MED D ELAN D EN FRAN
STATE NS SKEPPSPRQVNINGSANSTALT
(PUBLICATIONS OF TilE SWEDISH STATE SHIPBUILDING EXPERIMENTAL TANK)Nr74
GÖTEBORG
1975SINGLE- AND TWIN-SCREW
PROPULSION OF TANKERS
AND BULK CARRIERS
BY
AKE WILLIAMS
Paper presented at the
First Ship Technology and Research (STAR)
Symposium
Di8tributed by:
Liber Distribution
P.O. Box
S-16210 VÄLLINGBY Sweden
PRINTED IN SWEDEN BY ELANDERS BOKTRYCRERI ARTIEBOLAG
S tJMMARY
In this paper a comparative analysis is carried out regarding re-quired shaft horse power for large tankers and bulk carriers. In the first instance single- and twin-screw propulsion are compared generally,, but efforts are also made to judge some alternatives of twin-screw propulsion. The study is limited to large and full
ships, VLCCS and ULCCS, i e ship lengths above 300 metres (dead-weights above about 200 000 tons). The hull block coefficient is mostly higher than 0.80. The analysis is based on about 400 self propulsion test series carried out for such ships at the Swedish State Shipbuilding Experimental Tank (SSPA) . This material is supposed to be the most extensive on one hand outside Japan. All data for the present analysis are derived from the SSPA hull form data bank, which now comprises nearly 5 000 hull forms for merchant ships. During the investigation special interest has been devoted to the twin-skeg (or twin-gondola) form, as this afterbody has lately been very much discussed as a realistic alternative to the conventional twin-screw hull form. While the difference in total required shaft horse power is rather small between single-screw and conventional twin-screw propulsion, there are clear indications that gains in SHP of 5 per cent and more can be achieved by use of the twin-screw twin-skeg principle.
1. INTRODUCTION
The share of multiple-screw propelled merchant ships has varied over the years, which can be studied in the shipbuilding statis-tics. For a certain type and size of ship twin-screw propulsion can be predominant for some years, after which a period of single-screw propulsion will follow or vice versa.
In recent years the large container ships of 20 000-30 000 TDW have appeared as single-screw as well as twin-screw ships. Also a number of triple-screw container vessels have been built lately. In general, the total amount of shaft horse power has settled the number of propellers. In the boundary cases, however, factors as propulsive efficiency, risk of propeller cavitation and vthra-tions, etc have been determining. Sometimes the reasons for a certain number of propellers have not been visible to an out-sider. A special engine room layout (maintenance, safety, etc) or a certain engine type or manufacturer owner's experience, costs,
etc) may have been preferred.
The development of VLCCS and ULCCS as regards single-screw and twin-screw propulsion (triple-screw never mentioned) is easier to follow. Simply expressed: One propeller has been specified as soon as this has been possible. If a one-engine unit (steam or diesel), able to meet the demand, has been available on the mar-ket, this has been chosen with few exceptions. Thus, over the years, it has been a race between ship sizes and engine ratings and, also, a corresponding competition between diesel engine and steam turbine manufacturers to offer the largest units.
Similarly to the container ships, there are cases (but fewer in number) where single- as well as twin-screw propulsion cari be chosen. Sometimes it is a question of steam or diesel.
In practice the choice is topical mostly for ULCCS and some VLCCS. Therefore the present investigation concerns only ships larger than about 200 000 TDW (ship lengths above 300 metres) Thus, here only a small part of the indicated area in Fig 1, which
represents the range of deadweight and engine SHP for merchant ships in general, is treated.
Fig i
Extent of DWT and SHP for merchant ships. The present investi-gation includes only VLCCS and ULCCS (marked area) DW TONS lo LARGE TANKERS AND BULK CARRIERS MEDIUM-SIZED CARGO LINERS
\
N
i ENGINE SHP COASTAL AND SHORT-TRADE SHIPS ULCCS VLCCS CONTAINER AND RO-ROIt must be noted that the following discussion of single-screw and twin-screw propulsion is limited to required engine shaft horse power (SHP) under certain specified conditions for the ship,
such as hull main dimensions, displacement and propeller particu-lars. No account is taken to such important factors as propeller cavitation and propeller-induced hull vibrations. However, par-allel to this work other extensive investigations for large ships have been carried out at SSPA in the last-mentioned fields [1],
[21. The aim is to attain a more complete picture of the propul-sion problems of a large and full ship and, if possible, to solve them.
