Optimum performance screws for large tankers

note inside cover

Delft

NATIONAL PHYSICAL

LABORATORY

SHIP DIVISION

OPTIMUM PERFORMANCE SCREWS FOR LARGE TANKERS

by

T. P. O'Brien

(Reprint of article in Shipbuilding and

Shipping Record 5th May 1966)

A Station of the

Ministry of Technology

SHIP REP, 88

November 1966

Crown Copyright Reserved

Extracts fronithis report rna.y be reprOdnced

provided the source is aclthowledged.

_

Approved on behalf of Director, NFL by

Mr. A. Silverleaf, Superintendent of Ship Division

Reprinted from Shipbuilding and Shipping Record, May 5, 1966

Synopsis

This article discusses recent publica-tions on the effects in designing screws, of varying the screw diameter and rate of rotation; in particular, with

applica-tion to large diameter tanker screws

operating at low rates of rotation. It

gives a procedure for making perfor-mance estimates for series of screws of constant diameter and varying rate

of rotation. It refers to a systematic

series of propulsion experiments com-prising propulsion tests with model

hulls of varying hull shape and model screws of different diameter, the results

of which are given in a form enabling

the effects of variation in aperture size

and screw diameter to be studied.

It

describes optimum diameter and blade area charts used

in making screw

design calculations, and it comprises

worked examples on making propulsion

estimates and design calculations for

the screws for large tankers. The propulsion

estimates were made for

Propellers

Optimum performance screws for large tankers

Optimum diameter and blade area charts

71.11C109E6S RATIO 0.045

Fig. 1 Troost B-4 series, single screws

Ship three groups of screws, each of constant

diameter and over a range of rate of

rotation. The results of these estimates

were applied in making the design

calculations for a tanker screw of large diameter (D =29.5ft-8.99m) designed to

operate at a low rate of rotation (N,--75 revolutions per minute) and for a

delivered horsepower of 22,000 DHP. The vessel, powered by a turbine

installa-tion, was the sister ship of a vessel

powered by a diesel engine and propelled

by a screw of moderate diameter

run-ning at a moderate rate of rotation

(D 23.25ft-7.09m, N,--110 revolutions per minute). For the vessel with the

moderate diameter screw the trial speed

was:

V,=17 knots and

the screw

efficiency and propulsive efficiency were: = 0.485 and =0703, respectively.

For the vessel with the large diameter screw the corresponding values were

Vs=17i knots,' no = 0.575 and n = 0.77, respectively.

Thus the adoption of a

low speed turbine installation in lieu

of a moderate speed diesel engine

'4

I Z

07

06

0.5

Fig. 2 Van Manen

T. P. O'Brien, C.G.I.A., M.R.I.N.A., Division, National Physical Laboratory

enabled a larger diameter screw to be

fitted, and the corresponding improve-ments in performance were: an increase

in trial speed of 3%; an increase in

screw efficiency of 18% and a related

increase in propulsive efficiency of 10%.

1.

Introduction

RECENT PUBLICATIONS (References 1 to

4), some of which are based on research work at NFL, have shciwn that there are significant differences between the per-formance of moderate diameter screws running at moderate rates of rotation and that of large diameter screws running at low rates of rotation. In applying some of the results of this work to the design

of modem tanker screws, significant

improvements can be achieved by the

adoption of large diameter screws of up

to 30ft (914m) in diameter running at

low speeds (N about 75 revolutions per minute) and driven via geared turbines in

lieu of moderate diameter screws of

about 24ft in diameter running at moderate speeds (N about 110

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tions per minute) and directly driven by diesel engines. In the articles" 2 the

subject is discussed in general terms and

some comparisons are made between

alternative propulsion installations. In

the ardcle3

a procedure is given for

estimating the effects on screw efficiency due to varying the screw diameter and

the operating rate of rotation. In the

paper' data are given from which basic propulsion factors can be estimated and

the combined effects of aperture size

and screw diameter on propulsion factor values can be studied.

The object of the present article is to combine the data given in the publica-tions3. 4 with other screw design data in

the form of procedures

for making

screw performance estimates, propulsion estimates and screw design calculations for screws of different diameters operating at varying rates of rotation in propelling

hulls of varying aperture size, and to

apply the results obtained in making the propulsion estimates and design calcu-lations for a large tanker screw.

