The following article is designed to help you realize significant savings through pump energy and flow balance analysis.
It is a universally accepted fact that pumps are oversized for almost all industrial and commercial installations. A major reason for HVAC pump oversizing and the resulting power waste is a fear of flow unbalance (short circuiting) and a consequent lack of flow for some terminal units.
The following tips illustrate simple circuit flow analysis techniques that will enable readers to establish methods for flow balance in existing buildings and will permit a considerable reduction in pump power requirements. This analysis technique will also help identify and provide protection against short circuiting for new design and help reduce the tendency to oversize pumps.
But prior to describing the circuit flow analysis techniques, it is important to illustrate the relationship between pump power draw and operating cost. Given an electric motor drive operating at about 85 percent efficiency, each pump shaft bhp will draw about 0.88 kW. Thus, 10 pump shaft bhp will draw 8.8 kW. At a utility cost of $0.10 per kWh, the yearly cost for 10 bhp draw will be: (10 bhp) (0.88 kW per bhp) (24 hours per day) (365 days per year) ($0.10 per kWh) = $7,709 per year.
If we can reduce power consumption by 10 bhp, we will save $7,709 per year at $0.10 per kWh, $15,418 per year for a 20 bhp reduction, etc. The savings are substantial and, fortunately, can often be easily realized for oversized pumps if we establish and understand the basic physical correlations among pump power draw, the system flow-head relationship, circuit flow balance, and terminal heat transfer-flow relationships.
The pump shaft brake horsepower requirement is determined by dividing the amount of power a pump puts into the water by the pump efficiency. The correlation is as follows:
bhp = whp/Ep
Most pump horsepower curves assume water density of 8.33 pounds per gallon at 85 degrees F. For this condition, the brake horsepower relationship becomes:
bhp = H x Q/3960 x Ep
Where:
bhp = brake horsepower (hp)
whp = water horsepower (hp)
H = pump head (ft)
Q = pump flow (gpm)
Ep = pump efficiency at H, Q
Since many specifiers are interested in efficiency, they may become mesmerized by the term "pump efficiency" and neglect the terms in the pump power draw formula where true energy savings can be found (H x Q).
Pump power draw will increase directly with increases in either head or flow. Oversized pumps generally increase flow and head over that needed by the system, thus increasing system operating power draw over that actually required. The term used to describe oversizing effects is "system pump efficiency."
As an example, a pump may be selected to operate at 75 percent efficiency with the flow specified at actual system need, but with a head specification that is twice that actually required. Two separate operating conditions can be evaluated, each with a different system pump efficiency, even though in each case the pump operates at 75 percent efficiency.
If the terminals are exactly flow balanced to need, the pump power draw will be twice that required because the pump head input is twice that required. In this instance, one-half of the applied pump head is wasted because the excess head is used up by sharply throttled balance valves. Even though the pump itself operates at 75 percent efficiency, the actual system pump efficiency would decrease to 75/2 = 37.5 percent.
If the terminals are not balanced, the effect of increased head will be increased flow. In addition, a lack of balance will cause short circuiting, further increasing flow. Assuming flow has increased to twice that specified and that head is twice what is needed, the increased pump power draw over that required will be (H x 2)(Q x 2) for a power increase of four times over that needed. System pump efficiency will then become 75/4 = 18.75 percent even though the pump itself operates at 75 percent efficiency.
Wasted Power And Costs
Low system pump efficiency results in power and operating cost wastes. If in this unbalanced, overheaded system a 40 hp pump were used, the operating cost would be $30,836 per year at $0.10 per kWh. With a balanced system, a pump using only 10 hp (40/4) would be needed. The revised operating cost would be $7,709 per year, a $23,127 per year savings because 30 hp has been eliminated.
It is clear that flow and head are much more important variables in determining pump horsepower draw than pump efficiency alone. The system designer and/or balance engineer has a more important influence relative to system pump power draw than the pump designer does. The "ballpark" effect of overheaded pumps on system pumping efficiency is shown in Figure 1.
As a very rough approximation, for pumps in the 70 to 80 percent efficiency range and over a limited overhead range, each percentage point of overhead is equivalent to about one efficiency point decrease. Therefore, a 20 percent overhead is equivalent to a decrease from 80 percent efficiency to about 60 percent system pump efficiency.
If we are to decrease system pumping costs, we must first match pump input flow and head to minimum required system flow head needs. This objective can only be attained through balance and with an understanding of the relationship between flow and terminal unit heat transfer.
Water Flow
Water is used to convey heat to, or away from, the terminal unit, but water flow is not generally a major variable in terms of affecting full load terminal heat transfer. Heat transfer is governed more by unit surface area, supply water temperature, mean air temperature, wet bulb temperature, and maintenance of the external coefficient of heat transfer. The combination of variables is such that substantial variation in water flow is needed before terminal heat transfer is measurably affected.
Figure 2 illustrates a suggested flow tolerance range expected to provide from 97 to 101.5 percent design heat transfer. This is well within our heat loss or gain calculation and terminal heat transfer selection accuracy range. As an example in the use of the chart, a hot water terminal provided with 180 degrees F supply water temperature and designed for 20 degrees F delta-T has an allowable flow tolerance of ±22 percent. Similarly, a "standard" chilled water design base of 45 degrees F supply and a 10 degrees F rise has a terminal flow tolerance range of about ±12 percent.
In general, a ±10 percent flow tolerance should be established even though higher flow tolerances are allowable in many cases. Very low piping flow rates may cause piping air purging problems, while excessive flow rates simply increase pump power needs. It should be noted that some heat recovery systems and some special chilled water systems operating with water temperatures close to ambient require close (±5 percent) flow tolerance ranges to maintain required heat transfer.
The fact that flow varies approximately with the square root of head change provides a safety factor in terms of head estimation in piping systems. A flow tolerance of ±10 percent provides a head tolerance of about ±20 percent. We can certainly be more accurate than ±20 percent in head estimation because we now have accurate head estimation charts and calculators. The combination of variables is such that we can occasionally (perhaps on purpose) miss an elbow or two without affecting heat transfer results in HVAC system design.
Reprinted with permission from the ITT Industries newsletter TechTalk®, volume 19, issue 2, September 2004. This article was written by Bell & Gossett engineers. For more information, visit the Bell & Gossett Web site at www.bellgossett.com.
Publication date: 07/11/2005
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