noise and vibration, maintenance, installation, and control. These concerns affect primarily the mechanical design of the pump: material selection; auxiliary component specifications, such as bearings, seals, and instrumentation; and design for ease of assembly, service and maintenance.
Among the specifications affecting hydraulic performance, the most basic are the nominal efficiency, speed, flow rate, pressure or head, and power. Efficiency is important from the point of view of energy cost. Because of the maturity of pump design technology, the nominal efficiency of commercially available, competitive pumps is within 1 or 2% test data rarely being more accurate than I%. Power losses arising from piping system design and losses due to control considerations can amount to 30% or more of the energy supplied by the pump. In process applications pumps are often followed immediately by control valves, which produce a head loss of 20 or 30% at nominal conditions. Flow control is achieved by increasing or decreasing the head loss.
Considerable energy savings can be achieved by variable-speed drives, for example.
Consequently, energy conservation efforts should preferably be directed at these losses rather than at 1 or 2% pump efficiency improvements. The competitive advantage of high efficiency therefore appears to reside in reflecting high technical competency on the part of the manufacturer rather than in energy savings.
If the power of the drive motor is limited, nominal efficiency might also be a consideration. Depending on the application, motors are usually selected assuming a certain service factor, which multiplies the nominal power requirement of the driven equipment. A factor greater than l assures a reserve power margin. Actually, maximum temperature rise limits the ultimate power that an electric motor can deliver. In instances where the pump and motor are sold as an assembly, as in appliances for example, pump and motor are closely matched, assuming a service factor of 1. The motor load can reach the ultimate power, limited only by the permissible temperature rise. In such a case the pump efficiency determines the maximum pressure or flow rate
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that can be obtained from the assembly with a given motor.
Pump speed will depend primarily on the driver. Electric motor speeds follow the network frequency: 60 Hz in the United States and 50 Hz in Europe, corresponding to nominal 3600- and 3000 rpm synchronous speeds of single-pole motors. Smaller motors, up to perhaps 100 hp, usually turn at 3600 rpm. Larger motors turn at an even fraction of the synchronous speed: 1800, 1200, 900, or 720 rpm, depending on the number of poles of the motor.
Large pumps have large inlet diameters. The impeller inlet circumferential velocity increases with the diameter but must stay in reasonable proportion to the inlet flow velocity to avoid blade angles too much inclined toward the tangential direction. The danger of cavitation also increases with the velocities at the inlet. These reasons suggest lower rotational velocities for large pumps. Under full load, motor speed typically decreases 2% from the synchronous speed. At a price, variable-frequency drives are available, some of which can reach 5400 rpm, but the synchronous speed limits the maximum speed of most induction motors. Internal combustion engines drive pumps in remote, outdoor, often mobile applications, such as irrigation, dewatering, and flood control. Their rated speed varies but remains relatively low compared with that of electric motors, on the order from 1000 to 2000 rpm. Their speed is usually governed to remain constant regardless of the load.
The rated flow rate of pumps can match the required flow directly since leakage losses are normally insignificant. The nominal flow rate also determines the inlet and outlet pipe diameter and flange size, which are standardized. The code number of commercial lines of pumps often includes the inlet and outlet pipe diameters measured in inches. Pipe and flange diameters are sized to result in water flow velocities of 5 to 15 ft/sec (1.5 to 5 m/s), but velocities can reach 30 ft/sec (10 m/s), in large high-specific-speed pumps.
Specifying the pressure or head rise produced by the pump requires the most care. It is defined as the total head, static and dynamic, at the pump exit flange minus the total head at the inlet flange. The nominal head required by the system must be estimated by keeping in mind the maximum possible head that might ever be encountered. A reserve margin may be necessary. However, an excessive margin may make the pump operate normally at low efficiency or at some adverse condition. Most pumps are capable of operating over a range of higher or lower flow rates and pressures on either side of the nominal. Such off-design operation, at exceptional conditions, should be taken into account when estimating the nominal head, in order to minimize the need for a large
reserve margin.
The head rise required from the pump is estimated by adding up the suction head at the inlet, including pipe friction losses, valve and pipe friction losses at the exit, and the net head of the final load, such as the elevation head to an overhead tank (Fig. 5.1). If the pump inlet is below the water level of a sump or supply tank, the inlet head will be positive, from which the inlet piping losses, calculated at the rated flow rate, need to be subtracted. If the pump inlet is above the water supply level, the inlet head is negative and so are the additional inlet piping losses. The valve and pipe losses at the exit, also calculated at the rated flow rate, as well as the net head of the load, add to the head requirement of the pump.
