experimental investigation on unsteady pressure pulsation in a...

Ning Zhang1

School of Energy and Power Engineering,

Jiangsu University,

Zhenjiang 212013, China

e-mail: [email protected]

MinGuan YangSchool of Energy and Power Engineering,

Jiangsu University,



Bo GaoSchool of Energy and Power Engineering,

Jiangsu University,



Zhong LiSchool of Energy and Power Engineering,

Jiangsu University,



Dan NiSchool of Energy and Power Engineering,

Jiangsu University,



Experimental Investigationon Unsteady Pressure Pulsationin a Centrifugal Pump WithSpecial Slope VoluteRotor–stator interaction, a major source of high amplitude pressure pulsation andflow-induced vibration in the centrifugal pump, is detrimental to stable operation ofpumps. In the present study, a slope volute is investigated to explore an effective methodto reduce high pressure pulsation level, and its influence on flow structures is analyzedusing numerical simulation. The stress is placed on experimental investigation ofunsteady pressure pulsation inside the slope volute pump. For that purpose, pressure pul-sations are extracted at nine locations along the slope volute casing covering sensitivepump regions. Results show that distinct pressure pulsation peaks at fBPF, together withnonlinear components are captured. These peaks are closely related to the position ofpressure transducer and operating conditions of the pump. The improvement of rotationalspeed of the impeller results in rapid increase of pressure fluctuation amplitude at fBPF

and corresponding root mean square (RMS) value within 10–500 Hz. A comparison withconventional spiral volute pump is implemented as well, and it is demonstrated that slopevolute contributes significantly to the decline of pressure pulsation level.[DOI: 10.1115/1.4029574]

Keywords: centrifugal pump, slope volute, unsteady pressure pulsation, experimentalvalidation

1 Introduction

Unsteady pressure pulsation induced by rotor–stator interactionhas a direct impact on stable operation of the centrifugal pump.Even at nominal flow rate, flow discharged from the impellermatches the geometry of volute to a favorable extent, but pressurepulsation still occurs due to the interaction between the impellerand the volute [1]. Consensus has been attained that flow field dis-tribution is not circumferentially uniform in the volute, and theunevenness is closely associated with not just the geometricaldesign of pump components but also operating conditions. Somestudies have been dedicated to unsteady pressure pulsation byeither experimental or numerical method, and most of them laidtheir emphasis upon the influence of geometrical parameters onpressure pulsation characteristics [2]. In essence, the principalpurpose of probing into unsteady pressure pulsation is to open thepossibility of modeling the intense rotor–stator interactionbetween the impeller and the volute tongue. Furthermore, aneffective approach should be developed to control pressure pulsa-tion during pump design, which is sorely necessitated.

Spence and Aaral-Teixeira [3] numerically investigated theeffect of four geometric parameters, namely blade tip clearance,vane arrangement, shroud-to-casing radial clearance, and sidewallclearance, on pressure pulsation in a centrifugal pump. And thecontribution of the four parameters is thereby ranked, as isremarkably important for pump design. Yao et al. [4] analyzedtime-frequency characteristics of pressure pulsation in a double-suction centrifugal pump using fast Fourier transform (FFT)

algorithm and time-frequency representation methods. Under off-design conditions, some unexpected flow phenomena occurringinside pumps have considerable effect on pressure pulsation. Bar-rio et al. [5,6] predicted flow pulsations associated with rotor–sta-tor interaction using numerical simulation for different flow rates.The study facilitated a relation between geometrical parameters ofblade passage and pressure pulsations. And both tangential and ra-dial velocity components at reference locations in near-tongueregion were taken into account as well. Parrondo et al. [7] alsoinvestigated the effect of operation point on pressure pulsation ina centrifugal pump, and particular attention was paid to pressureamplitude at blade passing frequency. Increasing blade tip clear-ance serves as an effective measure for reducing pressure pulsa-tion, as illustrated by Yang et al. [8,9]. Besides, to lower pressurepulsation in the centrifugal pump, specifically devised impellerand volute structure such as the splitter blade and double volutehave been presented [10,11]. However, relative position of the vo-lute tongue with respect to the impeller has rarely been examined.Therefore, the major aim of this study is to seek an effectivemethod to alleviate rotor–stator interaction by using the slope vo-lute and then to investigate the unsteady pressure pulsation char-acteristics under various working conditions. Similar volutestructure has been employed in hygienic pumps [12], but relevantpressure pulsation has not been addressed.