2. Screw performance estimates

varying rate of rotation

A marine screw can be designed to absorb a given delivered horsepower

DHP at a given speed of advanceVA to

run at a stipulated rate of rotation N, or at a rate of rotation selected from a

stipulated range of values. If the rate of

rotation is specified then the diameter

can be chosen to correspond to the

advance coefficient at which the screw

efficiency is

the maximum i.e.

the

optimum diameter can be chosen

providing that its value is not greater

that determined by aperture size (single

screws) or related to minimum

hull-tip clearance (twin screws). Alterna-tively, if the screw diameter is specified, then the rate of rotation can be selected from within the stipulated range so that

its value corresponds to the advance

coefficient at which the screw efficiency is the maximum, i.e. the optimum rate of

rotation can be chosen. A convenient method for selecting either the screw

diameter or the rate of rotation is by the use of the B-8 charts, and methods for

doing this are discussed in the articles

(References 7 and 3, respectively).

If the rate of rotation is fixed, the

power coefficientByand speed coefficient 8 can be computed using equations 13

and 15 defined in the

article' and

repeated below ND 8

=

GDHP B

=

V24 5V A

where N is the rate of rotation in

revolutions per minute

D is the screw diameter in

feet

$R is the relative flow factor

defined by equation 9 DHP is the delivered

horse-power in British units

is the specific gravity of

the fluid in which the

screw operates (average

value for sea water s =

1-026)

VA is the speed of advance of the screw in 'mots which

can be related to

the

speed of the hull Vs using the wake fraction wr defined by

VA = (1 wr) Vs

If the screw diameter is fixed and the rate of rotation can be selected, values of8/andB,,1can be computed using the basic value of the rate of rotation 111,

and this enables corresponding values

of 8 and B, to be derived by applying a factor k defined by

N = kINT,

8

=

B,, = kB,

In applying this procedure, a range of values of rate of rotation is chosen, and a set of values of 8 and 132, are

derived. This enables corresponding

pairs of values of pitch ratio p and screw efficiency n o to be read from the B2,-8 chart. Values of propulsive efficiency 71, are derived from no and the results

plotted on a base of rate of rotation as

shown in Fig. 3. The propulsive

effi-ciency can be derived either from the

screw efficiency behind the hull 71B by

applying a hull factor ef., or from the screw efficiency in open water no by

applying an overall hull factor The

hull factors as defined by equation 12

of the article are repeated below. no

eH nR = eo0

and

ep

= G

eff

where eR, is the relative flow factor

defined by nB = eR

3. Propulsion estimatesvariation

in screw diameter

For small changes in screw diameter

HULL SCREWSERIES DIA. D (ft) TRIAL SPEED Vs (knots) PROPULSION FACTORS Wake

Fraction FactorHull RelativeFlow Factor Overall Hull Factor WT eyi eit. ep (I) (2) (3) (4) 1 A 23.25 17 0-43 1.42 1-02 1-45 2 B 3000 17 0-38 1.23 I .05 1-34 2 C 3000 174 0-38 1-28 1-05 1-34

SCREW PERFORMANCE DATA Rate of Screw Propulsion Available

Rotation Efficiency Efficiency Effective

h/power NF (r.p.m.) no np EHP, (5) (6) (7) (8) I A 23.25 17 110 0.480 0.695 15,300 2 B 3000 17 75 0-575 0-770 16,950 2 C 30-00 174 75 0.583 0-780 17,150 RATE OF

ROTATION Power Speed Pitch Screw Propulsive

(r.p.m.) Coeff. Coeff. Ratio Efficiency Efficiency

k NF N Bp 8 PT no ilp (I) (2) (3) (4) (5) (5) (6) 0-65 65 63-7 25-2 178 0-975 0.580 0-777 0-70 70 68-6 27-1 191-5 0-865 0-583 0-780 0-75 75 73.5 29-0 205 0-780 0-583 0-780 0-80 80 78-4 31-0 219 0-715 0-580 0-777 0-85 85 83-3 32-9 232 0-655 0-573 0-768 0-90 90 89-2 34-8 246 0-600 0-555 0-744 1-00 100 98-0 38-7 273-5 0-545 0-536 0-718

I) Ref. 4, Fig. 7 (5) Basic value (1) Set of values cover ng required_range (4) (equation 5) 5= kk

(2) (equation 10) &E -=

I t

(6).1 Values from Fig. 3 (NF=75 to 100 r.p m.) (5) Values from Bp 8 chart

value of t from Ref. 4, Fig. 8.