Particular attention must be given to the pump exit velocity head. The rated pump head rise usually includes the velocity head of the flow leaving the pump. In performance testing the pressure rise is measured "total to total," implying that the velocity heads at the inlet and exit flanges are added to the respective static head measurements to arrive at the total head at each location. This approach assumes that the kinetic energy of the flow leaving the pump remains useful, which may or may not be the case. If the pump produces a free jet, like a fire hose, then the velocity head can indeed be counted as useful energy. If the pump discharges into a stagnant tank, for example, the kinetic energy is lost and must be added to the piping losses at the pump exit.
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Figure 5.1 Pump system installation.
In addition to specifying the rated efficiency, speed, flow rate, and head, several other requirements need to be established. First, the pump must have sufficient suction capability without cavitating. Verifying satisfactory suction performance requires two steps: the suction head the net positive suction head available (NPSHA) from the system-must be estimated, and a comparison must be made with the net positive suction head required (NPSHR) by the pump. Pump catalogs list the NPSHR of the pump as a function of flow rate. The NPSHR values reach a minimum at the design flow rate and gradually increase at higher flow rates. The absence of cavitation must be assured at the greatest anticipated flow above the rated flow rate. Catalog data are usually obtained at conditions conforming to testing standards, typically when the pump head declines by
3 % with decreasing inlet pressure. However, cavitation vapor bubbles may appear earlier and cause damage to the impeller. Conditions in the actual application may differ from the standard testing conditions. Excessive aeration of the water ahead of the pump, supply sump vortexing, or air leaks into the suction piping may cause pump cavitation at inlet pressures that would appear sufficiently high according to catalog data. Therefore, the NPSH required should be estimated conservatively, and a higher inlet pressure than that absolutely necessary should be provided to the pump.
Often, further requirements are imposed, such as the minimum flow rate at which the pump may have to function, the head at shutoff, and the maximum power demand.
Several undesirable phenomena can appear at reduced flow rate: general instability of operation, inlet recirculation, exit recirculation, cavitation, and overheating. Details of such conditions are discussed in subsequent chapters. Some of these can be tolerated in case of an emergency but would lead to deterioration and failure of the pump if they were to persist for longer periods. The application will decide how often and how long the pump may operate at reduced flow and what the cost trade-off might be between a more expensive pump and damage from an emergency of low probability.
Manufacturers offer guidelines for the safe, minimum flow of their pumps (Cooper 1988).
Instability at reduced flow results from a positive slope of the head-flow curve of the pump. It depends not only on the pump characteristics but also on the system load curve, as described further below. It causes severe torque and power fluctuations as well as vibration and noise, which can damage the bearings or shorten their life. Inlet and exit recirculation may or may not create problems. It certainly disrupts the normal flow pattern in the pump, but can often be tolerated for a limited time. The flow rate corresponding to the onset of inlet and exit recirculation can be calculated with the procedures and computer programs given in this book. Recirculation can also combine with cavitation, which might aggravate vibration and power fluctuations, and will result in progressive cavitation damage. Recirculation is more likely to precipitate cavitation if the calculated NPSHA does not comprise a sufficient safety margin.
Overheating results when excessive losses, transformed to heat, are retained in the pump. Pumps handling cold water rarely encounter problems. However, a temperature rise will affect the vapor pressure of the fluid. When pumping hot fluids near boiling, cavitation may appear unexpectedly. The head at shutoff is difficult to estimate because the flow churns completely uncontrolled in the pump casing. It also depends on the configuration of the piping upstream from the pump. An approximate estimate can be obtained by extrapolating the performance curve to zero flow. The pressure at shutoff is
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specified, for example, in the case of fire pumps. Such a specification may intend to make certain that the head-flow curve does not have a range of positive slope, which would lead to instability.
If the power demand of the pump decreases with increasing flow rate, a reserve power margin will be needed if the pump is likely to operate below the flow rate at the rated point. The opposite holds for pumps whose power demand generally increases with the flow rate. Pump characteristic plots in catalogs show the power demand of the pump as a function of the flow rate. If the plot does not show the power directly, it can be calculated from the head, flow rate, and efficiency.