In present study, unsteady pressure pulsation characteristics ina centrifugal pump with slope volute are investigated from shutoffto maximum flow rates. Pressure pulsation signals are obtainedwith nine fast-response pressure transducers mounted on theslope volute casing. Detailed analysis of pressure spectrum is per-formed, and special attention is paid to pressure pulsation peaks atblade passing frequency and some nonlinear interaction frequencycomponents. Meanwhile, the influence of rotational speed on pres-sure spectrum is investigated. Finally, to validate the low pressure

1Corresponding author.Contributed by the Fluids Engineering Division of ASME for publication in the

JOURNAL OF FLUIDS ENGINEERING. Manuscript received August 5, 2014; finalmanuscript received January 11, 2015; published online March 11, 2015. Assoc.Editor: Bart van Esch.

Journal of Fluids Engineering JUNE 2015, Vol. 137 / 061103-1Copyright VC 2015 by ASME

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pulsation characteristic of the slope volute, pressure pulsation in acentrifugal pump with conventional spiral volute is measured forcomparison.

2 Experimental Setup

The single-stage model pump with an axial suction is equippedwith an impeller incorporating six backward curved blades.Figure 1 presents a structure comparison of the slope volute andconventional spiral volute. For the diffusion part of the slopevolute, a deviation of 15 deg from the vertical axis is adoptedbased on optimal results [13]. In Fig. 2, radial dimensions of theslope volute casing are identical, and cross-sectional area of thevolute passage develops along axial direction. Eight cross-sectional areas of the two volutes are compared in Fig. 3, whichshows small difference between the two volutes. Meanwhile, theinlet and outlet areas of the two volutes are equivalent. And thetwo impellers are identical as well. Besides, the same blade tipclearance is adopted for the two pumps to ensure comparableexperimental results. In Fig. 1, it also can be seen that the tongueof the slope volute locates at the right side of the impeller, so theflow emanating from the impeller does not interact directly withthe volute tongue. But the tongue of the spiral volute facesthe impeller immediately. Therefore, it is inferred that intenserotor–stator interaction in the slope volute pump would be attenu-ated obviously relative to the spiral volute pump, which had beenvalidated in previous study through numerical simulation [14].

Experiments were conducted in a closed-loop test rig to guaran-tee measurement accuracy, as shown in Fig. 4. Flow rate of themodel pump was measured by electronic flowmeter with anabsolute accuracy of 60.2%. The head of the model pump wasmeasured with an uncertainty of 60.1%. Flow rate of the modelpump could be regulated by means of a gate valve installed at the

inlet of the tank. The pump was driven by an alternating current(motor, and a frequency inverter was used to adjust the rotationalspeed of the model pump continuously. The rated rotational speedof the model pump is 1450 r/min, which is kept invariant fromshutoff to maximum flow rate conditions. But at some special tests(variable rotational speed experiment), the rotational speed wouldbe adjusted.

For the slope volute pump, nine fast-response pressure trans-ducers (PCB113B27 series) are mounted on the slope volute

Fig. 1 Structure comparison of the slope volute and the conventional spiral volute

