Performance and cavitation analysis of a waterjet system on a cavitation tunnel

Date Author Address

October, 2003

Prof.dr.ir. T.J.C. van Terwisga Delft University of Technology Ship Hydromechanics Laboratory Mekelweg 2, 26282 CD Delft

Phone: +31 15 2786873

TUDelft

Delft University of Technology

Performance and cavitation analysis of a

Waterjet system on a cavitation tunnel

by

T.J.C. van Terwisga

Report No. 1401-P

October 2003

FAST2003, The 7th International Conference on Fast Sea Transportation, Ischia, Italy

Università di Napoli "Federico II"

FAST 2003

Ischia (Italy)

7'

-

10' October 2003

The 7th International Conference

on

FAST SEA TRANSPORTATION

PAPERS (VoL 1)

Pasquale Cassella

Editor

Dipartimento Ingegneria Navale - Università di Napoli "Federico II"

Istituto di Navigazione "G. Simeon" - Università di Napoli "Parthenope"

"The Organizing Committee FAST 2003" shall not be responsible for statements

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These Proceeding,s consist of papers presented at the 7th International Conference on

FAST Sea Transportation. The Conference was held at the Hotel Continental Terme,

Ischia Porto (Italy), on 7-10 October 2003.

_{The Conference was Organised by:}

Department of Naval Architecture and Marine Engineering

University of Naples "Federico II"

Institute of Navigation "G. Simeon"

University of Naples Parthenope

Under the patronage of:

Senato della Repubblica Italiana

Marina Militare Italiana

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FINCANTIERI S.p.A.

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_Napoli

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VOLAVIAMARE

INTERNATIONAL COMMITTEE

Mr. K. Holden, MARINTEK A/S, Norway (Chairman)

Mr M. Bewick, The Harbor Consultancy Int. USA

Prof. L. Doctors, UNSW, Australia

Prof. O. Faltinsen, NTNU, Norway

Prof. G. Ferrara, University of Naples "Parthenope"

Mr. N. Gee, N. Gee and A. Ltd. UK

Prof. T. Moan, NTNU, Norway

Mr. J.A. Moret, IZAR Construcciones Navales, Spain.

NATIONAL COMMITTEE

Dr. Ing. C. Antonini

,

FINCANTIERI S.p.A

Ing. G. Bernardi, CETENA S.p.A.

Dr. U. Bulgarelli, INSEAN (Rome Towing Tank)

Prof. A. Campanile, University of Naples, "Federico II"

Prof. A. Cardo, University of Trieste

Ing. G. De Domenico, Tugboats Company of Naples

On. S. Lauro, Shipping-Line "Alilauro"

Prof. L. Nicolais, University "Federico II" of Naples

Adm. Gen. Isp. (G.N.) E. Piantini, Italian Navy

Prof. C. Podenzana Bonvino,University of Genoa

Adm. Gen.Isp. (C.P.) E. Sicurezza, Italian Coast Guard

Ing. N. Squassafichi Nicola, RINA (Italian Register of Shipping)

ORGANISING COMMITTEE

Prof. P. Cassella (Chairman), University of Naples, "Federico II"

Prof. R. Balestrieri, University of Naples "Parthenope"

Dr. C. Bertorello University of Naples, "Federico II"

Prof. G. Russo Krauss, University of Naples, "Federico II"

Prof. A. Scamardella, University of Naples "Parthenope"

FAST 2003 Secretariat

Dr. Carlo Bertorello

Mr. Pasquale Cioffi

Department of Naval Architecture and Marine Engineering

University of Naples "Federico II"

Via Claudio 21 80125, Naples (ITALY)

Tel. +39-081-7683310 /3714

Fax +39-081-2390380

e-mail

fast2003@unina.it

Table of Contents

Vol. 1

Preface

Opening Lectures

The Fast Ship Fever

From Trondheim To Ischia,The Gulf Of Naples.

