1. Condensate basics
2. Two values of NPSH
3. Calculating NPSHA
4. Increasing NPSHA
5. Condensate pump design
6. Summary
7. References and Bibliography

Anyone who has experienced cavitation damage to a pump impeller, shaft, or seals knows that problems relating to net positive suction head (NPSH) can also mean “not pumping so hot.”¹ To prevent damaging cavitation, it is necessary that engineers clearly understand the critical importance of NPSH when selecting pumps for steam condensate systems.

This article also discusses the various options available to an engineer to adjust the pertinent factors of NPSH to ensure proper system operation. The types of specially designed low NPSH pumps for condensate service are also reviewed along with their benefits and drawbacks.

Condensate basics

Steam condensate systems operate at elevated temperatures. Condensate typically will be returning to a boiler at temperatures of 180 to 200 F in a low pressure steam system. When a system is initially designed and installed, this temperature is easily maintained. Steam traps throughout the system prevent steam from entering the condensate return lines. However, if the traps are not properly maintained, they will eventually begin to fail and pass steam. If these traps are not replaced when necessary, steam will pass through to the condensate line where it will begin to condense. In turn, the energy in the latent heat of condensation will reduce the subcooling effect of the return piping. The result will be elevated temperatures in the condensate unit. Selecting a condensate pump for new or replacement applications requires the engineer to be aware of the importance of condensate temperature.

A critical factor that should be investigated in the selection of a condensate pump is the NPSH. This is extremely important when choosing a condensate pump due to the high temperatures in steam condensate returns. NPSH is determined by factors such as temperature, altitude, static head, and capacity. Condensate pumps are available that have been specifically designed to handle low NPSH, preventing cavitation, and thus able to operate at higher temperatures.

As liquid enters the eye of the impeller in a centrifugal pump, its pressure is reduced. If the absolute pressure at the impeller eye drops down to the vapor pressure of the fluid, vapor pockets begin to form. As these vapor pockets travel in the fluid along the vanes of the impeller, pressure increases and the pockets collapse. This collapse is called cavitation. Cavitation is not only noisy but also damages the pump impeller, shaft and seal, and over time, may reduce pumping capacity. NPSH refers to the minimum suction pressure, expressed in feet of water column, that is required to prevent the forming and collapsing of these vapor pockets.

Fig. 1 shows the change in system pressure (P_s) as the fluid travels through the impeller. To prevent cavitation, P_s must remain above the vapor pressure.

Fig. 1 Top curve shows system pressure (P_s) remaining above fluid vapor pressure as it passes through the pumps; cavitation cannot occur. Bottom curve shows P_s falling below the vapor pressure as it enters the impeller eye. This will cause cavitation. Cutaway view of a pump volute on the right shows the passage of flow through the impeller.

Two values of NPSH

There are two values of NPSH: NPSHR and NPSHA. NPSHR (required) is the amount of suction head required to prevent pump cavitation, is determined by the pump design, and is indicated on the pump curve. NPSHA (available) is the amount of suction head available or total useful energy above the vapor pressure at the pump suction. This is determined by the system conditions. NPSH typically is measured in ft of liquid.

Calculating NPSHA

NPSHA, measured in feet, can be calculated by using the equation noted below. The vapor pressure P_v of the liquid is subtracted from the system pressure P_a It is then converted from psia to feet by multiplying by 2.31 and dividing by the specific gravity of the liquid. The static head H_e, measured in feet, is determined by the elevation of the water line above the pump suction. This value can be negative if the application is a suction lift. Finally the friction losses H_f, measured in feet, are calculated and subtracted from H_e. This quantity is then added to the first term of the equation.

The following is an example of the calculation of NPSH. 100 gpm of condensate is returning at 210 F to a receiver that is vented to atmosphere. The static suction head is 2 ft. Pressure drop through the pipe is calculated for 1 ft of 2-1/2 in. pipe with a 90 deg elbow and a gate valve. The friction loss through a 2-1/2 in. iron pipe is 7.1 ft per 100 ft of pipe. The equivalent length of straight new pipe for turbulent flow has to be calculated for the elbow and valve. The 90 deg elbow has an equivalent length of 3.6 ft and the gate valve has an equivalent length of 1.7 ft. The total equivalent length is summed and multiplied by 7.1 ft per 100 ft. The calculated NPSHA for this example is 2.99 ft. The pump that is selected for this example must have a NPSHR less than 2.99 ft to prevent cavitation from occurring (Fig. 2).