Fig. 2 Configurations of cross section changing of twovolutes

Fig. 3 Comparison of cross section areas between two volutes

Fig. 4 Closed test platform

061103-2 / Vol. 137, JUNE 2015 Transactions of the ASME


casing, as shown in Fig. 5. All the transducers are located at themidspan of impeller outlet. From p1 to p7, the angle between twoconsecutive transducers is 15 deg, and from p7 to p9, thisangle increases to 45 deg. Frequency tailoring and high resonantfrequency (�500 kHz) enable an extremely wide available fre-quency range (beyond 0–100 kHz) of the pressure transducers.The response time of the transducer is less than 1 ls, and theuncertainty of measured signals is usually lower than 60.2%. Inthe centrifugal pump, frequencies excited by hydrodynamicphenomena usually fall into a low frequency range, which is nor-mally narrower than 1 kHz. During signal sampling process, thesampling frequency was set as 10 kHz to satisfy Nyquist samplingtheorem, so the detailed information associated with the originalones is well retained in obtained signals. Sampling resolution wasset as 1 Hz, and Hanning-window was applied during samplingprocess. Time domain signals were transformed into frequencydomain signals using auto-power spectrum algorithm, aspresented in the following equation:

Fig. 5 Configuration of pressure transducers mounted on the slope volute

Fig. 6 Performance of the model pump

Fig. 7 Absolute velocity distributions on the sixth cross section at different flow rates insidetwo volutes

Journal of Fluids Engineering JUNE 2015, Vol. 137 / 061103-3


SxðxÞ ¼ð1�1

RxðsÞe�jxsds (1)

Rx(s) describes the correlation between given signal x(t) andx(tþ s), as shown in the following equation:

RxðsÞ ¼ limT!0

1

T

ðT

0

xðtÞxðt6sÞdt (2)

where T represents observation time of x(t).

3 Results and Discussions

3.1 Performance of the Model Pump. Figure 6 showsexperimental performance results of the slope volute pump fromshutoff to 1.45Ad conditions. Nominal flow rate and head of themodel pump is 48 m3/h and 7.5 m, respectively. The best effi-ciency point of the model pump is near the nominal flow rate. Atlow flow rates, positive slope (curve instability phenomenon)occurs on head performance curve of the model pump, which

indicates that unsteady rotating stall phenomenon develops insidethe impeller channels under these working conditions [15]. So alarge-scale separate-flow region may occur in the impeller chan-nels, which would cause global instability within the model pump.Therefore, unsteady pressure pulsation characteristics may beaffected significantly. From comparison with the spiral volutepump, it is found that the efficiency of the slope volute pump islow at high flow rates. But at low flow rates, the discrepancy israther small, and the two pumps almost have equivalentefficiency.

To clarify the reason underlying low efficiency of the slopevolute pump, Fig. 7 presents velocity distributions on the sixthcross section of the two volutes. And the numerical simulationmethod can be found in our preceding study [13]. At nominal flowrate, it is evident that two counter-rotating vortices of nearly thesame scale are captured in the spiral volute, and the vortex centerskeep almost symmetrical with respect to the midspan plane of theimpeller. In the slope volute, similar flow pattern is obtained, butthe scales of the two vortices are obviously different. The small-scale vortex faces the impeller outlet, while the large vortexlocates at the left of the impeller. But with flow rate decreasing, at

Fig. 8 Comparison of flow structures of two pumps under Ad: (a) relative velocity at impellermiddle section, (b) absolute velocity at the sixth cross section, and (c) absolute velocity aroundthe volute tongue



0.6Ad, the small vortex facing the impeller disappears, and only alarge-scale vortex dominates the whole flow passage. The reasonfor low efficiency of the slope volute pump is probably related toflow structures in volute. At high flow rates, partial flow in thesmall-scale vortex region is suppressed by large-scale vortex,which could not be transported synchronously by the main streamtoward volute outlet. Therefore, high energy loss in this regionwould be expected. But this phenomenon does not occur at lowflow rates and in the spiral volute, all fluid would flow throughvolute smoothly causing low energy loss. So it is inferred fromnumerical results that low efficiency of the slope volute pumpmay be caused by vortex motion in the volute. To reduce the neg-ative effect of small vortex, changing the shape in the small vortexregion may be an effective way to improve the performance of theslope volute pump, which would be analyzed in the future study.