:corresponding to basic

(7) j value of NF _{(2)Wake scale effect N=0.98 NF (6) (equation 7) np= fpnc,}(Ref. 8)

(3) Ref. 4, Fig. 12. (8) (equation 18) EHP=np DI-IP (Ref. 6, Section 4-9)

4) (equation 8) np=en.e. DHP= 22,000. (3) (equation 6) Bp= kBp,

TABLE I.HULLS I AND 2PROPULSION FACTORS AND BASIC TABLE 2.PROPULSION ESTIMATES FOR 30 FEET DIAMETER

SCREW PERFORMANCE DATA SCREWS (GROUP C)

BASIC DATA: N.S.M.B. B-4-55 SERIES

V5= 17-5 knots VA= 10-75 knots D=30 feet

ciency due to departure from standard geometric features are assessed using the data given in the publications; 6 Finally, the correction factors are applied to the respective basic parameters enabling the screw particulars to be determined.

6. Comparison of results

The results of the calculations given in Table 4 show that there are significant differences between the performance of

moderate diameter screws running at moderate rates of rotation and that of large diameter screws running at low

rates of rotation. They also show that in designing modern tanker screws, signifi-cant improvements can be achieved by the adoption of large diameter screws of up to 30ft (914m) in diameter running at low speeds (NI, 75 revolutions per minute) and driven via geared turbines in

lieu of moderate diameter screws of

about 24ft (7.32m) in diameter running at moderate speeds (NF=110 revolutions

per minute) and directly driven by

diesel engines, as summarised below.

For the vessel powered by a diesel engine and propelled by a moderate diameter screw running at a moderate

rate of rotation, (Screw 1, D = 23-25ft

(7.09m), Np = 110 revolutions per

minute), the trial speed was Vs = 17

knots and the screw efficiency and pro-pulsive efficiency were no = 0.485 and

772, = 0.703, respectively.

For the

vessel powered by a turbine installation and propelled by a large diameter screw running at a low rate of rotation (Screw 5, D = 29.5ft (8.99m), NF = 75

revolu-tions per minute), the trial speed was

Vs = 17i knots and the screw efficiency and propulsive efficiency wereno= 0.575 andn,= 0.77, respectively.

Thus, the adoption of a low speed

turbine installation in lieu of a moderate

speed diesel engine enabled a larger

diameter screw to be fitted, and the

corresponding improvements in

per-formance were: an increase in trial speed

of 3 per cent., an increase in screw

efficiency of 18 per cent., and a related increase in propulsive efficiency of 10 per cent.

References

TODD, F. H. Some possibilities for im-proving the propulsive efficiency of large

tankers. London, Marine Engineer and

Naval Architect. Sept., 1965.

MOTT, I. K. Practical application of the large slow speed propeller, London,

Marine Engineer and Naval Architect,

Nov., 1965.

O'BRIEN, T. P. Design of tug

pro-pellers part 3optimum screw azameter and rate of rotation. London, Ship and Boat Builder International, 1966, 19.

PARKER, M. N. The B.S.R.A.

metho-dical seriesan overall presentation. Propulsion factors. Trans. Royal Inst.

Naval Arch. 1965, 108.

O'BRIEN, T.P. Comparative perfor-mance of four, five and six blade screws

for large tankers. London, Shipbuilding and Shipping Record, 1966.

O'BRIEN, T. P. The design of marine

screw propellers, London, Hutchinson Scientific and Technical Press, July, 1962.

O'BRIEN, T. P. Design of tug pro-pellers. London, Ship and Boat Builder International, April, 1965, 18, 22. WRIGHT, B. D. W. The N.S.M.B. standard series propeller data and their

application. British Ship Research Association, T.M. 213, June, 1965.

VAN MANEN, J.D. A review of research activities at the Netherlands Ship Model Basin, Rotterdam, Holland, International Shipbuilding Progress, Nov., 1963. 10, 111.

given by the relation

f =

EHPT

EHP1

where EHP, is the available effective

horsepower for Group 1

EHPTis the effective horsepower corresponding to trial

con-ditions and the effective

horsepower, and delivered horsepower are linked by the propulsive efficiency

defined by

EHP =

7h,DHP

The results are plotted on a base of

speed of hull V. together with the values of available effective horsepower for

Groups B and C (points B and C) as

shown in Fig. 4. The point of

inter-section of the line B-C with the curve of corrected values of effective

horse-power determines the speed at which a 30ft (9.14m) diameter screw of Group C

would propel the vessel on trial when running at 75 revolutions per minute.

