`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 1 of 72
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 1 of 72
`
`

`

`Second Edition
`
`Copyright @ 1985, 1992 by Butterworth-Heinemann.
`An Imprint ofElsevier
`P.mBllois m5y b. 6o!€nt di.cny f.om Els6,i.& Sci6rl.! and L.h.olog! Rrghrs t.p.rh.d in
`Odord, UK. Phon.: {44) 1866 843630, Fix: (a4) 1845 853333, 6{d[ p..mi$iomO6ta6d6im.uk.
`\bu n.y Jso c6dsL you. r.qusd on-lir. vi. th6 Elsai. hofi.paga; htp//ri**..ts.tbrcom by
`!.lrcrinq 'Cuaron.. SLrpport' lnd Ih.n robtaini@ Parnierions'.
`
`Originally published by Culf Professional Publishing,
`Houston, TX.
`
`For information, pleas€ contact:
`Manager of Special Sales
`Butterworth-Heinemann
`225 Wildwood Avenue
`Wobum, MA 01801-2041
`Tel: 781-904-2500
`Fax 781-904-2620
`For information on all Bufterwonh-Lleinemann publications
`available, contact our World Wide Web home page at:
`http:l'www,bh.com
`Printed on Acid-Free Paper (e)
`
`Transfe ed to Digital Printing,2010
`
`Printed and bound in the United Kingdom
`
`Ubrary of Cor4rtss Catologir4-iFhbtcrtim lhtt
`Lobanoff, Val S., 1910-
`Cenrifugal pumps: design & application/Val S.
`Lobanoff, Robert R. Ross.-2nd ed.
`p. cm.
`Includes index.
`ISBN-l3 : 978-0-87201-2N- 4 ISBN-lo: 0-87201-200-X
`1. Centrifugal pumps. I. Ross, Robert R.,
`1934- . tr. Tide.
`Tt9t9.L52 t9y2
`621.6'1-dcm
`
`9l-41458
`CIP
`
`ISBN-l3 : 978-0-87201-200-4
`ISBN-IO: G87201-200-X
`
`iv
`
`PralaGo
`
`Fant Bl Ei0iemnontr
`
`B ,r*u*on.....
`
`System Analysis for puml
`Head Capacity Curve. Pur
`ily. Conslruction. Pump S
`
`2 r**spoodnd
`
`Definition of Pump Spociti(
`ciric Speed Charts. Corec
`sion.
`
`$ mpomo*en..
`
`lmpeller Laloul. Developn
`ler End Viow. tmpelter lnt
`tions. Notation.
`
`rtl
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 2 of 72
`
`

`

`18
`
`=E:|<j=e<
`
`\
`
`by Frsd R. &erasl'
`Engineering Dynamics lncorporated
`
`V0brat[@m amd
`No0sc [m
`Fumops
`
`lntroduction
`Although a certain amount of noise is to be expected from cenhifugal
`pumps and their drivers, unusually high noise levels (in excess of 100
`dB) or particularly high frequencies (whine or squeal) can be an early'
`indicator of potential mechanical failures or vibration problems in cen-
`trifugal pumps. The purpose of this chapter is to concentrate on the
`mechanisms that may produce noise as a by-product; however, reduction
`of the noise, per se, is not the main concern. The main point of interest of
`this chapter is to study the mechanisms and their effect on the reliabilirr
`of the pump system. Methods will be presented to reduce the vibration
`(and noise) or eliminate the basic causes by modi$ing the pump or pip
`ing system.
`The occurrence of significant noise levels indicates that sufficient en-
`ergy exists to be a potential cause of vibrations and possible damage rc
`the pump or piping. Defining the source and cause of noise is the fim
`step in determining whether noise is normal or whether problems mar
`exist. Noise in pumping systems can be generated by the mechanical mc-
`tion of pump components and by the liquid motion in the pump and pip-
`ing systems. Noise from internal mechanical and liquid sources can k
`transmitted to the environment.
`Effective diagnosis and treatment of noise sources to control pur--
`noise require a knowledge of the liquid and mechanical noise-generanc--
`
`' Thg author wish6a lo acknolyledgp the contrlbutions by th€ engingerlng 6taff ol Englneoring Dyna..r i=
`lnc., who pedormed many ot the analFoi and fi€ld lesl8.
`
`422
`
`mechanisms and comrnon noisr
`transmitted. If noise itself is th
`acoustic enclosures or other tre
`
`Sourcel
`
`Mechanlcal Noi8e Sources
`
`Common mechanical sources
`pump components or surfaces b
`generated in the liquid or air. Im
`bearings, vibrating pipe walls,
`mechanical sources.
`In centrifugal machines, im
`causes mechanical noise at twi(
`speed is near or passes throueh
`generated by high vibrations rIs
`of bearings, seals, or imoellers.
`ized by a high-pitched squeal. V
`tor fans, shaft keys, andcouplir
`rngs produce high-frequency no
`and speed.
`
`Liquld Notse Sources
`
`These are pressure fluctuatior
`Liquid noise can be produced by
`(turbulence). pulsations, cavitatir
`ration, and impeller interaction
`pressure pulsations and flow mo
`or broad-band frequency compon
`any part of the structure includin
`cal vibration, then noise may b
`rypes of pulsation rou."e. ociu.
`o Discrete-frequency componentt
`as vane passing frequency and
`. Flow-induced pulsatlon ciused
`tlons and side branches in the ;
`r Broad-band turbulent energy r€
`. Intermittent bursts of broa"d'-bar
`ing, and water hammer.
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 3 of 72
`
`