3.2 Influence of Slope Volute on Flow Structures in theCentrifugal Pump. It is essential to analyze the influence ofslope volute on flow structures inside the model pump. Figure 8shows a comparison of flow distributions in slope and spiralvolute pumps at nominal flow rate. Relative velocity distributionson middle section of the impeller passage are presented inFig. 8(a). It is observed that flow patterns of the two pumps aresimilar and almost the same in different blade channels. There-fore, it is concluded that the slope volute has little influence onflow structures inside impellers compared with the spiral volute.Figure 8(b) shows absolute velocity distributions at the sixth crosssection of two volutes. According to the results indicated inFig. 7, different flow structures in volutes may be a reasoncausing low efficiency of the slope volute pump at high flow rates.Figure 8(c) shows flows near volute tongues. In the conventionalcentrifugal pump, it is well accepted that the interaction betweenflows discharged from the impeller with jet-wake pattern and thevolute tongue causes high amplitude pressure pulsation. However,in the slope volute pump flow pattern changes. It is evident thatflow is divided into two parts by the volute tongue, one at the leftof the volute tongue and the other at the right of the volute tongue.Flows discharged from the impeller do not interact immediatelywith the volute tongue but travel smoothly along the circumfer-ence of the volute toward diffuser outlet. And the interaction withthe volute tongue occurs at the inlet of the diffuser. But with flowmoving for a long distance, the pulsation level of the unsteadyflow would be attenuated significantly. So the interaction effectwith the volute tongue would be reduced rapidly. Therefore, it isinferred that pressure pulsation inside the slope volute is muchlower compared with that in the spiral volute pump.

3.3 Unsteady Pressure Pulsation of the Model Pump.Unsteady pressure pulsation experiments of the model pump

were conducted at different flow rates. Low frequency signals in0–10 Hz frequency band were filtered during signal processing.To evaluate pressure pulsation energy in particular frequencyband, RMS method was applied to deal with discrete pressuresignals as presented in Eq. (3)

RMS ¼ 1

0:5qu22

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi1

n

Xn

n¼1

ðAn � �AÞ2s

(3)

�A ¼ 1

n

Xn

n¼1

An (4)

where An represents pressure amplitudes at different frequencies,and �A is mean amplitude.

In the centrifugal pump, the gap between the rotating impellerand the stationary volute is small, leading to intense pressurefluctuating [16,17]. Figure 9 shows time history of unsteadypressure pulsation signals of sensor p7 at four typical flowrates. As observed, the pressure signals fluctuate obviously. Under

off-design conditions, especially at low flow rate of 0.2Ad, pres-sure pulsating amplitude is much larger than that at nominal flowrate, which is associated with the separate-flow structure develop-ing in the model pump. According to Fig. 6, unsteady rotating stallphenomenon occurs within the impeller channels around 0.2Ad.The global instability inflow, of course, influences the overallhydrodynamic performance characteristics of the model pump andinduces rather high amplitude pressure pulsation.

To investigate the influence of operating condition and the posi-tion of transducer on pressure pulsation characteristics, pressurespectra at sensors p3 and p7 are presented in Fig. 10 at four flowrates, namely 0.1Ad, 0.6Ad, 1.0Ad, and 1.3Ad. In the centrifugalpump, rotor–stator interaction is a primary reason for high ampli-tude pressure pulsation. And blade passing frequency fBPF and itshigher harmonics 2fBPF, 3fBPF, and so on are expected thereby.From Fig. 10, it is easy to identify fBPF and its higher harmonics atdifferent flow rates except the pressure sensor p3 at 0.1Ad due toits small amplitude. At extremely low flow rate of 0.1Ad, someprominent discrete components at fR, 2fBPF, 3fBPF, and 4fBPF canbe identified for sensors p3 and p7, but peaks at other higher har-monic of fBPF are not evident. Generally, it is accepted that pres-sure pulsation amplitudes at higher harmonics are lower than thatat basic frequency. But this argument is not always suitable forthe present case. As observed, at low flow rates of 0.1Ad and0.6Ad, predominant components in pressure spectra locate at3fBPF, and the corresponding amplitudes are larger than fBPF.Especially for p3 at 0.1Ad, it is difficult to identify discrete peakat fBPF. Besides, a large number of high amplitude components atlow frequencies are generated, which are probably caused byunsteady separate-flow inside the model pump.