1 Kjell O. Holden

High Speed Craft - Regulations And Efforts To Minimize Accidents

5

Eugenio Sicurezza

Keynote Lectures

The Optimisation Of Trimaran Sidehull Position For Minimum Resistance

1

Lawrence J Doctors, Robert J Scrace

Hydrodynamic Aspects Of High-Speed Vessels

₁₃

Faltinsen 0.M, Claudio Lugni, Maurizio Landrini

The Effects Of Length On The Powering Of Large Slender Hull Forms

₂₃

Nigel Gee, James Roy

Towards Structural Design Of High Speed Craft Based On Direct Calculations

31 Torgeir Moan

Simulation Based Resistance And Seakeeping Performamce Of High-Speed Monohull And Multihull

Vessels Equipped With Motion Control Lifting Appendages

₅₁

Paul D. Sclavounos, Shiran Pun;in, Ta/ha Ulusoy, Sungetin Kim

Developments On A Probabilistic Risk/Cost Model For Large-Scale Flooding Consequence Analysis

Of High Speed Monohulls

₆₇

Dimitris Konovessis, Daria Cabaj, Dracos Vassalos

A New Generation Of Large Fast Ferry - From Concept To Contract Reality

₇₅

Tony Armstrong, Kjell Holden

Session Al Hydrodynamic Performance

From Model Scale To Full Size. Investigation On Turbulence Stimulation In Resistance Model Tests

Of High Speed Craft

1

Carlo Bertorello (invited speaker,) Dario Bruzzone, Sebastiano Caldarella ,Pasquale Cassella, Igor Zotti

Hydrodynamic Improvements Of Catamaran Hulls When Using Streamlined Bodies Of Revolution

9

Igor Zotti (invited speaker)

Transom Hollow Prediction For High-Speed Displacement

_{Vessels 19}

Simon W Roba rds, Lawrence J Doctors

Series 64 Parent Hull Displacement And Static Trim

Variation ₂₇

Gabor Karafiath, Todd Carrico

Hullform Analysis And Optimisation For A Fast Ropax Ferry

₃₁

II

A Critical Survey On Performance Assessment For Classical And Advanced Ships

39 Luigi Iannone

The Effect Of Demihull Separation On The Frictional Resistance Of Catamarans

47

Tony Armstrong

Performance And Cavitation Analysis Of A Waterjet System On A Cavitation Tunnel

57

Klaas Kooiker Tom Van Terwisga, Rob Verbeek, Peter Van Terwisga

The Accuracy Of Wave Pattern Resistance Determination From Wave Measurements In Longitudinal

Cuts

63

Nastia Degiuli, Andreja Werner, Zdravko Doliner

Influence Of Hull Shape On The Resistance Of A Fast Trimaran Vessel

71

Antonio Cardo, Marco Ferrando, Carlo Podenzana-Bonvino

Hydrodynamic Performance And Exciting Force Of Surface Piercing Propeller

79 Kazuo Nozawa

Drag Reduction On A High Speed Trimaran

87 Robert Latorre, Aaron Miller; Richard Philips

Hydrodynamic Developments Of 154m Class Ro-Ro Passenger Ferry

93 Soon Ho Choi, Jung Joong Kim, Gun Ho Lee, Se Eun Kim, Sung Mok Ahn

On The Hydrodynamic Performance Of High Speed Craft

101

C. Bertorello, S.Brizzolara. D. Bruzzone, P Cassella, I. Zotti

A Comparative Analysis Of The Resistance Qualities Of A Series Of Semi-Displacement Hi-Speed

Mono-Hull Forms

109

Jacques. B. Hadler And Richard Vanhooff

Session A2 Propulsion

Applicability Of A Rim Drive Pod For High Speed Ship Propulsion

1

John Richards,Jon Eaton, Juergen Friesch, Michelle Lea, Donald Thompson, Bill VanBlarcom

Design Of A Waterjet Propulsion System For An Amphibious Tracked Vehicle

7

M. C. Kim, H. H. 'hun W. G. Park

Enhanced Performance Scaling Methodology

17

Jon E. Eaton, Eckhard Praefke, Friedrich Mewis

Erosion Problems On Fast High Powered Ships

27 Jiirgen Friesch

Design of optimal inlet duct geometry based on vessel operational profile

35 NWH. Bulten, R. Verbeek

Investigations About The Use Of Podded Drives For Fast Ships

41

Cornelia Heinke, Hans-JUrgen Heinke

Dynamic Performance Analysis of a Gas Turbine/Waterjet Propulsion System for a Fast Trimaran

Ferry

₄₉

Giovanni Benvenuto, Ugo Campora

Minimising The Effects of Transom Geometry on Waterjet Propelled Craft Operating In The