`

`Vibration and Noise in Pumps 423
`
`mechanisms and cotnmon noise conduction paths by which noise can be
`transmitted. If noise itself is the major concern, it can be controlled by
`acoustic enclosurcs or other treatment [1, 2].
`
`Sources ol Pump Nolse
`
`Mechanlcal Noise Sources
`
`Common mechanical sources that may produce noise include vibrating
`pump components or surfaces because of the pressure variations that are
`generated in the liquid or air. Impeller or seal rubs, defective or damaged
`bearings, vibrating pipe walls, and unbalanced rotors are examples of
`mechanical sources.
`In centrifugal machines, improper installation of couplings often
`causes mechanical noise at twice pump speed (misalignment). If pump
`speed is near or passes through the lateral critical speed, noise can be
`generated by high vibrations resulting ftom imbalance or by the rubbing
`of bearings, seals, or impcllers. If rubbing occurs, it may be character-
`ized by a high-pitched squeal. Windage noises may be generated by mo-
`tor fans, shaft keys, and coupling bolts. Damaged rolling element bear-
`ings produce high-frequency noise [3] related to the bearing geometry
`and s@.
`
`Liquld Nols6 Sources
`
`These are pressure fluctuations produced directly by liquid motion.
`Liquid noise can be produced by vortex fomation in high-velocity flow
`(turbulence), pulsations, cavitation, flashing, water hammer, flow sepa-
`ration, and impeller interaction with the pump cutwater. The resulting
`pressure pulsations and flow modulations may produce either a discrete
`or broad-band frequency component. If the generated frequencies excite
`any part of fte sEucture including the piping or the pump into mechad-
`cal vibration, then noise may be radiated imo the environnent. Four
`types of pulsation sources occur commonly in centrifugal pumps [2]:
`. Discrete-frequency components generated by dre pump impeller such
`as vane passing
`and multiples,
`o Flow-induced pulsation caused by turbulence such as flow past restric-
`tions and side branches in the piping system.
`. Broad-band turbulent energl resulting from high flow velocities.
`. Intermittent bursts of broad-band energy caused by cavitation, flash-
`ing, and watu hanrmer.
`
`l l
`
`l
`
`lf,hnatflom amdl
`!o[se [m
`rumops
`
`)tion
`is to be expected from centrifueal
`igh noise levels (in excess of l]m
`3rhine o1 squeal) can be an early
`res or vibration problems in cen-
`chapter is to concentrate on the
`a by-product; however, reduction
`cern. The main point ofinterest of
`s and their effect on the reliabiliry
`presented to reduce the vibration
`es by modiffing the pump or pip
`
`levels indicates that sufficient en-
`'ibrations and possible damage to
`ce and cause of noise is the first
`rormal or whether problems may
`generated by the mechanical mo-
`pid motion in the pump and pip
`nnical and liquid sourcas can be
`f noise sources to control DumD
`and mechanical noise-geneiation
`
`th6 sngln.lrlng steff ol Englneering Dynlmlcs
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 4 of 72
`
`