Fig. 2 Example of the calculation of NPSH. Piping and appurtenance sizes are 2-1/2 in. See text for system description and equivalent lengths of the piping components used in the H_f calculation.

The vapor pressure of a liquid drops as its temperature falls. Positive head increases as the temperature decreases. Table 1 shows the relationship of temperature, vapor pressure, and positive head.

Typically, condensate pumps are designed to handle condensate with temperatures up to 200 F. The NPSH available at 200 F is 7.35 ft (Table 1). If the system return temperatures are higher than 200 F, either the NPSHA in the system has to be increased or the NPSHR of the pump has to be reduced to prevent cavitation.

Increasing NPSHA

The NPSHA can be increased by changing one or more of the factors that determine it. The temperature of the condensate can be lowered to decrease the vapor pressure. This can be accomplished by having the condensate go through either a cooler or a flash tank. However, lowering the temperature of the condensate through the use of these can result in energy losses unless this energy can be recovered.

Fig. 3 shows how energy is wasted in a typical heat exchanger (HX) and trap arrangement. In this example, 50 psig steam at 298 F enters the heat exchanger. The latent heat of condensation is used to heat the water in the tube bundle, and the condensate leaves the HX at 298 F. This high temperature condensate flows into the bucket trap. When the trap opens, the condensate drains into a flash tank, which is then vented to the atmosphere. At atmospheric pressure, the saturation temperature is 212 F. At the lower pressure, 9 percent the condensate flashes to steam. The heat of vaporization of the vented flash steam sensibly cools the remaining 91 percent of the condensate from 298 F down to 212 F. If the flash steam is not recovered for some other process, it equates to a 9 percent energy loss.

Fig. 3 Heat exchanger and trap showing flash losses.

Maintaining sufficient NPSH to avoid pump cavitation usually is not a serious factor in closed hot or chilled water systems because these systems are pressurized. The system fill pressure can usually be increased to satisfy the NPSHR of the pump. This same concept of a pressurized system can be applied to a steam condensate system. The NPSHA can be increased by pressurizing the receiver where the condensate is collected. Pressurized condensate units are available that are designed with pumps to handle temperatures up to 250 F. By using a closed system, flash losses are also avoided.

A pressurized condensate unit, however, is not always the solution for low NPSHA. If the condensate unit collects condensate from multiple returns that operate at different pressures, a pressurized unit is not appropriate. The pressure(s) in the higher pressure return line(s) will prevent the draining of condensate from lower pressure return(s). For this situation, a vented receiver must be used.

The static elevation can be increased by elevating the receiver that collects the condensate. However, if the condensate returns are low or below grade level, this factor is difficult to change. Sometimes the only solution is to put the condensate unit in a pit deep enough to provide the necessary NPSHA. In an existing system, however, there is often no way to change the level of return lines.

There are situations where the NPSHA value cannot be increased to the desired point. For these times, selecting a pump with a low NPSHR is often the only way to handle high temperature condensate. For steam condensate systems with temperatures higher than 200 F, pumps have been designed with lower NPSH values than those of centrifugal pumps used in closed hot and chilled water systems.

Condensate pump design

Pump manufacturers must seriously concentrate on the NPSH characteristic when designing pumps for condensate service. The NPSHR of a pump is specifically affected by the impeller inlet design. The impeller NPSH characteristic is based on, among other things, the inlet flow angle, which is taken at the outer diameter of the impeller eye. The inlet flow angle refers to the angle of the fluid stream entering the impeller. A larger angle gives higher efficiency; a smaller one causes a lower NPSH. A flow angle of 17 deg with approximately five to seven vanes is often used as a compromise between the two. NPSH characteristics of different pumps can be compared by their suction-specific speeds. Suction-specific speed (N_ss) is a dimensionless value that is determined by Equation 2, below.

A lower N_ss number indicates a higher NPSHR value for a given pump.

A typical range for the suction-specific speed is 7000 to 8000 for centrifugal impellers. For condensate and boilerfeed applications where low NPSH is critical, the suction-specific speed increases from 12,000 to 18,000. To reach such high values, the flow angle decreases to as low as 10 deg with as few as four vanes. The advantage to fewer and thinner vanes is a reduction in blockage effect. Blockage is caused by water vapor obstructing or choking off the flow of liquid through the vanes. While vapor formation is to be avoided, this vane arrangement can better handle it should this occur. The disadvantage of low flow angles or large inlet diameters is that the pump may run roughly at capacities below 50 percent.

