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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. Theprinciples explained here are applicable to many industries including chemical processing, water,and wastewater treatment. The student is expected to already have a basic understanding ofNPSH and how to calculate losses in suction piping. Definitions for these key terms and conceptswill 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 bettersuction piping by applying the guidelines presented. After reading this material and completingthe 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 Contents

I. 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 determinedat 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 absolutepressure scale, 0 is a perfect vacuum, and approximately 33 feet of water corresponds to a 0gauge pressure.

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

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

NPSHA =(Pt Pv) / sg +Z Hfwhere 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 VelocityHead term [V 2/(2g)] in the equation above. This is because velocity head energy is lostaccelerating the fluid from the sump or tank into the suction pipe. That energy is then recoveredin the suction pipe. When using this equation at the design stage, the velocity head terms cancelout. 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. Thecavities may be bubbles or vapor-filled pockets, or a combination of both. The local pressuremust be at or below the vapor pressure of the liquid for cavitation to begin. And the cavities mustencounter a region of pressure higher than the vapor pressure to collapse. Bubbles whichcollapse 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 inletedge of the impeller vane because it's the bubble collapse that does the damage, not the bubbleformation.


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

I I. NPSH Available vs. NPSH Required

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

NPSH AvailablemustALWAYS exceedNPSH Required (NPSHA >NPSHR). Margin is theamount 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 TIMESHIGHER than the published NPSHR (3% head drop) value. The deterioration of pump

performance as suction conditions approach the NPSHR value will be graphically presented inthe next section.

2) Frequently, the ACTUAL pump operating flow rate exceeds the DESIGN flow rate becausesystem heads are often over-estimated. Also, the actual head developed by the pump at the ratedflow will exceed the rated head because of test code requirements. Both of these factors reduceNPSH 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 theone most often used is the SUCTION SUPPRESSION test. Here a constant level open sump isused, and NPSH Available is slowly reduced by partially closing a suction valve. To obtain themost accurate results, the flow must enter the impeller eye uniformly, therefore there must be atleast 5 10 diameters of straight pipe between the pump suction flange and the suppressionvalve. The second NPSHR test setup is the CLOSED LOOP test with vacuum control. This setupoften 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 theNPSHA.

With either the SUCTION SUPPRESSION or CLOSED LOOP setup for NPSH testing, airentering 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 atmosphericpressure during most of the NPSH test. Therefore suction piping joints and suction valve stemsmust be air tight. Pump shaft packing should be flushed with external water and adjustedrelatively 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 at1000, 1500, 2000, and 2250 GPM respectively, while reducing the NPSH Available on thesuction 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 FLOWvs. NPSHR curve.

Once again, to emphasize the point that cavitation is well established when a pump is operatedwith 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 suctionpressure 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 publishedhead performance is expected at 2000 GPM flow, an NPSHA value above the published NPSHRvalue must be maintained. Remember that 32 feet NPSH (absolute pressure) would roughlycorrespond to a suction gauge reading just below zero (gauge pressure) for cool water. Pleasestudy 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 howthe 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, therated condition point specified by the system designer is 2000 GPM (gallons per minute) at 100feet of head.

Figure 2 shows the actual head vs. flow (H-Q) performance curve for the pump mentioned aboveat 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, not100 feet. Typical test standards used in the USA allow only a positive tolerance on head. Thepump 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 actualsystem head curve. The dashed system head curve in Figure 2 is the one calculated duringsystem 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 thecalculated system head curve being above the actual system head curve. This factor, plus thepositive test tolerance on head that the pump manufacturer must meet, has led to an actualoperating 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 showsNPSH Required vs. Flow. This information is supplied by the pump manufacturer, and is basedon the data given in Figure 1. Figure 2 also shows the NPSH Available as calculated by thesystem 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 ledto 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 betweenavoiding potential cavitation damage and initial cost. However, the cost of correcting cavitationproblems 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 thatexample was a wastewater pump. The original design at 2000 GPM did not quite meet theminimum recommended NPSH margin given above (1.3). At the design flow of 2000 GPM, themargin was 1.25 (40' / 32'). After construction and commissioning, the actual system headturned 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 interpolatebetween the 2000 and 2250 GPM NPSH test data to see that at 37.7' NPSHA, the developed headis about 1.5% below the full head you could expect at high NPSHA levels (like 45' NPSHA orhigher).

