Terry Henshaw is a retired consulting engineer who designs pumps and related high pressure equipment and conducts pump seminars. For 30 years, he was employed by Ingersoll Rand and Union Pump. Henshaw served in various positions in the Hydraulic Institute, ANSI Subcommittee B73.2, API 674 manufacturers' subcommittee and ASME Performance Test Code Committee PTC 7.2. He authored a book on reciprocating pumps, several magazine articles and the two pump sections in Marks' Handbook (11th Edition). He has been awarded six patents. Henshaw is a registered professional engineer in Texas and Michigan, is a life fellow of ASME and hold engineering degrees from Rice University and the University of Houston.
Dampening the Pulsations
Any characteristic of the suction system that tends to absorb the pulses from the pump will reduce acceleration head. The suction stabilizer, therefore, helps those systems with excessive acceleration head and/or with gas entrained in the liquid. (The most effective suction stabilizer is a flow‑through type that also separates free gas from the liquid. Any excess gas must be vented from the stabilizer, possibly piped back to the vapor space in the suction vessel.)
According to Hydraulic Institute standards (2), a properly selected, installed and maintained dampener will reduce the effective length of the suction line in the above equation to about 10 pipe diameters, i.e., for a 6 in suction pipe, L would be about 60 in (5 ft). This would result in a calculated acceleration head in Problem 1 of only 0.7 psi.
Another effective method of reducing acceleration head on an atmospheric‑pressure suction system is to install a section of soft hose as part of the suction line, adjacent to the pump.
Shortcomings of This Equation
Equation 19‑2 is not sophisticated enough to compensate for such factors as system elasticity and the velocity of a pressure wave in the pumpage (sonic velocity due to liquid elasticity). It is therefore recommended for use only for relatively short, non‑elastic suction lines.
Miller (3) reported that his tests indicated acceleration head to be much less than calculated with the above equation. Some field installations also operate satisfactorily with NPSHa considerably less than this equation indicates as necessary. On the other hand, some installations require NPSH that agrees favorably with this equation.
The reason for these discrepancies is not known, but, in addition to the above, it may be due to gas, such as air, being liberated (or trapped) in the suction line. Any gas entrained in the liquid, or collected at a high point in the suction piping, tends to absorb the pulsations from the pump, and thereby reduces acceleration head.
Some pump operators have reported that suction stabilizers, which were designed to also separate and accumulate gas, have, to their surprise, required periodic venting. If the stabilizer had not been in the suction line (or did not have this separation feature), the pump would have ingested gas, possibly resulting in shock operation, or in the extreme case, causing one or more pumping chambers to become gas bound or vapor locked. Without the stabilizer, the agitation in the suction line would have been greater, and more gas could have been liberated. The pressure shocks caused by gas ingestion can cause failure of pump and system components (1).
The Water Hammer Equation
For a quick closing (or opening) valve, reference 4 provides the equation for water hammer as follows:
h = Head increase or decrease
c = Sonic velocity in liquid
V = Change in velocity
g = Acceleration of gravity
This equation provides the maximum head that a quick operating valve can generate. Note that the length of the pipe is absent from the equation, and enters into the evaluation only to the extent that it determines how fast the valve must close (or open) to be considered as quick operating. This equation could therefore be used to calculate a more accurate pump acceleration head if we could accurately determine the change in velocity of the pumpage in the pipe.
Unfortunately, the velocity change is more complex than shown in Figure 1, because it is dependent on the fluid compressibility, the clearance volume in the pumping chamber and the effectiveness of the valve springs in closing the pump valves quickly enough for smooth pump operation. (A weak or broken spring, on either a suction or discharge valve, will cause a significant velocity change.) All these factors are difficult to establish for field installations.
At one time, one pump vendor produced power pumps that, because of unique construction of the fluid end, could not be equipped with springs on the suction valves. A vendor of pulsation dampeners once remarked (without knowing the reason) that that pump brand required twice the discharge dampener of other vendors' pumps. Sound level tests, on a number of brands of power pumps, also revealed that the springless pumps were noisier than equivalent pumps with valve springs. Adequate springs are required on both suction and discharge valves to provide a quiet, smooth running power pump.
1. Henshaw, Terry L., Reciprocating Pumps, Van Nostrand Reinhold Co., Inc., 1987.
2. Hydraulic Institute Standards, Hydraulic Institute, 9 Sylvan Way, Suite 360, Parsippany, NJ 07054‑3802.
3. Miller, J. E., "Experimental Investigation of Plunger Pump Suction Conditions", ASME Paper 64‑PET‑14, 1964.
4. Daugherty and Ingersoll, Fluid Mechanics, 5th Edition, McGraw‑Hill Book Co., NY, 1954.
5. Marks' Mechanical Engineers' Handbook, 6th Edition, pg. 14‑6, McGraw‑Hill Book Co, New York, 1958.