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Good Practice in Suction Piping Design - Avoiding Hydraulic Noise

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Good Practices in Suction Piping Design – Avoiding Hydraulic Noise

Robert J. Meyer, P.E.

BSME, University of Cincinnati

MBA, Xavier University

Professional Engineer, State of Ohio

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Course Overview This course is useful for engineers involved in designing systems using centrifugal pumps. The principles explained here are applicable to many industries including chemical processing, water, and wastewater treatment. The student is expected to already have a basic understanding of NPSH and how to calculate losses in suction piping. Definitions for these key terms and concepts will be reviewed at the beginning.

This course will explain two design objectives for avoiding cavitation damage, hydraulic noise, and the maintenance expenses associated with these problems-

1. Deliver fluid to the pump suction at a pressure that avoids cavitation damage.

2. Deliver fluid to the pump suction that has a uniform flow distribution.

The student will acquire specific knowledge from this course that can be used to design better suction piping by applying the guidelines presented. After reading this material and completing the quiz, the student should:

• have a better and more practical understanding of NPSH available and NPSH required. • have the tools to make better judgments on a safe margin between NPSH available and

NPSH required. • be able to make wise suction piping design choices that produce uniform flow at the

pump inlet.

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Table of ContentsI. Review of Term and Definitions ................................................................................4 II. NPSH Available vs. NPSH Required.........................................................................5 A. Consider two factors when deciding on the margin between NPSHA& NPSHR.5 B. How NPSHR is determined by the pump manufacturer........................................5 Fig. 1 - Typical NPSH Testing at Several Flow Rates ........................................7 C. System head calculations can affect actual NPSH margin ....................................8 Fig. 2 - Pump Performance & System Head........................................................9 D. NPSH Margin Guidelines ....................................................................................10 III. Good Practices in Suction Piping .............................................................................12 A. Suction pipe velocity ...........................................................................................12 B. Pipe slope, reducers, and air pockets ...................................................................13 Fig. 3 - Suction Lift, Air Pockets and Reducers ................................................13 C. Elbows and tees...................................................................................................14 Fig. 4 - Flow Streamlines at an Elbow...............................................................14 Fig. 5 - Recommended Use of Elbows ..............................................................15 Fig. 6 - Double Suction Casing Inlet .................................................................17

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I. Review of Term and Definitions

Net Positive Suction Head (NPSH) – the total suction head in feet of liquid absolute determined at the suction nozzle and referred to datum, less the vapor pressure of the liquid in feet absolute. Note that NPSH is an ABSOLUTE pressure, not a relative or gauge pressure. In the absolute pressure scale, “0” is a perfect vacuum, and approximately 33 feet of water corresponds to a “0” gauge pressure.

Net Positive Suction Head Required (NPSHR) – the amount of total suction head in feet of liquid absolute, less the vapor pressure, required to prevent more than 3% loss in total head when operating at a certain flow rate. NPSH Required values are determined at various flow rates by the pump manufacturer. Plots of typical NPSHR tests and a general description of the most common test methods will be given later.

Net Positive Suction Head Available (NPSHA) – the total suction head in feet of liquid absolute, determined at the impeller datum, less the absolute vapor pressure of the liquid. The pump system designer must calculate the NPSH Available, which changes with flow rate and liquid level in the sump or suction tank. The general formula used by system designers is:

NPSHA = (Pt – Pv) / sg + Z – Hf where Pt = absolute pressure on free surface of liquid (ft.) Pv = vapor pressure of the liquid at pumping temperature (ft.) sg = specific gravity of the liquid (water = 1.0) Z = vertical distance between free surface and pump datum (ft., + or -) Hf = friction loss in suction line and entrance losses

USING CONSISTENT UNITS IS IMPORTANT, as always. Note that there is no Velocity Head term [V^2/(2g)] in the equation above. This is because velocity head energy is lost accelerating the fluid from the sump or tank into the suction pipe. That energy is then recovered in the suction pipe. When using this equation at the design stage, the velocity head terms cancel out. When taking actual field test data with gauges, velocity head must always be added in. Gauges always measure static pressure.