In this example, the pump actually operates about half way towards its 3% head loss NPSHRcondition. Some cavitation is well established at that point. Note also from the NPSH test curvesin Figure 1, just below the 3% head drop point (the published NPSHR value) the developed headstarts 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 theNPSHR 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 tankc) 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 lowerspeed. Both pump and motor will have higher initial costs, but operating costs may belower 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 RPMwas selected. NPSHR at 2000 GPM was 32 feet. An alternate selection could have been a6 suction and 6 discharge pump with 16.5 impeller at 1180 RPM, which has anNPSHR of 16' at 2000 GPM (40'/16' = 2.5 NPSH Margin). This is huge increase inmargin 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, andlower 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 theflow range they are designed for. Inducers are sometimes used in chemical pumpapplications on clear liquids. Because of their low inlet angles and high vane overlap,their ability to pass spheres is very limited.

I I I . Good Practices in Suction Piping

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

To avoid hydraulic noise and the associated vibration, you must also deliver a uniform flowvelocity distribution to the pump inlet. A centrifugal pump that lacks a straight and uniform flowpattern 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, andpremature bearing failures.

A. Suction pipe velocity

Suction pipe size should generally be at least one size larger than the pump inlet. Suction pipeinlet velocities at the sump should be limited to 5 ft./sec. If there is a suction manifoldarrangement, the main line should also be limited to 5 ft./sec. Branch suction lines off the mainmanifold should have flow velocities in the range of 5 8 ft./sec. If the designer must take someliberties with other suction pipe guidelines concerning elbows, straight lengths of pipe, and thelike, 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 settlingof solids and roping of stringy materials. This settled material might later be pulled up into thepump during higher flow conditions and overwhelm the pump's solids handling ability thuscausing 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 liftcondition, 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 avoidcollecting air or gas at high points. Reducers will generally be required just ahead of the pumpsuction flange, since the suction pipe will be at least one size larger than the pump suction. Theymust be installed to avoid air pockets, as shown in the lower half of Figure 3. Reducers at thepump 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 theirexits, it is strongly recommended that 5 to 10 pipe diameters of straight pipe be provided in frontof 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, axialloading oscillation, possible bearing failure, and cavitation.

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

Figure 4 shows typical flow streamlines through a short radius elbow at higher velocities. Notethat 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 bean undesirable flow pattern for the pump inlet. With such an uneven pattern, it is impossible tohave the flow line up smoothly with impeller blade inlet angles in all four quadrants of theimpeller eye. Large radius elbows with lower flow velocities would show much less tendencytoward 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 certainrecommendations that must be followed. For double suction pumps, do NOT place an elbow nearthe suction with its plane parallel to the pump shaft. Because of the non-uniform flow at theelbow exit, as seen in Figure 4, an elbow oriented this way will overload one side of the impellerwhile 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 flowvelocities low and only use long radius or reducing long radius elbows. Plus the plane of theelbow 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 shouldbe in the same plane, as shown in the middle, left side of Figure 5. Orientation of the turns in thesame 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) caninduce 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 andcavitation as the result.

Now that the evils caused by elbows near the pump suction have been explained, you might bethinking hold on just a minute. There are several cases where pump manufacturers incorporatean 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 treatmentindustry. 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, ordriven 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. Herethe suction elbow is a necessary evil. However, we can diminish the evil by making smartchoices.

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

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

Once again, convergent flow in the suction passage of the pump casing (due to reducing areasleading to the impeller inlet on each side) should result in reasonably uniform flow at theimpeller eye. Good double suction pump inlet designs also feature anti-rotation baffles tosuppress 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 theaxial direction (shaft axis). These designs could produce greater hydraulic noise.

The pump manufacturer has the primary responsibility for supplying proper suction elbows forvertical pumps, and designing suction passages in split-case pump casings. But better informedpump 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 areavoided.

Fig. 6 Double Suction Casing Inlet

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