With flow rate increasing, pressure amplitude at 3fBPF decreasesrapidly at high flow rates, exhibiting an opposite tendency relative

Fig. 9 Unsteady pressure signals of sensor p7 at four typicalflow rates



to that at fBPF. Especially at 1.3Ad, pressure amplitude at 3fBPF ismuch smaller than that associated with fBPF. Also, it is easy to findthat pressure spectrum is sensitive to the position of pressuretransducer. At 1.0Ad and 1.3Ad, predominant components corre-spond to fBPF at p3, but spikes at fR remain predominant at p7.From the above analysis, it is concluded that component at fBPF

does not always dominant pressure spectrum, and its magnitudeis often less than other frequencies. The reason for low fBPF

magnitude probably lies in the fact that energy at fBPF is oftenredistributed among other higher harmonics and nonlinear interac-tion components hence lowering the fBPF intensity [18].

As illustrated in Fig. 10, at some operation conditions, theamplitude at component fR is often larger than fBPF, so fR is a criti-cal basic frequency. A large number of nonlinear componentswould be generated due to nonlinear interaction between fR andfBPF, as well as their higher harmonics, having the form ofmfBPFþ nfR, where m and n are integers. At low flow rate of0.1Ad, several nonlinear frequencies could be identified corre-sponding to higher harmonic of fBPF and to its sum anddifference frequencies with higher harmonic of rotating frequencyfR. Table 1 presents some evident nonlinear components. Withflow rate increasing, flow field inside the model pump ismuch uniform compared with that at low flow rates, therefore

large-scale separate-flow structure no longer exists. At nominalflow rate, both the number and amplitude of nonlinear compo-nents decrease significantly. As a consequence, only one evidentnonlinear component corresponding to 3fBPF� fR could be found.On the contrary, energy at discrete components fR and fBPF isimproved rapidly at p7. So it is concluded that the increase ofamplitudes at fR and fBPF is at the expense of energy decrease attheir higher harmonics and nonlinear components [19,20].

According to Fig. 10, frequencies associated with remarkablepeaks in pressure spectra are usually lower than 4fBPF. To illus-trate the relationship between pressure pulsation energy and flowrate, RMS method is used to process pressure signals rangingfrom 10 to 500 Hz. Figure 11 presents RMS values as a functionof flow rate for different pressure transducers. It is seen in Fig. 11that the position of pressure transducer influences not only pres-sure spectrum but also energy distribution of discrete signals inparticular frequency band. In Fig. 11(a), RMS values of differentpressure sensors almost have similar variation tendencies. Whenthe model pump operates near nominal flow rate, pressure pulsa-tion energy reaches a minimum level. At off-design conditions,either high or low flow rates, RMS values increase rapidly, espe-cially for sensors p5, p6, and p7. Generally, at partial loadunsteady flow phenomena would be easy to develop inside theimpeller passage. These flow phenomena are represented by flowseparation from blade suction side, vortex shedding at the trailingedge of the impeller, which would result in a sharp increase ofpressure pulsation energy. But at extremely low flow rates from0.1Ad to shutoff conditions, pressure pulsation energy decreases.Also it is noted that RMS values of different sensors show appa-rent difference at low flow rates, especially at 0.2Ad, RMS valueof transducer p7 is almost 1.6 times of its counterpart associatedwith p3.

However, in Fig. 11(b), pressure pulsation energy exhibits asignificant difference compared with that shown in Fig. 11(a).

Fig. 10 Pressure spectra of pressure sensors p3 and p7 at different flow rates

Table 1 Some identified nonlinear components

Value (f/fBPF) Nonlinear component

2.83 3fBPF� fR3.17 3fBPFþ fR3.83 4fBPF� fR3.66 4fBPF� 2fR3.49 4fBPF� 3fR



Pressure sensors p1and p2 have similar variation tendencies, andRMS values increase first, then experience a local maximum valuearound 0.5Ad, finally fall as flow rate increases. At high flowrates, rapid increase in RMS value is not found. For pressure sen-sors p8 and p9, the trends change again. As observed, at low flowrates, sharp increase in RMS value does not occur. At p9, RMSvalue almost keeps increasing from shutoff to maximum flowrates.