Cavitation – the formation and subsequent collapse of vapor-filled cavities in a liquid. The cavities may be bubbles or vapor-filled pockets, or a combination of both. The local pressure must be at or below the vapor pressure of the liquid for cavitation to begin. And the cavities must encounter a region of pressure higher than the vapor pressure to collapse. Bubbles which collapse on a solid boundary (such as an impeller vane or shroud wall) will cause pitting, damage, and some vibration. Cavitation pitting is evident slightly down-stream from the inlet edge of the impeller vane because it's the bubble collapse that does the damage, not the bubble formation.

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System Head - the sum of the static head between suction and discharge liquid levels, the pipe friction head, and the head lost through fittings and valves. In many systems, the static head varies because suction and discharge liquid levels vary. Friction head generally increases at a rate approximately equal to the square of the flow through the system. Friction head is affected by changes in pipe condition and valve opening.

II. NPSH Available vs. NPSH Required

A. Consider two factors when deciding on the margin between NPSHA and NPSHR

NPSH Available must ALWAYS exceed NPSH Required (NPSHA > NPSHR). Margin is the amount that NPSHA exceeds NPSHR. Margin can be expressed in two ways:

• as a difference (NPSHA - NPSHR)

• or as a ratio (NPSHA / NPSHR)

An adequate MARGIN is both necessary and important because:

1) Cavitation has already begun and is well established at the published NPSH Required (3% head drop) value. Incipient cavitation usually starts at suction pressures TWO or MORE TIMES HIGHER than the published NPSHR (3% head drop) value. The deterioration of pump performance as suction conditions approach the NPSHR value will be graphically presented in the next section.

2) Frequently, the ACTUAL pump operating flow rate exceeds the DESIGN flow rate because system heads are often over-estimated. Also, the actual head developed by the pump at the rated flow will exceed the rated head because of test code requirements. Both of these factors reduce NPSH margin, as we will see later when the interaction of these factors is presented.

B. How NPSHR is determined by the pump manufacturer

There are two types of NPSH test setups generally used by pump manufacturers. Probably the one most often used is the SUCTION SUPPRESSION test. Here a constant level open sump is used, and NPSH Available is slowly reduced by partially closing a suction valve. To obtain the most accurate results, the flow must enter the impeller eye uniformly, therefore there must be at least 5 – 10 diameters of straight pipe between the pump suction flange and the suppression valve. The second NPSHR test setup is the CLOSED LOOP test with vacuum control. This setup often gives more accurate results at low NPSH values. The suction tank is a closed vessel, and a

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vacuum pump is used to reduce the pressure in this closed vessel, and thereby reduce the NPSHA.

With either the SUCTION SUPPRESSION or CLOSED LOOP setup for NPSH testing, air entering the pump suction is always a possible problem, and the enemy to an accurate test. Suction pressures, and possibly pressure at the pump shaft seal, will be below atmospheric pressure during most of the NPSH test. Therefore suction piping joints and suction valve stems must be air tight. Pump shaft packing should be flushed with external water and adjusted relatively tight during testing, or use a double mechanical seal with flush.

The NPSH test results, as shown in Figure 1, were produced by holding the flow rate constant at 1000, 1500, 2000, and 2250 GPM respectively, while reducing the NPSH Available on the suction side of the pump. Differential head was measured to determine the 3% head drop point. Each constant flow rate is a separate test, and represents one data point on the published FLOW vs. NPSHR curve.

Once again, to emphasize the point that cavitation is well established when a pump is operated with a suction pressure equal to its NPSH Required, let's look at the 2000 GPM data in Figure 1. At ample suction pressures (NPSHA above 42 feet), the head developed is 105 feet. As suction pressure is lowered, the total developed head is reduced until it reaches just under 102 feet (3% drop off) at an NPSH value of 32 feet. At 2000 GPM, the NPSHR = 32 feet. But if full published head performance is expected at 2000 GPM flow, an NPSHA value above the published NPSHR value must be maintained. Remember that 32 feet NPSH (absolute pressure) would roughly correspond to a suction gauge reading just below zero (gauge pressure) for cool water. Please study the NPSH data presented in Figure 1 at all four flow rates.