Thus enabling the speed of hull corres-ponding to Screw 5 to be estimated.

Screw 5design calculations

In making the design calculations given in Table 3; first, values of

pro-pulsion factors and basic coefficients

are obtained from Tables 1 and 2 and

the cavitation number a and power

coefficient Bu are evaluated. Next, the

speed coefficient a is read from the

optimum diameter and blade area chart (Fig. 1), the screw diameter is evaluated

and the cavitation number cr4.8 is

derived fron 04; this enables basic

values of blade area ratio aE, pitch ratio

T and screw efficiency 110 to be read

from this chart. The blade thickness

ratio T is determined using the Taylor

strength criterion", and the correction factors to pitch ratio and screw

effi-k0.7

2 0.6

derived from a A using equation 14.

Finally, the blade area ratio aE is obtained from the contour chart, and the

corres-ponding values of pitch ratio p and

screw efficiency no (by interpolation for change in expanded area if required) are read from the chart.

The concluding stages of the design calculations comprise the determination

of blade thickness and the application

of correction factors for departure from

basic blade thickness and from basic

blade section shape. The blade

thick-nesses can be based on blade stress

estimates made using the Taylor strength criterion following the procedure given

in the book6 or in the article, and the

-0

NUM I GROUP A V1. 17 KNOTS D. 23.25FT - HULL2 GROUP B 556I 'KNOTS 6.300 F7

NULL 3 GROW c veITJIANoTs 0.30.0 FT

z 0.11 0 ^

,1,11T,ILI

I 70 60 90 ice 120 RATE OF ROTATION N, (pp.) HULLS I AND 2 PROPULSION ESTIMATES

20,000 19,000 14,003 17,000 I 6,000 '15,000 14,000 00 16 15 400 1.006 E HP, 5, 300 A 0.2325 N, . 110 rpm 0.300 6,- 75 rpm

AVAILABLE EFFECTIVE. HORSEPOWER E HP,'

EFFECTIVE HORSEPOWER

TRIAL CONDITION EHP,

I I 1 I I

161,/ 17 17%2 113 le l'a

SPEED OF HULL V (K NO T S)

'FIG.4. HULL PERFORMANCE DATA & 'SPEED ESTIMATES

correction factors for departure from

standard geometric features

can be

determined using the data also given in the book6 and in the article.

5. Worked examples

It is required to prepare the

propul-sion estimates for

three. groups of

screws each of constant diameter over a

range of rate of rotation.. The results of these estimates are to be applied in making the design calculations for a

tanker screw (Screw 5) of large diameter

designed to operate at a low rate of

rotation.

The vessel for which this

screw is required will be similar to the

one for which the design calculations

for 4-, 5- and 6-bladed screws (Screws

1 to 4) are given in an article recently

published6. The first vessel (Hull 1)

was powered by a diesel engine the

delivered horsepower was 22,000 DHP and Screw diameter and rate of rotation were D = 23.25ft (7-09m) and N = 110 revolutions per minute, respectively.

The Second vessel (Hull 2) is to be

poweied by a turbine installation, the

delivered horsepower will be the same as for the first vessel, and the screw will

run at low rate of rotation (N about

75 revolutions per minute) enabling_ a

large diiimeter screw (D about 30ft

9.14m) to be selected.

The propulsion estimates for the first

group of screws (Group A) are to be made for Hull 1, the screw diameter will be 23.25ft (7-09m) and the trial

speed will be 17 knots. Those for the

Groups B and C are both to be made for

Hull 2, the screw diaMeter will be

30ft (9.14m) and the trial speeds will be 17 and 1711- knots, respectively. The values of the propulsion factors are to be derived from those of Hull 1 using the

data given in the paper4 referred to in

Section 3, above. The screw performance

estimates are to be made using the

revised B,-8 charts8 and applying the

procedure described in Section 2, above.

The screw design

calculations for Screw 5 are to be made using the optimum

diameter and blade area charts6 and applying the procedure discussed in

Section 4, above. The blade thicknesses

are to be determined using the Taylor strength criterion6, and the correction

factors for departure from standard geometric features are to be estimated

using the data given in the

publica-done. 6 referred to in Section 4, above.