`

`incorect application, hydrauli
`sign and manufacturing flaws
`vibrations and failures are [5]
`lnstelhtlon/ilalntenance
`
`Unbalance
`Shaft +o-shaft misalignment
`Seal rubs
`Case distortion caused bv oi
`Piping dynamic r".ponr" ('r.
`Support structural response
`Anchor bolts/grout
`Improper assembly
`
`Appllcarion
`
`Operating off of design point
`Improper speed/flow
`Inadequate NPSH
`Entrained air
`
`Hydraullc
`
`Interaction of pump (head-flc
`Hydraulic instabilities
`Acoustic resonances (Dressun
`rtr/ater hammer
`Flow distribution problems
`Recirculation
`Cavitation
`Flow induced excitation (turb
`High flow velocity
`
`O€Cgn/[anutacturlng
`
`L:teral critical speeds
`lbrsional critical speeds
`Improper bearings or seals
`Rotor insrability
`Shaft misalignment in iournah
`Impeller resonances
`Bearing housing/pedestal reso
`
`424 Centrifugal Pumps: Design and Application
`
`A variety of secondary flow patterns [4] that produce pressure fluctua-
`tiors are possible in centritrgal pumps, as shown in Figure l8-1, particu-
`-operation
`at off-design flow. The numbers shown in the flow
`larly for
`stream are the locations of the following flow mechanisms:
`l. Stall
`2. Recirculation (secondarY flow)
`3. Circulation
`4. Leakage
`5. Unsteady flow fluctuations
`6. \lhke (vortices)
`7. Thrbulence
`8. Cavitation
`
`Caulcs of Vlbrrtlon3
`
`Causes of vibrations are of major concern because of the damage to the
`pump and piping that generally results from excessive vibrations. Vibra-
`iioni in pumps may be a result of imProper installation or maintenance,
`
`?,
`
`8
`
`t
`
`IIIIEI GUI DT
`
`\-s-l
`
`I
`
`1o
`
`ROTATIOT{AL ATIS
`
`l.
`v.
`Flgurc l&1. Secondary llow around pump imPoller ofl-deBign flow EPRI Re
`search Projoct 120er8, Roport CS'1445 [41.
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 5 of 72
`
`

`

`Application
`
`4l that produce pressure fluctua-
`s shown in Figure l8-1, particu-
`lbe numbers shown in the flow
`g flow mechanisms:
`
`Vibration and Noise in Pumps 425
`
`incorrect application, hydraulic interaction with the piping system, or de-
`sigo and manufacturing flaws, Some of the common causes of excessive
`vibrations and failures are [5]:
`
`lnstallation/Malntenance
`
`Unbalance
`Shaft-to-shaft misalignment
`Seal rubs
`Case distortion caused by piping loads
`Piping dynamic response (supports and restraints)
`Support structural response (foundation)
`Anchor bolts/grout
`Improper assembly
`
`Appllcation
`
`Operating off of design p<iint
`Improper speed/flow
`Inadequate NPSH
`Entrained air
`
`Hydraullc
`
`Interaction of pump (head-flow curve) with piping resonanoes
`Hydraulic instabilities
`Acoustic resonances (pressure pulsations)
`rirhter hammer
`Flow distribution problems
`Recirculation
`Cavitation
`Flow induced excitation (turbulence)
`High flow velocity
`
`Ileslgn/ilanulscturlng
`
`Lateral critical speeds
`Torsional critical speeds
`Improper bearings or seals
`Rotor instability
`Shaft misalignment in journals
`Impeller resonances
`Bearing housing/pedestal resonancas
`
`ratlons
`
`ern because of the damage to the
`om exccssive vibrations. Vibra-
`,per installation or maintenance,
`
`7,
`
`I
`
`t
`
`4
`
`-+-
`
`impeller off-do8ign flow EPHI Rs-
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 6 of 72
`
`