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Fig. 1 Typical NPSH Testing at Several Flow Rates

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C. System head calculations can affect actual NPSH margin

The previous NPSH Required data came from a wastewater pump with 6” suction and discharge, 13” impeller, operating at 1780 RPM. Continuing on with that same example, we will show how the interaction between the pump head – capacity curve, the calculated system head curve, and the actual system head curve can have an adverse effect on NPSH margin. For this example, the rated condition point specified by the system designer is 2000 GPM (gallons per minute) at 100 feet of head.

Figure 2 shows the actual head vs. flow (H-Q) performance curve for the pump mentioned above at 13” impeller diameter at 1780 RPM. This pump would meet the 2000 GPM at 100 feet rating. The first thing to note is the actual head produced at 2000 GPM is approximately 105 feet, not 100 feet. Typical test standards used in the USA allow only a positive tolerance on head. The pump supplied will always meet, or more likely exceed, the rated head. Common International (ISO) test standards provide for a bi-lateral (+ / -) tolerance on head.

The actual pump operating point will be where the actual H-Q curve intersects with the actual system head curve. The dashed system head curve in Figure 2 is the one calculated during system design. The solid system head curve is the actual result once the system is in operation. Typically the system designer will be conservative in estimating friction losses, resulting in the calculated system head curve being above the actual system head curve. This factor, plus the positive test tolerance on head that the pump manufacturer must meet, has led to an actual operating point of 2160 GPM, instead of 2000 GPM in this example.

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Fig. 2 – Pump Performance & System Head

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What has this done to the NPSH margin that the system designer expected? Figure 2 shows NPSH Required vs. Flow. This information is supplied by the pump manufacturer, and is based on the data given in Figure 1. Figure 2 also shows the NPSH Available as calculated by the system designer. At the rated flow of 2000 GPM, the calculated or expected NPSH margin was: NPSHA – NPSHR = 40 feet – 32 feet = 8 feet OR NPSHA / NPSHR = 40' / 32' = 1.25

However, at the actual flow rate of 2160 GPM, the actual NPSH margin is:

NPSHA – NPSHR = 37.7 feet – 34.5 feet = 3.2 feet OR

NPSHA / NPSHR = 37.7' / 34.5' = 1.09

The reality of the ACTUAL operating flow rate being greater than the RATED condition has led to a loss of NPSH safety margin in the design. Study Figure 2 carefully to understand how these factors have interacted.

D. NPSH Margin Guidelines

The following table presents recommended NPSH margins (ratios) for various applications. Safety margins are always subjective, and the actual margin used is always a balance between avoiding potential cavitation damage and initial cost. However, the cost of correcting cavitation problems after construction can be significant.

APPLICATION MARGIN (NPSHA / NPSHR)

Chemical 1.1* - 1.3

Electric Power 1.1* - 1.5

Water / Wastewater 1.3 – 1.7

General Industry 1.2 – 1.7

Pulp & Paper 1.1* - 1.4

Building Trades 1.2* - 1.5

Cooling Tower 1.4* - 1.7

* = or 5 feet difference, whichever is greater

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Remember our previous example for a design of 2000 GPM at 100 feet. The pump used in that example was a wastewater pump. The original design at 2000 GPM did not quite meet the minimum recommended NPSH margin given above (1.3). At the design flow of 2000 GPM, the margin was 1.25 (40' / 32'). After construction and commissioning, the actual system head turned out to be lower than calculated, and the actual operating point was 2160 GPM. At that actual flow rate, the NPSH margin further deteriorated to a 1.09 ratio or 3.2' (difference).

Using Fig. 2, we can see there is 37.7' NPSHA at 2160 GPM. From Fig.1, we can interpolate between the 2000 and 2250 GPM NPSH test data to see that at 37.7' NPSHA, the developed head is about 1.5% below the full head you could expect at high NPSHA levels (like 45' NPSHA or higher).

In this example, the pump actually operates about half way towards its 3% head loss NPSHR condition. Some cavitation is well established at that point. Note also from the NPSH test curves in Figure 1, just below the 3% head drop point (the published NPSHR value) the developed head starts dropping more rapidly. This is typical for many centrifugal pump designs, and this “slippery slope” is one you don't want to get near.

If increased NPSH margin is desired, either the NPSHA of the system must be raised or the NPSHR of the pump must be lowered.