2. SOME RESULTS FROM SYSTEMATICAL TESTS
As mentioned before, the present discussion includes only large tankers and bulk carriers. However, it may be interesting, for a moment, to see what a comparison between single- and twin-screw
propulsion for ships of different sizes looks like. From other systematical investigations at SSPA in this field [3], [4], the following has been extracted:
Exemplified ships:
Ordinary resistance and propulsion tests have been carried out, extrapolation by ITTC 1957 friction line, trial allowances accor-ding to SSPA's present standard procedure.
The twin-screw ships were model tested also with different boss angles and direction of propeller rotation. For each of the ship types the SSPA standard hull forms were used. The results are presented in Figs 2-4.
Considering the best combination of boss angle and direction of propeller rotation the results indicate the following differences
from single-screw, approximately expressed:
20 000
r'.
SHP ON TRIAL 15 000 4o
10000 200 40°BOSS ANGLE o< ABT 115 RPM AT 20 KNOTS
20 KNOTS
o
O
rwIN-SCREW OUTW TURN PRIP TWIN-SCREW INW TURN PRO - SINGLE SCREW60°
Fig 2
13 000 TDW cargo liner, single- and twin-screw
propuls ion
13 000 TOW cargo liner abt 20 knots 70 000 TDW tanker abt 17 knots 300 000 TDW tanker abt 16 knots
70 000 TDW tanker. Single- and twin-screw propulsion SHP ON TRIAL 25 000 20 000 50 000 SHP ON TR IAL 40000 15000 I 350 ABT 105 RPM AT 17 KNOTS I
O---TWIN-SCREW OUN TURN PR
Fig 3 - -- TWIN-SCREW INW TURN PRO
SINGLE SCREW
ABT 70 RPM AT 17 KNOTS'.,=:
-L 550 BOSS ANGLE o 18 KNOTS 17 KNOTS 17 KNOTS 16 KNOTS! 75 P Fig 4 300 000 TDW tanker. Single- and twin-screw propulsion30000
O
TWIN-SCREW OUTW TURN PRIP-
- TWIN-SCREW INW TURN PROSINGLE SCREW
30 50° 700
13 000 TDW cargo liner:
Twin-screw requires 10-15% higher SHP 70 000 TDW tanker:
Twin-screw requires O-2% higher Sl-IP 300 000 TDW tanker:
Twin-screw requires 2-5% lower SHP
Later results from other individuals in the 300 000 tdw-series indicated more or less agreement between single- and twin-screw
propulsion
-A first conclusion of this part of the investigation is that twin-screw propulsion is better (or less unfavourable) the larger the ships are. To this there may be two explanations:
The size of the propeller is usually relatively smaller for a larger ship. Hence, the shape of the afterbody is less propeller-adapted, which means that the afterbody hull forms for single- and twin-screw propulsion will be more alike. The resistance will be more the same and the afterbody will give the propellers in both cases less different flow resulting in more similar quasi-propulsive coefficients.
For an increasing size (and fullness) of single-screw ships the propeller load in question, expressed as the number of advance, gives as a result a rapidly decreasing propeller efficiency in
the behind condition. This drop in propeller efficiency cannot be balanced by the relatively high hull efficiency for this
type of ship.
It must be kept in mind, that in this case, when the problems con-cern flow around full ship forms, the model test results must be analysed with great caution. The different propulsive coefficients are affected by scale effects, some of them to a high degree, which have been the subject for some general investigations at SSPA [5]. It seems, however, that the comparison of required SHP between single- and twin-screw propulsion is not severely influ-enced. In any case, the tendencies reported above and in the following can be considered reliable.