Displacement and Pre-Planing Regime

59

James Roy, John Bonafoux

Latest Developments in Fast Ship Gear Technology

₆₇

Franz Hoppe

Session A3 CFD

Fully Nonlinear Wave-Wave Interactions Between A High-Speed Vessel And

Incident Waves 1

Ray-Oing Lin, Arthur M. Reed. William Belk-nap

A Numerical Method For Performance Prediction Of Hydrofoil-Assisted

_Catamarans

₉

PERFORMANCE AND CAVITATION ANALYSIS OF A WATERJET SYSTEM ON A

CAVITATION TUNNEL

Klaas Kooiker, Tom Van Terwisga, Rob Verbeek, Peter Van Terwisga,

SUMMARY

To study the effect of intake working point (defined by Intake Velocity Ratio IVR) on pump and waterjetperformance, experiments with a Waterjet System mounted on a Cavitation Tunnel have been conducted. The effect of the intake on efficiency and head and torque coefficient for both the pump and the complete jet system were studied, as well as the

inception of cavitation on the pump. The tests have been performed on a mixed flow waterjet system that is

representative for current industrial standards. The effect of IVR on pump head coefficient and pump efficiency was tested for two working points. Inception of cavitation was determined for the design IVR and for an off-design IVR. At these IVR's, cavitation inception was found for several flow rate coefficients by lowering the ambient pressure in the tunnel. The paper clearly shows that pump-intake interaction effects are present and need to be accounted for. The test set-up described appears suitable to determine the overall jet system characteristics and to assess theeffect of the intake on pump performance.

MARIN, The Netherlands

MARIN, Delft University of Technology, The Netherlands Wärtsilä Propulsion Netherlands, The Netherlands

Royal Netherlands Navy, The Netherlands

1. INTRODUCTION

Little is known about cavitation inception of waterjet

pumps when mounted in the complete wateijet system. Most often, when cavitation inception is discussed in the

literature, the pump working point is meant where the

pump head starts to decrease for lower ambient pressure

(or increasing specific suction speed nss). Head

breakdown criteria of 1 or 3% are often used. In this situation however, the impeller usually shows already a lot of cavitation and consequently produces a lot of noise. Furthermore, these "inception points" are usually determined in a pump loop set-up without an intake that absorbs boundary layer flow, that was built up along the

hull of the ship.

Figure 1: Test set-up of waterjet system in MARIN

Large Cavitation Tunnel

Session Al

57

For naval applications, cavitation is to be avoided as long

as possible. The question remains however at what

specific suction speed the impeller, or any other part of

the jet system, shows the first signs of cavitation and

starts to produce excessive noise. Because cavitation

inception is very sensitive to local flow distortions, it is important to model the local flow conditions at the pump

inlet as good as possible. This was already argued by

Kruppa [1993], who cites results on the effect of several upstream bend configurations on pump efficiency and

torque and head coefficient. And these results dealwith

integral quantities! The effect of an intake distorted flow on cavitation inception is likely to be more pronounced than it is on the integral quantities.

The objective of the study presented here is twofold; The first objective is to obtain more knowledge on cavitation inception limits of a representative waterjet pump. The second objective is to obtain insight in the effect of the

distorted flow, provided by the intake to the pump, on

pump performance.

This paper presents the results of an experimental study of the effect of intake working point on both the integral quantities such as pump efficiency and torque and head

coefficient, as well as on cavitation inception number

based on

visual identification of cavitation in a

representative waterjet system. A detailed description of

the test set-up is included.

Furthermore,

results of the current test

set-up are

compared with those from a traditional pump loop set-up without the presence of a waterjet intake and an ingested

2. TEST SET-UP

The test set-up consisted of a waterjet system with retour

conduit mounted on top of a cavitation tunnel. A

photograph of the test set-up is shown in Figure 1. A more detailed drawing of the waterjet mounted in the tunnel cover is presented in Figure 2. The aeometry of

the

waterjet was provided by Wärtsilä Propulsion

Netherlands.

+ 4 + 1- +

1-+

PORSHLOL - A

Figure 2: Waterjet system mounted on tunnel cover The experimental set-up enabled independent variation of the intake operating point (IVR) and the pump working point (KQ), allowing for a study of the effect of IVR on pump performance Visualisation of cavitation inception

on the impeller and the stator was made possible by using a Perspex pump housing, containing both the

impeller and the stator. The working points of the pump and the intake could be adjusted independently.