Design data

HullSingle-screw tanker; length 830ft (253m), breadth 125ft (38.1m), draft level 45ft (13.7m), block

coefficient 0.8. Speed of hull V, (knots)

16-i 17 1Th 18 Effective horsepower EHPr

14,000 15,400 16,900 18,700

(trial condition)

Trial speed not less than 17 knots. .

Enginesteam turbine; delivered

horse-power

22,000 DHP, rate

of rotation to be selected from

within range NF 75 to 100

revolutions per minute.

Stern detailsStreamlined rudder, shaft

immersion I = 26.5 ft (8.08m) StipulationsMaximum screw diameter

30ft (914m), number of blades

four. Screw material, nickel aluminium bronze.

Design condition-22,000 DHP, value

of rate of rotation to be selected to give optimum performance. Trial speed not less than 17 knots.

Propulsion factors (Hull 1)Wake

frac-tion wr -= 0.43, relative

flow

factor eft =1.02, hull factor eri= 1-42.

(Hull 2)to be estimated using

the data given in the paper4.

g

c/

Propulsion estimates

In deriving the propulsion factors, the

values oi which are summarised in

Table 1; first, the values of wake fraction

wr, hull factor G and relative flow

factor eR for Group A (speed of Hull

V, = 17 knots, screw diameter D =

23-25ft-7:09m) were obtained from the data for Hull 1 and Screw 1 given in the

article. Next, three sets of values, the

first corresponding to Group A, the

second and third

corresponding to

Groups B (V,=-17 knots, D=30 feet) and Group C (V8=17i knots, D

=30ft-9.14m), respectively, were obtained using

the charts given in the paper4. This

enabled a series of ratios to be derived

each of which was applied to the res-pective parameter of the basic group thus enabling corresponding values of

propulsion factors for Groups B and C to be derived.

In making the three sets of propulsion

estimates one of which is

given in

Table 2, first, the basic values of power coefficient, B51 and speed coefficient 8/ are evaluated. Next, a series of values of power coefficient B, and speed coeffi-cient 8 are derived covering a range of rate of rotation. This enables a series of

corresponding values of pitch ratio p

and screw efficiency no to be obtained

from the B5-8 charts given in the

memorandums. Finally, a set of values of propulsive efficiency n, are derived

from no, and the results plotted on a

base of rate of rotation NF as shown in Figure 3. Values of screw efficiency and propulsive efficiency and derived values of available effective horsepower EHP corresponding to each appropriate rate of rotation are summarised in Table 1. In making the speed estimates; first, the values of available effective

horse-power EHP are corrected by applying

correlation factor f the value of which is 04₆₀

and for no variation in hull form the

values of propulsion factors can be

assumed constant. However, large

changes in screw diameter need to be

associated with variations in hull form,

aperture shape and shaft height which

result in variations in propulsion factors,

typical values of which are given in

Table 1. An accurate evaluation of these variations

would require model

ex-periments using different diameter screws associated with modifications in

hull form, but these

effects can be estimated using data given in a recent

paper by Parker4. This gave the results of propulsion experiments for a series of

hulls the geometric characteristics of

which covered variations in block coeffi-cient, position of longitudinal centre

of buoyancy, breadth-draft ratio,

length-displacement ratio and aperture size,

the three model screws of different

diameter.

The results were given in the form of a series of charts from which values of

wake fraction

wr, thrust

deduction

fraction t and relative rotative efficiency

77, for a desired hull form and screw

combination could be estimated. The

wake fraction tor is that defined by

equation 3 above and the " relative

rotative efficiency" and relative flow

factor (equation 9) are identical. The hull factor can be evaluated using the wake

fraction wr and the thrust deduction

fraction t and applying the equation (10) eH =

1 -t

1 -W T

4. Screw design calculations

There are a number of design methods

TABLE 3.-SCREW 5-DESIGN CALCULATIONS

BASIC DATA: Optimum dia. and blade area chart.

Troost 13-4 series single screws Fig. I.

DESIGN CONDITION: Delivered horsepower 22,000 DHP. Rate of rotation Np= 75 r.p.m. Trial speed Vs= 174 knots. No. of blades, 4. Shaft immersion 1=26.5 feet.