`

`produce torsional or lateral vi
`coupling are similar to those j
`small angular misalignment or
`not constant. If one shaft spee
`has a faster rotational rate [6] f
`tional rate for part of the revol
`sults in a second harmonic (tu
`
`iE=
`=F:[=
`
`(.)
`
`,lTN
`n@,
`
`t N,
`
`,.)',r.rnd
`{
`(b) Gl
`
`(c) Pol.r
`Flgurc l&.2. E tecls o, ang
`
`426 Contrifugal Pumps: Design and Application
`
`Many of these causes are a result of an interaction of the pump (or its
`driver) with the fluid or the structure (including piping). This interactive
`relationship requires that the complete system be evaluated rather than
`investigating individual components when problems occur. Although
`prototype pumps or a new design may run the gambit of these problems,
`standard design or "off-the-shelf" pumps are not immune, particularly to
`system problems.
`
`lnstallatlonruaintenanca Ef lects
`
`UnbaLnc!. Unbalance of a rotating shaft can cause large transverse vi-
`brations at certain speeds, known as critical speeds, that coincide with
`the lateral natural frequencies ofthe shaft, l,ateral vibration due to unbal-
`ance is probably the most cornmon cause of downtime and failures in
`centrifugal pumps. Damage due to unbalance response may range from
`seal or bearing wipes to catastrophic failures of the rotor. Excessive un-
`balance can result from rotor bow, unbalanced couplings, thermal distor-
`tion, or loose parts. All too often, field balancing is required elen after
`careful shop balancing has been performed.
`Although a pump rotor may be adequately balanced at startup, after a
`period of operation the pump rotor may become unbalanced by erosion,
`corrosion, or wear. Unbalance could also be caused by non-uniform plat-
`ing of the pumped product onto the impeller. In this instance, cleaning the
`impeller could restore the balance. Erosion of the impeller by cavitation
`or chemical reaction with the product may cause permanent unbalance
`requiring replacement of the impeller. rlbar of the impeller or shaft
`caused by rubs will require the repair or replacement of dre dern ged
`component. Another cause of unbalarrce can occur if lubricated couplings
`have an uneven build-up of grease or sludge.
`Assembly or manufacturing procedures may cause a new pump rotor
`to be unbalanced because of slight manufacturing imperfections or toler-
`ance build-up resulting in the center of mass of the rotor not being exactly
`at the center of rotation. Forging or casting procedues cao produce local
`variations in the density of the meal due to inclusions or voids. On large
`cast impellers, the bore for the shaft may not be exactly centered with the
`casting geometry. Stacking a rotor can result in thermal disortions of the
`shaft or impellers that can result in a cocked impeller. Nonsymmetries of
`just a few mils caused by these manufacturing or assembly methods can
`result in significant forces generated by a high speed roor. Mo6t ofthese
`nonsymmetries can be compensated for by balancing the rotor,
`
`Mlsalignment. Angular misalignment between two shafts connected
`with a flexible coupling introduces an additional &iving force that can
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 7 of 72
`
`

`

`)n
`
`on of the pump (ot its
`ping). This interactive
`evaluated ralher than
`.ems occur. Although
`rbit of these problems,
`nmune, particularly to
`
`lse large transverse vi-
`ds, that coincide with
`vibration duo to unbal-
`rntime and failures in
`Ponse may r4nge from
`te rotor. Excessive un-
`rplings, thermal distor-
`I is required ovetr after
`
`nc€d at startuP, after a
`nbalanced by erosion,
`d by non-uniform plat-
`s instanc€, cleaning tlrc
`impeller by cavitation
`I pennanent unbalance
`the impeller or shaft
`:ment of the damaged
`if lubricated couplings
`
`use a new pump fotor
`imperfections or toler-
`rotor not being exactly
`ures can produce local
`ons or voids. On large
`actly centered with the
`rmal distortions of the
`ler. Nonsymmetries of
`assembly methods can
`,d rotor. Most of these
`ring the rotor.
`
`two shafts connected
`driving forc€
`that can
`
`vibration and Noise in PumF 427
`
`produce torsional or latoral vibrations. The forces in a typical industrial
`coupling are similar to 0rose in a universal joint (Figure l8-2). When a
`small angular misalignment occurs, the velocity ratio across the joint is
`not constant. If one shaft speed is assumed constant, then the other shaft
`has a faster rotational rate [6] for part ofthe revolution and a slower rota-
`tional rate for part of the revolution. This variation of rotating speed re-
`sults in a second harmonic (twice shaft speed) vibrational component.
`
`B
`
`{!) Univ.rs.l Joint
`
`0
`
`(b) Ge.red coupllnq {ith
`An-oular [lisa] i gnmert
`
`(() Pol.r Angul.r Velocity Di.grrn
`Flgure '18-2. Ettects o, anguler misalignmsnl in shatt couplingE.
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 8 of 72
`
`