Changes to the system that will increase the NPSH Available include:

a) Increase the supply tank elevation, or raise the minimum tank (wet well) level.

b) Lower the pump relative to the supply tank

c) Increase suction pipe size to reduce flow velocity and friction loss.

d) Add a booster pump.

e) Reduce the liquid temperature, thus reducing the vapor pressure.

Changes to the pump that will reduce the NPSH Required include:

a) Select a different pump with lower NPSHR. This may mean using a larger pump at lower speed. Both pump and motor will have higher initial costs, but operating costs may be lower with reduced cavitation, reduced vibration, reduced wear if abrasives are present, and longer seal life. Going back to our previous example, for the design point of 2000

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GPM at 100 feet, a 6” suction and 6” discharge pump with 13” impeller at 1780 RPM was selected. NPSHR at 2000 GPM was 32 feet. An alternate selection could have been a 6” suction and 6” discharge pump with 16.5” impeller at 1180 RPM, which has an NPSHR of 16' at 2000 GPM (40'/16' = 2.5 NPSH Margin). This is huge increase in margin from the original 1.25 margin, along with a higher initial cost for the pump and motor. But these additional upfront costs will likely result in longer pump wear life, and lower noise and vibration due to cavitation.

b) Add an inducer to the impeller inlet. This is sometimes done in low NPSH applications. Inducers act as first stage impellers with low inlet angles, and reduce the NPSHR in the flow range they are designed for. Inducers are sometimes used in chemical pump applications on clear liquids. Because of their low inlet angles and high vane overlap, their ability to pass spheres is very limited.

III. Good Practices in Suction Piping

The previous section dealt with delivering sufficient suction pressure (NPSH) to avoid cavitation damage and its detrimental effects on vibration, bearing life, and seal life. That turns out to be only half the battle.

To avoid hydraulic noise and the associated vibration, you must also deliver a uniform flow velocity distribution to the pump inlet. A centrifugal pump that lacks a straight and uniform flow pattern at its inlet will not respond properly, or perform to its maximum capability. A non-uniform or swirling flow profile can lead to noisy operation, random axial load oscillations, and premature bearing failures.

A. Suction pipe velocity

Suction pipe size should generally be at least one size larger than the pump inlet. Suction pipe inlet velocities at the sump should be limited to 5 ft./sec. If there is a suction manifold arrangement, the main line should also be limited to 5 ft./sec. Branch suction lines off the main manifold should have flow velocities in the range of 5 – 8 ft./sec. If the designer must take some liberties with other suction pipe guidelines concerning elbows, straight lengths of pipe, and the like, then it is more important to stay near the low end of these flow velocity guidelines.

When using solids handling pumps, horizontal line velocities below 3 ft./sec. can cause settling of solids and roping of stringy materials. This settled material might later be pulled up into the pump during higher flow conditions and overwhelm the pump's solids handling ability – thus causing a plug.

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B. Pipe slope, reducers, and air pockets

Flooded suction for centrifugal pumps is always preferred. This eliminates any priming issues, and also provides higher NPSHA. However, for pumps that must operate with a suction lift condition, the suction line must slope constantly upward toward the pump. Refer to the top half of Figure 3.

Also, any valves that are installed in the suction line should have their stems horizontal to avoid collecting air or gas at high points. Reducers will generally be required just ahead of the pump suction flange, since the suction pipe will be at least one size larger than the pump suction. They must be installed to avoid air pockets, as shown in the lower half of Figure 3. Reducers at the pump suction should be the conical type. Contoured eccentric reducers are NOT recommended, as they can disturb flow right in front of the pump.

Also, more than one pipe size reduction in a single reducer fitting should be avoided. Large reductions over short lengths can result in non-uniform flow patterns.

Fig. 3 – Suction Lift, Air Pockets, and Reducers

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C. Elbows and tees

Because flow through elbows and tees can create non-uniform velocity distributions at their exits, it is strongly recommended that 5 to 10 pipe diameters of straight pipe be provided in front of the pump suction. Uneven and swirling flow velocities at the pump inlet will result in poor angles of attack between the flow and the impeller blades, leading to hydraulic noise, axial loading oscillation, possible bearing failure, and cavitation.