2.1 BOUNDARY LAYER AND ARTIFICIAL THICKENING

LDV measurements on the boundary layer velocity

distribution have been conducted in the past for a smooth tunnel wall and a tunnel wall fitted with a serrated strip

as depicted in. w

Strip 1: w=30 mm, h=19 mm, 11=40 mm, 12=30 mm,

b=510 mm, a =approx. 33 deg

Figure 3 Serrated strips for boundary layer thickening The serrated strip is fitted to the tunnel top, at the point

where the contraction in the tunnel connects to the

parallel part of the tunnel. This

is some 630 mm

upstream of the intake ramp tangency point.

Session Al

58

The boundary layer velocity profile has been measured with LDV equipment by Van Terwisga [1993] for two transverse locations and two streamwise locations. The sen-ated strip resulted in a boundary layer thickness of approx. 28 mm at a downstream position of the strip of 480 mm at a tunnel velocity of 7 m/s.

In line with MARIN's best experience, carborundum

grains were alued in the leading edge area of the impeller blades with an average size of 40 Jim over a chord length of some 4 mm f-rom the leading edge to induce transition to turbulence in the boundary layer.

2.2 TRANSDUCERS AND DATA REDUCTION

The objective of the tests is to determine the waterjet system hydraulic performance characteristics and to

determine its cavitation characteristics. To this end, the

following transducers were mounted:

Flowrate transducer

An Electromagnetic Flow Meter (type IFM 4080 K from Krohne-Persenaire) was used during the tests to measure

the flow rate. This meter was, following the

specifications, mounted in the return duct, well away

from disturbances in the conduit that might have affected

the measurements.

Head measurement

To determine the hydraulic performance of the system,

total head over the pump and the waterjet system are

important parameters. The total head is defined by the

following equation:

H = -1 p(t7,25, (P, 7153) p g

-2 where

= average energy velocity in longitudinal direction at position i;

1 f f

Q jAj,

= flow rate

= velocity along longitudinal co-ordinate

= unit normal vector rectangular through stream

tube cross sectional area

= Area

= mean static pressure at position i (see figure 4)

= mean z

co-ordinate (z co-ordinate positive pointing vertically downward)

It is seen from eq. (2) that the velocity distributionacross

the relevant cross section A, is needed to determine the

average energy velocity Fie4.

In the absence of the

detailed flow field, the average energy velocity ii-eg is approximated by the average volumetric velocity Wig

Figure 4: Definition of station numbers and normalized energy fluxes according to ITTC [1996]

The mean volumetric velocity is however only equal to the average energy velocity in a uniform flow field. As

the actual flow field at the pump inlet (pos. 3) and the pump outlet (pos. 5) is non-uniform, an error is

introduced here. Estimates of the error introduced by this

simplification can be calculated for distinct velocity

distributions. Scherer et al. [2001] present the following table of deviations:

Table 1 Typical energy factors for various velocity distributions (mean velocity based on unit free stream

velocity)

Session Al

59

the measuring system. The total absolute pressure at any pressure tap i can thus be obtained from:

= pa,, + Po+ pgzi (5)

where

P,,

= atmospheric ambient pressure measured with

a mercury barometer mounted near the cavitation tunnel

Po pressure measured at the centreline cavitation

tunnel, relative to the atmospheric pressure

= pressure measured at station i, relative to pressure at centreline tunnel Po

zi vertical co-ordinate of station i, relative to centreline cavitation tunnel. Positive z direction pointing downward.

Torque and rotation rate

A torque transducer was mounted between the sealing of the watetjet pump and the electric motor drive (after the

1:1 transmission). The torque transducer was designed to measure a maximum average torque of approx. 125 Nm. During most of tests, the torque transducer measured in the range of 10-20% of its maximum. The calibration of the torque transducer in its measuring range was checked with a static load in "built in" condition, so as to make sure that no residual torque due to the set-up would affect the calibration. The results from the static load test were

all within 2% of the calibrated torque. As the original

calibration was considered more accurate, it was decided to use this original calibration. A rotation rate transducer was mounted on the outgoing shaft of the electric motor (in front of the 1:1 transmission to the pump shaft). This rotation rate transducer was checked with an independent transducer prior to testing. The difference in results of these two meters appeared to be well within 0.5%.