`

`428 Centrifugal Pumps: Dssign and Application
`
`Piplng and Structure. The pump should be relatively isolated from the
`piping. The weight and thermal loading on the suction and discharge con-
`nections should be minimized. The American Petroleum Institute (API)
`Standard 610 [7] specifies allowable external nozzle forces and mo-
`ments. Most pump manufacturers specify allowable weight and thermal
`loads transferred from the pipe to the pump case. Static forces ftom the
`piping may misalign the pump from its driver, or for excessive loading,
`the pump case may become distorted and cause rubs or seal and bearing
`damage. Thermal flexibility analyses of the piping should be performed
`to evaluate piping loads and to design the necessary supports and re-
`straints to minimize the transfer of piprng loads to the operating equip-
`ment.
`Vibrations of the piping or the support structure can be mechanically
`transferred to the pump. The piping and lhe structure should not have
`their resonant frequencies coincident with any of the pump excitations
`such as vane passing frequency or multiples. The vibrations transferred
`from the pipe to the structure can be minimized by using a visco-elastic
`material (i.e., belting material) between the prpe and the pipe clamp.
`
`Applicalion
`
`The initial stage of pump system design should include the task of de-
`fining the range of operating conditions for pressure, flow, temperatures.
`and the fluid properties. The vendors can provide the correct pump ge-
`ometry for these design conditions. Expected variations in operating con-
`ditions and fluid composition, if a significant percentage, may influence
`the design.
`Improper application or changing conditions can result in a variety of
`problems. Operation at high-flow, low-head conditions can cause vibra-
`tions of the rotor and case. Inadequate NPSH can result in cavitation tha!
`will cause noise and vibration of varying degrees.
`
`Bearings. General purpose, small horsepower pumps in process plars
`generally have rolling element beatings. Noise and vibrations are com-
`monly a result of bearing wear. As the rolling elements or races wear, tht
`worn surfaces or defects initially produce a noise and as wear increa-.e-.
`vibrations may become noticeable. Several vibrational ftequencies m.a..
`occur that depend on the geometry of the bearing components and the::
`relative rotational speeds [3]. The frequencies are generally above o5:-
`ating speed.
`Many ball bearing failures [8] are due to contaminants in the lubn*:r
`that have found their way into the bearing after the machine has be=
`placed in operation. Common contaminants include moisture, dirt. ar
`
`I
`
`other miscellaneous particles
`may cause wear or permanent
`tremendous stresses generated
`purpose pumps and I
`-.,speclal
`fllm (hydrodynam ic) bearings.
`rotllng element bearings forii;
`orodynamic bearing suppors t
`geometry of the hyarodynamic
`portant role in controlling the I
`vtbrational characteristici of ti
`
`Seals. The fluid dynamics of I
`ll- d9r.9ynryu", t9l. Hydrodr
`me stabilization of rotatinq ma
`targe axial flow in the turbilent
`ro produce Iarge stiffness and d,
`robr vibrations and shbiliry. Wr
`and-cause greater leakage and p
`renstrcs of the seal resu-iting in
`
`Hydraullc Eftectg
`
`.Hydraulic effects and pulsatior
`vroratron of the pump oipiping :
`pas$ng frequency arO it" t armo
`c-an, be. caused by acoustical reso:
`Hil:HX.:"fiix",1i#*Ti1
`an_ uneyen pressure d'istributiln r
`rotor.
`
`Ii:::r:T". srarring and stoppins
`r{l,#iilil::l:r#3ffH;
`suoden rmpact.force to thJ pump
`warcr hammer has caused cracks ir
`was anchored.
`closure of con yendonal v:
`" --Rapid
`-€vere water haruner. Increasing tl
`[rE[1[:;l"*d';n*.X
`oon tor various closure rates [ib].
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 9 of 72
`
`