PROPULSION FACTORS AND BASIC COEFFICIENTS:coT=0.38 6=1.34 (Tables 1 and 2)

no= 0583 77p = 0.70 VA= 10-75 knots B=290

(1) (equation 15) (1) -e)=2084-1- 62.4s1= 3780 (2) (equation 16) 0A2 (;6%-v eA). -1075 (3) (equation 12) Bu= Bp Vv.= 22.15

BASIC GEOMETRIC DATA:

Values from the chart (Fig. 1)

(equation I) D= 4

(equation 14) D ad.s

L (Fo-e)

available ranging from simple charts to detailed calculation methods. A

con-venient form of chart which has proved useful in making screw design calcula-tions is

the optimum Diameter and

Blade Area Chart given in Section 3.12 of the book') two of which are reproduced in Figs. 1 and 2.

Each chart comprises two parts: a

graph of speed coefficient 8, pitch ratio p and screw efficiency no on a base of

square root of power coefficient /B;

a chart of contours of cavitation number

o-A.8 on co-ordinates of VB u and

ex-panded area aE.

The speed coefficient has already been defined (equation

1) and the power

coefficient is defined by

N THP

Bu =TA-i

sVA

where N is

the rate

of rotation in

revolutions per minute D is the screw diameter in feet

VA

is the speed of advance in

knots

THP is the thrust horsepower

applied by the screw

is the specific gravity of the

fluid in which the screw operates (an average value for sea water is s = 1-026)

If desired, the power coefficient Bu can be derived from the power

coeffi-cient Bu using the relation

Bu = 14-070

where B,, is the power coefficient as defined by equation 2

no is the screw efficiency in

uniform flow in open water The cavitation number is defined by

Basic values corrected for

de-parture from standard geometric

features. (Ref. 3 and 6.)

Blade thickness based on Taylor strength criterion.

(equation 7) np = fpno

(13)a48 -

2(p8-e)_{p V A'}

where (p.8-e) is the static pressure

measured at the x = 0-8 radius fraction of the screw when at minimum immer-sion.

vA

is the speed of advance in feet

per second

p

is the mass density of the fluid

in which the screw operates (for

fresh water p = 1938,- for sea

water p = 1.988)

If desired, the cavitation number

a A.8 can be derived from the basic form crA by applying the relation

ri -25sD1

I_ (130-e)

I

cr4

where (p, -e) is the static pressure

measured at the screw axis. and its value is given by

(p0-e) = 2084 + 62-4 sI

where I is the depth if immersion of the screw axis in feet.

The basic form of cavitation number is defined by

2(p, -e) =

(p0

-e)

crA =

(14) a4.8=

The procedure for designing screws

using the optimum diameter and blade area charts is as follows:

First, the cavitation number crA is

evaluated using equation 16 and the

power coefficient Bu is either directly determined using equation 11 or derived from B,, using equation 12. Next, the

speed coefficient 8 is read from the

chart, and this enables the screw

dia-meter to be evaluated using equation (1) and the cavitation number <74.8 to be

TABLE 4.-SCREWS 1 TO 5-GEOMETRIC FEATURES AND

PERFORMANCE DATA Delivered Horsepower 22,000 DHP HULL SCREW No. of BLADES RATE of ROTA-TION TRIAL SPEED GEOMETRIC FEATURES Doia' (feet) Blade Area Ratio

--SE Pitch Ratio Thick Ratio B Ny. (r.p.m.) v. (knots) pT T I 1 I I 2 I 2 3 4 5 4 5 6 6 4 110 110 110 100 75 17 17 17 17 174 23-25 23.25 23.25 23.25 29.50 0.625 0.700 0.780 0.800 0.530 0-770 0.750 0.740 0.840 01320 0.055 0-052 0.050 0.047 0.049 PERFORMANCE DATA Screw Effi-ciency Propul-sive Effi-ciency %increase above Screw 1 no np no 1 1 I I 2 1 2 3 4 5 4 5 6 6 4 110 110 110 100 75 17 17 17 17 174 0.485 0.485 0.475 0.490 0.575 0.703 0-703 0.689 0-710 0.770

-0 -14 4 18

-

_o -1i 10 VBIT E D 1 aii e 1 PT ILE no 4.71 (4) (5) (6) I (4) (4) (4) 201 29.5 8.6 1 0.81 0.53 0.575 SCREW PARTICULARS No.of Blades Dia. D (feet) Blade Area Ratio Pitch

Ratio ThicknessRatio Screw Efficiency Propulsive Efficiency SE p1 T n 0 np (5) (4) (7) (8) (7) (9) 4 29.5 0.53 0.82 0.049 0.575 0.77 v4 2-76sV 2