From the above analysis, it is clear that different positions ofthe slope volute are associated with different pressure fluctuationcharacteristics. In the centrifugal pump, due to the asymmetricgeometry of the volute, circumferential pressure distributioninside the volute is not uniform. As demonstrated, even at nominalflow rate, the unevenness of static pressure distribution along thevolute casing exists, as also applies to unsteady pressure pulsation.So it is concluded that pressure pulsation energy does not have thesame tendency along the slope volute casing due to nonuniformflow field distribution.

Angle distributions of pressure amplitudes at fBPF on peripheryof the slope volute for various flow rates are presented in Fig. 12.As observed, from 75 deg to 108 deg, pressure amplitudes havesimilar variation tendencies with respect to flow rate. Usually,rotor–stator interaction is intense near the slope volute tongue,namely around 45 deg, where maximum pressure amplitudescould be detected for flow rates ranging from 0.8Ad to 1.4Ad. Butan opposite trend is observed at low flow rates ranging from0.2Ad to 0.6Ad. And at 45 deg, pressure amplitude is smaller thanthose at other measuring points. At low flow rates, the exit angle

of flow discharged from the impeller deviates significantly fromthat at nominal flow rate, which has an obvious influence on flowdistribution near the slope volute tongue. According to our previ-ous work [13], partial fluid flows back from the diffusion sectionpassing the volute tongue, hence the interaction pattern betweenjet-wake flow structure and the volute tongue would probablybe disturbed due to backflow structure. According to Akin andRockwell [19], pressure amplitude may be decreased as a result ofturbulent mixing in the centrifugal pump. So it is inferred that thedecrease of pressure amplitude around the slope volute tongue atlow flow rates is probably related to turbulent mixing of backflowstructure and rotor–stator interaction.

As illustrated in Fig. 10, some evident peaks occur in pressurespectrum at given flow rate, and Fig. 13 shows two distinct peaksat impeller rotating frequency fR and higher harmonic of bladepassing frequency 3fBPF. It is noted that pressure amplitudes at fRand 3fBPF fluctuate slightly at flow rates lower than 1.0Ad. At highflow rates (A> 1.0Ad), pressure amplitude at fR increases gradu-ally, while an opposite trend is seen for 3fBPF. Furthermore, pres-sure amplitude drops rapidly as flow rate increases. Situations atfR and 3fBPF support the energy redistributing conjecture amongdiscrete components in pressure spectrum. As observed, energyincrease at fR is at the expense of rapid energy decrease at 3fBPF.

Fig. 11 RMS trends of different measuring points versus flowrate

Fig. 12 Angle distributions of pressure amplitudes at fBPF

along the slope volute casing for different flow rates

Fig. 13 Pressure amplitudes at distinct peaks fR and 3fBPF ver-sus flow rate



3.4 Influence of Rotational Speed on Pressure PulsationCharacteristics. In the centrifugal pump, strong rotor–statorinteraction, together with intense pressure pulsation, is excited asthe impeller blades pass the volute tongue successively. So pres-sure spectrum characteristics, such as the amplitude of discretecomponent and energy distribution in particular frequency band,are closely related to rotational speed of the impeller [21]. For themodel pump operating at nominal flow rate, Fig. 14 presentspressure spectra of pressure sensor p3 at three rotational speeds,namely 1450 r/min, 1200 r/min, and 1000 r/min. At nominal rota-tional speed of 1450 r/min, some discrete spikes corresponding tofR, fBPF, and their harmonic frequencies can be identified. Asobserved, peaks at fBPF always dominate pressure spectra at p3 fordifferent rotational speeds. When rotational speed decreases to1200 r/min, amplitudes at fBPF and 3fBPF decrease evidently. Withrotational speed decreasing further, at 1000 r/min, amplitude atfBPF decreases continuously, while the component at 3fBPF couldnot be identified easily. But pressure amplitude at 4fBPF increasesrapidly as rotational speed decreases.