`

`Application
`
`I be relatively isolated from te
`n the suction and discharge e-r--
`rican Fetroleum Institute (AlT
`rternal nozzle forces and m;-
`y allowable weight and therm:-
`mp case. Static forces ftom Se
`lriver, or for excessive loadrnr
`I cause rubs or seal and bearir.:
`the piping should be performc
`lhe necessary supports and re-
`g loads to the operating equip
`
`: structure can be mechanically
`Ithe structure should not have
`th any of the pump excitations
`e.s. The vibrations transferred
`rimized by using a visco-elastic
`the pipe and the pipe clamp.
`
`r should include the task of de-
`r pressure, flow, temperatures,
`r provide the correct pump ge-
`ted variations in operating con-
`)ant petcentage, may influence
`
`itions can result in a variety of
`ad conditions can cause vibra-
`SH can result in cavitation that
`degrees.
`
`Dwer pumps in process plants
`Noise and vibrations are ornn-
`ing elements or races wear, the
`a noise and as wear increases
`rl vibrational frequeacies may
`bearing components and their
`cies are generally above oper-
`
`) contaminants ifl the lubricant
`rg after the machine has been
`ts include moisture, dirt, and
`
`vibration and Noiso in Pumps 425
`
`cher miscellaneous particles which, when tr-?ppeq insid: lle.-bearing'
`;;;;;;;;
`p,i..r*otrv ina"nt rhe balli irnd racewavs under the
`;;;;;" Gt;es'generated bv tte operating load'
`m'*rxtm+:mgm*x*e*:**;+
`;"'itiil;ffiil;#il.." fr" 'it'i " a fttm ot o:iias lt rotates' rhe
`Il]j"iI"rir'. rt"[.o.iin"*i" be;in! and the oil properties ptay a1 ry-
`:ffi;ii#;:ffiiii; d;i"t"a "?ti"a
`speedi and consequentlv the
`iitrational draracteristics of the pump'
`Secls. The fluid dynamics of flow tlrough sealslrave a dramatic effect
`;;;;"il;-;;;r-i91 Hvdrodvnamic fories involved mav contribute to
`;;;'$^bilil"ti* oi iotuting .h.t'i""ty or make it unstable' s€als with
`i"t" *i"r nt"' in ttre turtitem r"ngei such as in feed water Pumps' tend
`fi ;;"fr; ilil;l+;;;; "d
`e*''pt'i coefficients. that are-be^nefi cial to
`;;fi;il;il;"t-# stabititv. wearbr 6e seals will increase theclearance
`ffi;;;;Hleakage and possibly change tre rotordynamic charac-
`teristics of-the seal resulting in incroased vibrations'
`
`Hydraullc Effects
`Hvdraulic effects and pulsations can result in almost any frequeircy of
`p"rp-o.'pipirg from once per r.wo.lu$on up tothe vane
`oassinq frequencv and its trarmonics. frequencias below running sqed
`"iirLrrii"r,iiJ
`Hfff;tHil""outti""ii*n"nce s' Generally' these effects are due
`;,#illl*;;;;;thJit"ilse
`diftuser.or some other discontinuitv
`in the case. Any nonsynrmetry of ihe internals of the pump may produce
`p;rrt" disttiuuti'on that can result in foices applied to the
`;;;;;
`rotor.
`
`Ttamlents. Starting and stopping pumps with the attendant opening and
`closing of valves is a major cause of severe transients in piping systems'
`the rJsulting pressue surge, referred to as water hammer' can apply a
`r"Ja", ir"pi"f f"rce to rhJ pump, its internals, and the piping' Severe
`fr".it"t tt.. caused oracks in concrete structures to which the pipe
`t
`"t"i
`was anchored.
`n""iJ"t*rr" of conventional valves used in fecdwater lines can cause
`.";;[;iln
`,nii. tncreasing the ctosure time of the valve can reduce
`thc severitv of the surge pr"st,it". Analytical rnethods arc available to
`*."irity oi *'utt hammer in a porticular piping configura-
`"i,"ilJui,,
`tion for various closure rates [10].
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 10 of 72
`
`

`

`430 Centritugal Pumps: Design and Application
`Cavltatlon and Fla3hlng. For many liquid pump piping systems, it is
`common to have some degree of flashing and caviation associated with
`the pump or with thc prcssure control valves in the piping system. High
`flow rates produce more severe cavitation because ofgreater flow losses
`through restrictions.
`Cavitation produces high local pressures that may be traosmitted di-
`rectly to the Frmp or piping and may also be transmitted through the
`fluid to otlrcr areas of the prping. Caviation is one of the most commonly
`occurring and damaging problems in liquid pump systems. The term
`cavitation refers !o the formation and subsequent collapse of vapor bub-
`bles (or cavities) in a ligid caused by dynamic pressure variations near
`the vapor pressure. Cavitation can produce noise, vibration, loss of head
`and capecity as well as severc erosion of the impeller and casing sur-
`faces,
`Before the pressure of the liquid flowing through a centrifugal pump is
`increased, the liquid may experience a pressure &op inside the pump
`case. Ihis is due in part to acceleration of the liquid into the eye of the
`impeller and flow separation from the impeller inlet vanes. If flow is in
`excess of design or the incident vane angle is incorrect, high-velocity,
`low-pressure eddies may form. If the liquid pressure is reduced to the
`vaporization prcssure, the ligid will flash. Later in the flow path the
`pressure will increase. The implosion which follows causes what is usu-
`ally referred to as cavitation noise. The collapse of the vapor pockets,
`usually on the nonpressure side ofthe impeller vanes, causes severe dam-
`age (vane erosion) in addition to noise.
`When a centrifugal pump is operated at flows away from the point of
`bast efficiency, noise is often heard amund the pump casing. The magni-
`tude and freguency of this noise may vary from pump to pump and are
`dependent on the magdtude of the pump head being generated, the ratio
`of NPSH required to NPSH available, and the amount by which actual
`flow deviates from ideal flow. Noise is often generated when the vane
`angles of the inla gurdes, imFller, and diffuser are incorrect for the ac-
`tual flow rate.
`Cavitation can best be recognized by observing the complex wave or
`dynamic pressure variation using an oscilloscope and a pressure tratrs-
`ducer. Ihe pressure waveform will be non-sinusoi&l with sharp maxi-
`mum peaks (Cpike$ and rounded minimum peaks occurring at vapo'
`plessure as shown in Figurc 18-3. As the pressure drops, it canncr
`produce a vacuum less than the vapor pressure.
`Cavitation-like noise can also be heard at flows less than design, evea
`when available inlet NPSH is in excess of purp required NPSH, and this
`has been a puzzling problem. Ar explanation offered by Fraser I I, 12]
`suggests that noise ofa very low, random ftequency but very high inten-
`
`'o]
`
`_.1
`
`lr
`
`(.) c.!i
`
`{tz
`!sa
`.9 36
`618
`
`0
`
`(b) ConDt.x !r
`Shoiing ar
`Flgwo t8\r. Cavh
`
`sity results from backflow r
`charge, or both. Every centri
`tain conditions of flow reduc
`can be d"maging to the pressr
`Ier vanes (and also to casinp ,
`ctease in loudness of a barisi
`suction and/or discharse orlr
`Sound levels mea"rrfo 'rt
`rt
`slction piping during cavitati<
`tron prcduced a wide_band si
`ever, in this case, the vane pa
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 11 of 72
`
`