Tunnel velocity

The tunnel velocity or free stream velocity was measured

from the pressure difference in the contraction of the

tunnel, upstream of the measuring section. This pressure

difference signal has been calibrated

by LDV

measurements to arrive at the most accurate estimate of

the tunnel flow velocity (outside the tunnel boundary

layer).

3. REVIEW OF TESTS

Three types of tests have been conducted:

1. Jet system performance

'"). Cavitation Inception

3. Cavitation observations

The Jet System Performance tests can be subdivided into three distinct types:

Exploratory tests where pump rpm and tunnel

velocity were fixed (at n = 1250 rpm and Uo = 6 m/s)

and where the pump working point (in terms of KO

Type of velocity distribution Energy

factor l3E

Mean

velocity Uniform vel. distribution 1.0 1.0

Linear Distribution

O - Free stream 2.0 0.50

1/3 Free stream - Free stream 1.25 0.67

1/2 Free stream - Free stream 1.11 0.75

Exponential distribution

Exp = 1.0 (linear) 2.0 0.50

Exp= 0.5 (parabolic) 1.35 0.67

Exp = 1/7 (boundary layer) 1.05 0.88

The lcinematic contribution to the energy flux _canE/cm be expressed in terms of the energy factor by:

43Ek, = Q. pV 2 0,1

2

(4)

The average static pressure T? is approximated from 5

pressure taps fitted circumferentially at the required station at equidistant positions. The z co-ordinate is measured relative to the centreline of the cavitation tunnel, where the origin of the co-ordinate system is positioned, the z axis being positive when pointing

vertically downward. All pressure transducers measured a pressure difference between the actual pressure tap and the ambient tunnel pressure at the centreline tunnel. This ambient tunnel pressure is referred to as Po. Prior to every

has been varied by gradually shutting of the valve in the return duct

Jet system performance tests at two distinct operating points of the intake (IVRri values of 1.6

and 1.2)

Jet system performance test to get an appreciation of the effect of IVR, on jet system performance. To this

end, two distinct pump operating points were

adjusted (1(Q=0.40 and 0.37)

Cavitation inception tests were conducted to determine at

which specific suction speed number nss, cavitation

would first start to occur visually.

To this

end,

visualisation

of the impeller and stator was made

possible by the Perspex pump housing, containing both the impeller and the stator (see Figure 5).

Figure 5 Photograph of Perspex pumphousing

For each cavitation inception test, a certain flow rate coefficient K.Q at a prescribed IVRi. was adjusted by adjusting the valve for a pre-set tunnel and impeller

velocity. Cavitation inception was subsequently found by lowering the ambient pressure in the tunnel. The limits in the flow rate coefficients over which inception could be found were set by the limits of the water tunnel and test set-up. The non-dimensional pressure at which cavitation inception would occur is defined by the specific suction

speed nss, defined by:

nfd

nss= (gHs)Y4 where 1 , Hs

=-Q9t+

Po+ P3meos + pgT3 Pg

Pv = vapour pressure of water

The suction height Hs is the pressure at which incipient cavitation was observed first. One can understand that for

an increasing non dimensional rotation rate, the

resistance against cavitation decreases, that is for

increasing suction specific speed.

Session Al

60

Finally, cavitation observation tests were conducted with the aim to detect possible cavitation forms that could be

erosive in nature.

4. DISCUSSION OF RESULTS 4.1 JET SYSTEM CHARACTERISTICS

Figures 6, 7 and 8 show the jet system characteristics in

non-dimensional form: torque coefficient, pump head

coefficient and pump efficiency for three distinct tests. The scatter in results of these tests are essentially caused

by differences in intake working point IVR. as will be

discussed in the following. For reference purposes, the pump characteristics as derived from tests in a different

set-up on a slightly larger pump diameter (factor 1.6

times bigger) have been plotted in the same graphs. It is

shown that significant differences with these earlier results occur. These differences may be caused by the different set-up of the tests. It has been demonstrated

here that the intake worldng point does have a significant

effect. This is caused by the intake velocity profile,

offered to the pump. As the tests in the pump loop circuit did not have the possibility to easily and truthfully model the intake working point, it is considered to be the major cause for the discrepancies measured. Another possibility for the deviation could be the presence of roughness on

the leading edge of the impeller blade. Based on

MARIN' s experience with propellers, this effect is

considered of secondary importance.