For the model pump operating at different rotational speeds,Fig. 15 shows both RMS values and amplitudes corresponding tofBPF at p1 and p3. It is seen that both RMS value and amplitude atfBPF increase as rotational speed increases. At nominal flow rate,from 1000 r/min to 1480 r/min, RMS value has an increment of103% at p1 and 62% at p3. Under off-design conditions, espe-cially at low flow rate 0.8Ad, the increment increases rapidly, andRMS value increases by 150% at p1 and 113% at p3. Amplitudeat fBPF shows a similar trend relative to RMS value. At nominalflow rate, from 1000 r/min to 1480 r/min, amplitude increases by183% at p1 and 46% at p3.

In the centrifugal pump, low frequency vibration signalsinduced by pressure pulsation have a great influence on mechani-cal components of the pump. As shown in Fig. 15, pressurepulsation energy decreases rapidly as rotational speed of theimpeller decreases. So it is concluded that adjusting rotationalspeed of the impeller is an effective measure for reducing vibra-tion energy. Finally, the negative impact of vibration on mechani-cal components of the pump and even the pumping system wouldbe weakened significantly.

3.5 Validation of Low Pressure Pulsation LevelCharacteristic of the Slope Volute. To relieve rotor–stator inter-action in the centrifugal pump, a special slope volute is investi-gated. Based upon theoretical analysis of flow structures in Fig. 8,the slope volute is characterized by low pressure pulsations, espe-cially at fBPF. To validate our conjecture, pressure pulsation insidea centrifugal pump with conventional spiral volute was measured

for comparison. The same impeller was used for the two pumps.In the presence of the size of the spiral volute, only three pressuretransducers were mounted on periphery of the spiral volute forcomparison, as shown in Fig. 16. The transducers were located atmidspan of impeller outlet, as was implemented for the slopevolute pump. And rotational speed of the impeller was kept invari-ant 1450 r/min for all examined flow rates.

In the centrifugal pump with the conventional spiral volute,fBPF often dominants pressure spectrum, which is also an impor-tant forcing frequency inducing low frequency vibration signals in

Fig. 14 Pressure spectra of p3 at three rotational speeds undernominal flow rate

Fig. 15 Influence of rotational speed on pressure pulsation atthree different flow rates: (a) RMS value in 10–500 Hz frequencyband and (b) amplitude at fBPF

Fig. 16 Detail of pressure transducers mounting on theconventional spiral volute



fluid machineries [22–24]. To illustrate the influence of slopevolute on fBPF, Fig. 17 presents a comparison of amplitude at fBPF

between the slope volute and the spiral volute pumps. Asobserved, experimental results agree with our conjecture verywell. It is found that pressure amplitude of the spiral volute pumpis much larger than that in the slope volute pump from shutoff tomaximum flow rate conditions. Besides, variation of pressure pul-sation for the slope volute pump is relatively smooth, as impliesthat flow parameter distributions inside the slope volute pump aremore uniform compared with that inside the spiral volute pump.For p7, evident discrepancy is present between the two pumpsespecially at low flow rates. At 0.4Ad, amplitude for the spiralvolute pump nearly reaches 4.9 times of that for the slope volutepump, and 4.3 times at 0.5Ad. At nominal flow rate, there is still adifference of 1.9 times. For measuring points p8 and p9, similarresults are observed for all flow rates in consideration. Addition-ally, amplitude of the spiral volute pump is three times of that forthe slope volute pump for p8 and 3.2 times for p9.

To evaluate the influence of the slope volute on low frequencysignals in 10–500 Hz band, a comparison of RMS values betweenthe two pumps is performed in Fig. 18. It is noted that RMS valueof the spiral volute pump is much larger than that of the slopevolute pump at p7 for all measured flow rates. In particular, underoff-design conditions, the difference is salient. For pressure sen-sors p8 and p9, similar trends can be found at low flow rates, andRMS value of the spiral volute is larger than that of the slopevolute pump. But at nominal flow rate, the difference between the

two pumps is not significant, and RMS values of the two pumpsare pretty close.