`

`lltl
`
`ication
`
`Pump ptptng systems, it is
`I cavitation associated with
`in the piping system. High
`nuse of greater flow losses
`
`hat may be transmitted di-
`be trartsmited through the
`i one of the most commonly
`pump systems. The term
`rent collapse of vapor bub-
`tic pressure variations near
`rise, vibration, loss of head
`e impeller and casing sur-
`
`rough a centrifugal pump is
`surc drop inside the pump
`e liquid ino ttre eye if thi
`cr inlet vanes. If flow is in
`is incorrect, high-velocity,
`pressure is rcduced to the
`Later in the flow path the
`bllows causes what is usu-
`rpse of the vapor pockets,
`vanes, causes severe dam-
`
`ws away from tho point of
`r pump casing. The magni-
`)m pump to pump and are
`being generated, the ratio
`e anrount by which actual
`generated when fte vane
`lr are incorrect for the ac-
`
`ring the complex wave or
`ope and a pressure trans-
`tusoidal with sharp maxi-
`peaks occurring at vapor
`rressure drops, it camot
`
`rws less than deqign, even
`r rcquircd NPSH, and this
`rffered by Fraser [l l, 12]
`ency but very high inten-
`
`Vibration and Noise in Pumps 431
`
`lr
`
`tt
`
`lt
`
`rr
`
`rt
`
`,=0
`
`(.) C.vtr.tion Theor),
`lf Pd, P3 - Plp than cavitltion rill occur
`P5 = St.tic Prarru..
`Pd' 0ra.riC P!lratio.s, te.o.0..l
`Pvp= !rpo.0resturc
`
`ll
`
`t
`ilI
`IilTfl
`
`?
`
`08
`
`90
`
`72
`
`54
`
`36
`
`lg
`
`0
`
`5c i!/div
`{b) Conplet I.veforn or P.esrvre
`snori.g Iff€.t. or C.vit.aiot
`Flguro 1&3. Cavitalion etfects on lhe dynamic praosur€
`
`sity results from bacKlow at the impeller eye or at the impeller dis-
`charge, or both. Every centdfugal pump has this recirculation under cer-
`tain conditions of flow reduction. Operation in a recirculating condition
`can be damaging to the pr€ssure side of the inlet and/or discharge impel-
`ler vanes (and also to casing vanes). Recirculation is evidenced by an in-
`crease in loudness of a banging type, random noise, and an increase in
`suction and/or discharge pressure pulsations as flow is decreased.
`Sound levels measured at the casing of an 8000 hp pump and near the
`suction piping during cavitation [2] are shown in Figure 18.4. The cavita-
`tion ptoduced a wide-band shock that excited many frequencies; how-
`ever, in this case, the vane passing frequency (number of impeller vanes
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 12 of 72
`
`