Figure 6: Jet system characteristics: Torque coefficient

1.4 1.3 1.4 .3 O 1.2 1.1 t 1.0 20 0.8 a 0.7 Ic 0.5 Torque coefficient 1, 1.1 .11P1111... -.11PP%111.

' , '41IM

I

1.0 o 8

Mmi...

"

.16M it 0.7 -o- 254 mm intake 0.6 -0-160 mm intake 1 0.5 . -6- ivrr=1.6

=

0.4 II -8-ivrr=1.2 05 0.6 0.7

Pump head nondimensional

-- 254 mm intake

0160 mm intake

-ivrr=1.6

0.4 -,--ivrr=12

05 0.6 0.7

Ko/Kc best performance

(6)

2P1732 Pv) (7)

0.8 09 1.0 11 1.2

08 0.9 1.0 1.1 1.2

Ko/K. best performance

Figure 7 Jet system characteristics: Pump head

; 1.0 0.9 0.8 0.7 of' -è 0.8 1.20 0.92 0.90 1.1 0.5 254 mm intake -6- 160 /TIM intake s- ivrr.1.6 Pump efficiency K./Kb beat performance

Figure 8: Jet system characteristics: Pump efficiency 4.2 REPEATABILITY OF RESULTS

In order to get an appreciation of the repeatability of the results, results that were obtained at the same pump and intake worlcing point (KQ and IVR respectively) were compared with each other. IVR identity was selected as an additional constraint, as it was clearly demonstrated that this intake working point did significantly affect the

results.

An indication of the repeatability of the torque

coefficient could be obtained from the jet performance test series at constant IVR, the cavitation inception tests at constant IVR and the corresponding IVR values in the

IVR variation tests at constant KQ. This gave a number of

21 pairs of results that could directly be compared with each other with the same independent variables KQ and IVR. From this comparison, a sample standard deviation in torque coefficient was found of 0.83%.

4.3 EFFECT OF IVR ON PERFORMANCE

The effect of the Intake Velocity Ratio IVR, on head

coefficient and pump efficiency is presented in Figure 9. Pump characteristics as RIVR)

Figure 9: Effect of IVRR on head coefficient and pump

efficiency

The coefficients are normalized relative to the

coefficient's value

at an IVRR=1.2, which can be

regarded as the design point of the intake, where the most uniform flow is delivered by the intake. It is seen that the head coefficient increases with 3% for an IVRR=1.6 and

even up to 8% for an IVRR=2.0. The effect of IVR is

greatest for the lightest pump loading (KQ=0.40).

Session Al 61

A similar trend is seen in the pump efficiency 17e, although the effects are somewhat increased due to

similar trends in torque coefficient.

The effects of IVR on torque coefficient are presented in Figure 10. The torque coefficient decreases for increasing

IVRR. A decrease of 2.5% at an IVRR=1.6 occurs,

whereas the torque coefficient decreases even with some 7% for the IVRR=2.0. Azain, the effect is strongest for

the lightest pump loading.

Figure 10: Effect of IVRR on torque coefficient 4.4 EFFECT OF IVR ON CAVITATION INCEPTION

Figure 11 shows the results of the cavitation inception

tests. It is seen that tip vortex cavitation occurs first,

followed suit by sheet cavitation at the outer radii of the impeller. In addition, cavitation was observed first in the blade top position, where the axial velocity is relatively low due to the intake flow and impeller shaft wake, and where the pressure is also at a low due to its high vertical

position.

Cavitation could only be noticed on the stator at very high specific suction speed numbers nss, which were

adjusted during the cavitation observation tests.

2 1.8

Pump characteristics as t( IVR)

Cavitation inception

09 0.92 0.94 0.96 0,98 1 1.02 1.04 1.06 1.08 11 K01K0 best pertorrnance

Figure 11: Results of cavitation inception tests

Although the results in Figure 11 show a corresponding tendency, the reliability of two points can be questioned; Sheet cavitation inception at the highest loading (lowest

KQ) and tip vortex cavitation (tvc) inception at the lowest

loading (highest K,Q). Both questionable points were

measured for an off-design rvRR of 1.6.

...4-Kh,Kg.0.40 -s--etap. 6,0.40 _

_-..