If suction pipe velocities are near the maximum recommended values, then it is more important to provide a longer straight run of pipe in front of the pump suction (i.e. follow the 10 pipe diameter recommendation). See the top diagram in Figure 5. Bends at 30 – 45 degrees are always preferred over 90 degree bends.

Figure 4 shows typical flow streamlines through a short radius elbow at higher velocities. Note that the flow shifts toward the outside wall during the second half of the turn. Flow becomes separated from the inside wall, and there can be pockets of eddies in this area.

If a short radius elbow with higher flow velocities were near the pump inlet, the result would be an undesirable flow pattern for the pump inlet. With such an uneven pattern, it is impossible to have the flow line up smoothly with impeller blade inlet angles in all four quadrants of the impeller eye. Large radius elbows with lower flow velocities would show much less tendency toward non-uniform flows at the elbow exit.

Fig. 4 – Flow Streamlines at an Elbow

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When elbows must be incorporated in the suction piping design, there are certain recommendations that must be followed. For double suction pumps, do NOT place an elbow near the suction with its plane parallel to the pump shaft. Because of the non-uniform flow at the elbow exit, as seen in Figure 4, an elbow oriented this way will overload one side of the impeller while starving the other side. This upsets the axial balance of the rotor and may result in cavitation on the starved side. High axial fluctuating loads and noisy operation are likely results. See Figure 5. If an elbow must be used near a double suction pump inlet, keep the flow velocities low and only use long radius or reducing long radius elbows. Plus the plane of the elbow MUST be perpendicular to the pump shaft.

Fig. 5 – Recommended Use of Elbows

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If the suction piping arrangement must contain two or more 90 degree turns, those turns should be in the same plane, as shown in the middle, left side of Figure 5. Orientation of the turns in the same plane allows the second turn to rectify the non-uniform flow coming from the first turn. Orientation of the two turns in perpendicular planes (shown as NOT RECOMMENDED) can induce a rotational swirl pattern to the flow entering the pump inlet. Once again, this will lead to a poor match between liquid flow angles and inlet blade angles, with hydraulic noise and cavitation as the result.

Now that the evils caused by elbows near the pump suction have been explained, you might be thinking – hold on just a minute. There are several cases where pump manufacturers incorporate an elbow right in front of the impeller inlet. How do they get away with that?

One such group of pumps are dry-pit vertical wastewater pumps used in the sewage treatment industry. Vertical pumps are often preferred here for their smaller floor space requirements. Vertically oriented pumps can be direct-driven by a motor mounted on top of the pump, or driven through line shafting connected to a motor at a higher elevation. In either case, the pump manufacturer must supply a suction elbow mounted directly in front of the impeller inlet. Here the suction elbow is a necessary evil. However, we can diminish the evil by making smart choices.

One principle to remember is that flow through CONVERGENT channels produces uniform flow, while flow through DIVERGENT channels produces non-uniform flow patterns. Therefore, the best designs for vertical pumps incorporate REDUCING ELBOWS on the suction, or LONG RADIUS ELBOWS. These two choices result in the most uniform flow for the impeller inlet. Short radius, non-reducing elbows would be the least desirable choice for an elbow just in front of the impeller inlet.

Another pump design that incorporates “elbows” just ahead of the impeller inlet is a double suction or split-case pump, such as previously shown in Figures 3 and 5. Again, the turn in the suction passage (built into the casing), just ahead of each impeller inlet, is a necessary evil in this design. Casing suction passage areas must be generous. See Figure 6. The area at “AF” should be at least 1.5 times the impeller eye area “D”.

Once again, convergent flow in the suction passage of the pump casing (due to reducing areas leading to the impeller inlet on each side) should result in reasonably uniform flow at the impeller eye. Good double suction pump inlet designs also feature anti-rotation baffles to suppress swirl at the impeller inlet. Note the anti-rotation baffle shown in Figure 6. Be wary of double suction pump designs where the casing suction passage looks unusually compact in the axial direction (shaft axis). These designs could produce greater hydraulic noise.

The pump manufacturer has the primary responsibility for supplying proper suction elbows for vertical pumps, and designing suction passages in split-case pump casings. But better informed pump users and system designers can assure that the best choices are made for suction elbows on

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vertical pumps. And that double suction pumps with unusually small casing suction passages are avoided.

Fig. 6 – Double Suction Casing Inlet