To further reduce the NPSH and increase the suction-specific speed above 18,000, an axial flow impeller or inducer may be used ahead of the centrifugal impeller (Fig. 4). Its flow angle is between 5 and 10 deg. The vane angle is 3 to 5 deg larger, and the number of vanes is between two and four. The vane angle is the angle of the leading edge of the impeller vane. The difference between flow angle (fluid stream) and vane angle (impeller vane) is caused by slippage. The axial flow inducer develops a positive pressure at the eye of the centrifugal impeller. This pressure typically is 5 to 10 psig.

One of the benefits of using an inducer is that efficiency at the operating point is not sacrificed. As stated earlier, the inlet flow angle affects the efficiency and NPSH. Since it allows the centrifugal impeller to have a larger inlet flow angle and smaller eye area, the efficiency is higher. Combining the inducer with a standard centrifugal impeller provides the best combination of low NPSH without sacrificing efficiency.

A drawback to using a pump with an inducer is a reduced working flow range of the pump. The inducer reduces the NPSH of the pump, but it is only effective within a certain range. Beyond this range, the NPSH could increase dramatically. Fig. 5 shows how the NPSH curves compare between a pump utilizing an inducer versus one that does not. It is clear that as long as the pump is running between Q₁ and Q₂, the inducer is effective. Beyond this range of flow, using the pump with an inducer would not be a sensible choice.

Fig. 4 Pump with inducer that develops pressure into the eye of the impeller.

Let us review the equation for suction specific speed (Equation 2). One of the factors in the equation for suction specific speed is the rpm of the motor. Using a lower rpm motor will reduce the NPSH required by a given impeller. It must be recognized that the discharge head that is developed will also be reduced, and this may not be practical on all applications. If a lower speed is used and you want to maintain the flow and head developed at the higher speed, the impeller tip velocity must be increased. This is achieved by increasing the impeller diameter. An impeller is the approximate shape of a disc. The formula for the moment of inertia (I) of a disc about its axis is I = Wr²/2 g or half the weight (W) times the radius (r) squared, divided by the gravitational constant (g). To develop the same amount of pressure, a 1750 rpm impeller must have twice the diameter of a 3500 rpm impeller.

Fig. 5 NPSH with an inducer and without an inducer.

There are both definite advantages and disadvantages to operating at the lower speed. Comparing a 3500 rpm pump to a 1750 rpm pump, the 3500 rpm pump is substantially smaller and therefore less expensive. But at 3500 rpm, the pump’s noise level and NPSHR are increased. The designer must evaluate the obvious tradeoffs–lower cost but higher noise versus lower NPSH and noise but increased cost.

Summary

In summary, NPSH is a critical factor in selecting condensate pumps due to the high temperatures experienced in steam condensate systems. The NPSHA consists of many factors. These include condensate temperature, pressure in the condensate receiver, elevation of the condensate receiver, and pipe friction losses. Changing any of these has a direct effect on the NPSHA.

The NPSHR is dependent on the manufacturer of the pump and its design of it. The manufacturer needs to take into consideration such factors as motor speed, impeller inlet flow angle, number of vanes, and the use of an inducer.

When selecting a condensate pump, the consulting engineer should be aware of these various factors, specifically the use of an inducer and the speed of the motor. Taking these factors into account will ensure that the correct choice is made to best fit the specific application. HPAC

Reference

1. Carlson, G. F., “NPSH: It Shouldn’t Mean Not Pumping So Hot,” Heating/Piping/Air Conditioning, April 1979.

Bibliography,

1. American Society of Heating, Refrigeration and Air-Conditioning Engineers, ASHRAE 1996 HVAC Systems and Equipment Handbook, I-P Edition, 1996, pp. 38.8-38.9.
2. ITT Fluid Handling Training and Education, “Domestic High Temperature Condensate Return Systems, Bulletin No. TES-582,” 1982, pp. 13, 25, 29.
3. Karassik, I. J., Krutzch, W. C., Fraser, W. H., Messina, J. P., Pump Handbook, Second Edition, McGraw-Hill Book Co., 1986, p. 2.15.This article originally appeared in HPAC Magazine.

   Reprinted from TechTalk September 1998