-KR 69=0.37 --/, -etap, Kg-0.37 em ...." ...., ...-- ...--onti=1.6 tvc inc

--4-ivrr=1.6, sheet inc --ii-ivrr=1.2, tvc inc

--11-ivrr=1.2, sheet inc

...10 :idge."'"AgelEl . -1.10 1.30 1.50 1.70 1.90 2.10 VRh 05 0.6 07 0.8 09 1.0 1.1 1.2 13 1.10 R 1 00

/-..

Kcm0.40 -ea -61g, 6,0.37 0.80 0.90 1.10 1.30 1.50 1.70 1.90 2./0 IVR

For decreasina, pump loadinas (increasing flow rate

coefficient KO, cavitation inception occurs at higher nss values. For the flow rate coefficient KQ at the maximum pump efficiency point, the specific suction speed number

at which incipient tip vortex cavitation is found is the

approximately the same for both IVRR values of 1.2 and 1.6. Incipient sheet cavitation is found for slightly lower ambient pressures. The sheet cavitation seems to be more

sensitive to the pump input velocity field than the tip

vortex cavitation, as the difference between nss values for IVRR = 1.6 and 1.2 at maximum efficiency is about

12%.

For a 8% lighter pump loading, cavitation inception can

be delayed until approx. 1.7 times the nss value at the

pump loadina at maximum efficiency.

During the cavitation inception and observation tests.

only suction side cavitation has been observed.

Figure 12: Photograph of developed cavitation

5. CONCLUSIONS

The current study has demonstrated the applicability of waterjet system tests on a waterjet that was mounted on

top of a cavitation tunnel. The complete range of flow

rate coefficients of interest could be investigated, without the additional use of another pump in the circuit to adjust intake and pump working point independent from each

other. The conduit resistance was relatively small and

further reduced by some 50% by the suction effect of the outlet scoop. By applying a transparent pump housing,

cavitation could be observed on both the impeller and

stator . In addition, the intake lip could be observed

through Perspex windows in the intake and throual

windows in the cavitation tunnel.

The pump characteristics derived from the present tests

are _{in good qualitative agreement with the pump}

characteristics measured earlier on a sliahtly bigger

pump (scale factor 1.6) in a conventional pump loop set-up. The absolute values differed however (some 2 % in efficiency at maximum efficiency point).

This is most likely ascribed to differences in the test set-up. The repeatability of the new test set-up was found to

Session Al

₆₂

be adequate. The torque coefficient was checked and showed a sample standard deviation of s=0.83% on a

sample of 21 pairs of measurements that could be

compared directly.

Clear effects of the intake working point, expressed in

IVRR, on pump performance and cavitation inception were measured. Effects in head and torque coefficient

were measured up to some 3% between a design and an off design condition for the intake at the lightest pump loading condition tested. Due to accumulating effects, the effect in efficiency appeared to be some 5% for this case. Cavitation inception started with tip vortex cavitation at the suction side of the impeller blade, followed suit by

sheet cavitation at

the outer radii of the

impeller.

Inception of sheet cavitation appeared to be more

sensitive to intake working point than inception of tip

vortex cavitation. For a 8% lower loading of the pump,

the cavitation inception number found for sheet

cavitation appeared to be some 70 % higher. Based on

the cavitation extent, the character and position of the cavitation at inception, it is expected that cavitation

inception can be sigmfficantly improved by modifying the

impeller design. It is, however, likely

that such a

modification will be at the cost of pump efficiency.

Based on the present study, it

is concluded that for

reliable measurements of pump performance, including cavitation, it is important to properly model the complete waterjet system, including a hull boundary layer.

REFERENCES

1. _{British Standard BSI Guide 7405:1991; 'Guide to}

selection and application of flow meters for the

measurement of fluid flow in closed conduits', Aug. 30 1991, ISBN 0 580 19335 7

ITTC Quality Manual; 'HMI Speed Marine Vehicles

- Waterjets', Section 4.9-03-05.2, 21st ITTC 1996

Kruppa, C.; 'Contribution to the 20th ITTC

workshop on waterjet propulsion Pump Installation

efficiency', 1993

Scherer, O., Mutnick, I and Lanni, F.; 'Procedure for conducting a towing tank test of a waterjet propelled craft using laser doppler velocitimery to determine the momentum and energy flux', 26th ATTC, Webb

Institute, July 23-24, 2001

Van Terwisga, T.J.C., Radstaat, G and Van der

Weerd, G; 'LDV measurements on waterjet intake

MARIN Data Report', MARIN Report