`

`432 Centrifugal Pumps: Design and Application
`
`5 ll(lat tior lwllon
`----- 6 llrclltt ,tot qnl .lst
`-
`
`i6
`
`t^xt trt'tc
`Itt0u[,r(r
`
`^il)
`
`Flgure lH. Noiso spsctra of cavitation in cenlrifugal pump.
`
`times revolutions per second) and multiples of it predominated. Cavita-
`tion noise of this type usually produces very high ftequency noise, best
`described as " crackling."
`Flashing is particularly common in hot water systems (fe€dwater pump
`systems) when the hot, pressurized water experiences a decrease in pres-
`sure through a restriction (i.e., flow control valve). This reduction of
`pressure allows the liquid to suddenly vapotae, or flash, which results in
`a noise similar to cavitation. lb avoid flashing after a resricdon, suffi-
`cient back pressure should be provided. Alternately, the restriction could
`be located at the end of the line so that the flashing energy can dissipate
`into a larger volume.
`
`Flow TUrbulence, Pump generated dynamic pressure sources include
`turbulence (vortices or wakes) produced in the clearance space between
`impeller vane tips and the stationary diffrrser or volute lips, Dynamic
`pressure fluctuations or pulsations produced in this manner can cause im-
`peller vibrations or can result in shaft vibrations as the pressure pulses
`impinge on the impeller.
`Flow past an obstruction or restriction in the piping may produce tur-
`bulence or flow-induced pulsations [2]. These pulsations may produce
`both noise and vibration over a wide-frequency band. The frequencies
`are related to the flow velocity and geometry of the obstruction. These
`pulsations may cause a resonant interaction with other parts of the acous-
`tic piping system.
`Most of these unstable flow patterns are produced by shearing at tle
`boundary b€tween a high-velocity and low-velocity region in a fluic
`
`field. Typical examples of rl
`obstructions or past deadwatr
`bidirectional flow. The shear
`are converted to pressure Defl
`localized vibration excitaton
`acoustic natural response mor
`the turbulence has ; strong in
`ttls voriex shedding. Experir
`tex flow is more severe when
`the generation frequency of th
`turbulent energy centeied ar,
`with a dimensionless Strouha.
`f:$y
`D
`where f: vortex frequency,
`S" = Strouhal numbei.
`V = flow velocitv in t
`D:acharacteriiticdi
`For flow past tubes, D is the
`past a branch pipe, D is the ins
`Strouhal equation is firrrher de
`an example, flow at 100 ft/se
`produce broad-band turbulencr
`slub were acoustically resonani
`tlon amplitudes could result.
`Pressure regulators or flow r
`ated with both turbulence and f
`ating with a severe pressure dro
`ate significant turbulence. Althr
`broad-band, it is characteristic
`sponding to a Strouhal number
`Pu.lsatlons. pumping systems
`pulsatlons through normal ounr
`tions occur from mechanisms w
`rn a centrifugal pump are generi
`-betwe
`lp9_n the clearance space
`drttuser or volute lips, the instal
`symmeffy of the pump rotor an(
`Tculatgly known, predicting th
`loenucal pumps often have diffr
`
`Exhibit 1130
`Bazooka v. Nuhn - IPR2024-00098
`Page 13 of 72
`
`

`

`ld Application
`
`6 lxfltt tiot Sucnoli flnic
`6 |lrortt rlm! PUi' cisa
`
`rz
`
`alion in contrifugal pump,
`
`of_it predominated. Cavita-
`'Ies
`/ery tugl frequency noise. besr
`water systems (feedwater DumD
`exDenences a decrease in'Drc"._
`rrot valve). This reductitin of
`rrze, or flash, which results in
`:mng atter a restric(on. suffi_
`rr:rn_atglx the restriction could
`, rusnrng energy can dissipate
`
`uic pressure sourcts include
`r lne cleaEDce space between
`user
`_or volute lips. Dynamic
`r ln 0lis manner can cause im_
`?uons as the pressurc pulses
`the piping may produce tur-
`tese pulsations may produce
`'ency- qand. The freqlencies
`rl.ot the obsEuction. These
`vrm other parts of the acous_
`prc{uc<d by shearing at the
`'-veroclty region itr a fluid
`
`Vibration and Noiso ln Pumps 433
`
`field. Typical examples of this type of turbulence include flow around
`obstructions or past deadwater regions (i.e., a closed bypass line) or by
`bi{irectional flow. The shearing action produces vorlices, or eddies that
`are converted to pressure perturbations at the pipe wall that may result in
`localized vibration excitation of the piping or pump components. The
`acoustic natural response modes of the piping system and the location of
`the turbulence has a strong influence on the frequency and amPlitude of
`this vortex shedding. Experimental measurements have shown that vor-
`tex flow is more severe when a system acoustic resonance coincides with
`the generati