The main purpose to investigate unsteady pressure pulsationcharacteristics is to seek an effective way to reduce pressureamplitude in the centrifugal pump. In most relevant literatures,attentions are often paid to the influence of geometric parameterson pressure amplitude. In this paper, by using a special slopevolute, the intense interaction between flow discharged from theimpeller and the volute tongue would be attenuated obviously.According to the comparison results, the slope volute has an evi-dent impact on reducing pressure energy in the centrifugal pump,especially for the amplitude at fBPF.

4 Conclusions

An experimental investigation on unsteady pressure pulsationfeatures of a centrifugal pump with slope volute is presented. Theanalysis focuses on pressure pulsations at nine different measuringpositions on periphery of the slope volute. Pressure pulsation sig-nals measured from shutoff to maximum operating conditions areanalyzed using auto-power spectrum algorithm and RMS method.Pressure spectrum characteristics, both discrete component andRMS value, are easily affected by operating conditions and meas-uring position on the slope volute casing. At low flow rates, thepredominant component in pressure spectrum often deviates fromfBPF to its high harmonic frequency. And some nonlinear compo-nents are captured due to nonlinear interaction between fR andfBPF. With flow rate increasing, both the number and amplitude ofnonlinear components decrease obviously.

Meanwhile, the influence of rotational speed on pressure pulsa-tion is investigated. It is found that pressure amplitude at fBPF andRMS value increase rapidly as rotational speed increases. Gener-ally, low vibration signals are caused by pressure pulsation, whichare critical to stable operation of the pump. Thus, low rotationalspeed of the impeller may be more appropriate during pumpdesign process considering low vibration characteristic of the cen-trifugal pump.

To assess low pressure pulsation level of the slope volute, pres-sure pulsation of a conventional spiral volute pump is measuredfor comparison. It is observed that pressure pulsation inside theslope volute pump is much smaller than that inside the spiralvolute pump, especially for amplitude at fBPF. So it is confirmedthat intense rotor–stator interaction in the centrifugal pump is sig-nificantly attenuated by employing the slope volute.

Finally, it is expected that the present work will provide a dif-ferent view considering lower pressure pulsation requirement ofthe centrifugal pump and contribute to a better understanding ofunsteady pressure pulsation. In further study, experimental inves-tigation on optimal design and unsteady flow field associated withthe slope volute will be conducted to further lower pressure pulsa-tion level and to improve high pump efficiency as well.

Acknowledgment

The authors gratefully acknowledge the financial support ofNational Natural Science Foundation of China (51476070 and51206063), a Project Funded by the Priority Academic ProgramDevelopment of Jiangsu Higher Education Institutions (PAPD),and the Research and Innovation Project for College Graduates ofJiangsu Province (KYLX_1036).

Nomenclature

A ¼ pressure amplitude (Pa)b2 ¼ impeller outlet width (17 mm)d2 ¼ impeller outlet diameter (172 mm)cp ¼ pressure coefficient (A/0.5qu2

2)fR ¼ rotating frequency of impeller (24.2 Hz)

fBPF ¼ blade passing frequency (145 Hz)g ¼ efficiency

Fig. 17 Comparison of amplitudes at fBPF between two pumpsat various flow rates

Fig. 18 Comparison of RMS values between two pumps



nd (x) ¼ rotational speed (1450 r/min (151.8 rad/s))ns ¼ specific speed (ns ¼ 3:65nd

ffiffiffiffiffiffiQd

p=H0:75

d , 130)u2 ¼ tangential velocity at blade outlet (13 m/s)Z ¼ blade number (6)q ¼ water density (1000 kg/m3)

Ad ¼ nominal flow rate coefficient Qd= xd22b2

� �� Wd ¼ nominal head coefficient gHd= x2d2

2

� �� References

[1] Brennen, C. E., 1994, Hydrodynamics of Pumps, Concept ETI Inc./OxfordUniversity Press, White River Junction, VT/Oxford, UK.

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