3. ANALYSIS OF COLLECTED PROPULSIVE DATA FOR LARGE SHIPS
Available systematical data concerning comparison between single-and twin-screw drive are not sufficient as a basis for a decision whether or not the twin-screw alternative is the most favourable for propulsion of large ships. Any serial tests with systemati-cally varied standard hull forms have the drawback of suffering also from systematical errors. For example: At the edges of the investigated field the tested hull forms may keep the propulsive quality but become unpractical to realize. Such limitations of the serial tests will often be revealed years afterwards when shipowners and shipyards start to project ships of such size, hull form and fullness.
For the above reason it is most valuable to include in this in-vestigation also the results from all unsystematical tests carried out by customers for their different alternatives of large ship projects. Another reason to rely upon these tests is that the collected material represents a volume much larger than the syste-matical one. In all some 400 series of propulsion tests for ships
(built, to be built or projected) longer than 300 metres have been performed at SSPA up to now.
The results from all propulsion tests carried out at SSPA are collected, analysed and stored in a computer-operated data bank. The concentrated data from each test series include direct or codified figures for, among other things, customer, ship type, hull main particulars, hull form, propeller arrangement, kind of investigations, resistance and propulsion. The two last-mentioned quantities are expressed as actual effective horse power (EHP) and propeller shaft horse power (SHP) in relation to a corresponding ship with SSPA standard (single-screw) hull form and propeller.
We are now going to make a frequency test of the collected mate-rïal with regard to propulsive quality. Thus, the division into classes is to be made with regard to actual required SHP versus standard. In short, this relative figure tells us how 'good the topical ship is regarding propulsion. Consequently, a figure higher than 1 means higher required engine output than a standard single-screw ship meeting the same hydrostatical and hydrodynami-cal demands, i e the same hull main dimensions, displacement,
SSPA SHIP PROPULSION REGISTER
SPECIFICATION OF DEMANDED DATA FROM REGISTER
8 SSPA HLE CUS T OME R SHIP TYPE SHIP MODEL NO LENGTH LPP M DISPLACEMENT M3 BLOCK COEFFICIENT LENGTH/DISPL 1/3 BREADTH/DRAFT
SPECIAL HULL FORM NO OF PROPELLERS SPECIAL PROPELLER ARR SPECIAL INVESTIGATION SHIP SPEED KNOTS RESISTANCE VS STANDARD SHAFT HP VS STANDARD STANDARD SHIP SERIES
00 00 000 00 3 00-5 0 0 00 01 01 01 01 i 00 00 00 00 .98-1.01 T2
propeller RPM and propeller blade area.
An example of the use of the data bank for a topical frequency test with regard to propulsion is shown in Table 1. How many tankers (code 300), single-sòrew, belong to the class engine SHP vs standard 0.98-1.01? Part of the answer is given as Table 2, where also other demanded data appear on the printed sheet from
the computer. Standard ship series T2 means SSPA standard hull forms for VLCCS and ULCCS. All model test results for tankers and bulk carriers in the range 200 000 to 400 000 TDW (single- and twin-screw) are judged against this standard series, which con-tains single-screw forms only.
The result of the frequency test for single-screw ships is shown in Fig 5. A small number of systematical data are included, namely the individual hull forms of some serial tests. This will contribute, to a certain degree, to concentrate the material around engine SHP vs standard = 1, as some of this material forms the basis of standard series T2. Likewise, all figures engine SHP vs standard less than 0.82 and above 1.18 (rather few in number) are excluded. In these ranges the analyses have often proved to be unsafe, since the limits of the standard series with regard to
fullness and main dimensions ratios have been too far exceeded.
Using the same procedure for the twin-screw ships we obtain re-sults such as Fig 6. The distribution of engine SHP vs standard is here more concentrated around 1, but the standard deviation seems to be larger. Table 1 Example of input to register. 00 indicates no information needed. 01 indicates demanded information without restrictions
Example of computer output from the data bank. Demanded data as per Table 1 NO OF SHIPS 50 25 Fig 5 o f I I Distribution of engine 0.82 0.90 0.98 1.06 1.14 SHP vs standard. 0.86
O.4
1.02 1 10 1.18 Single-screw ships L > 300 m ENGINE SHP VS STANDARD 9SSPA SHIP PROPULSION REGISTER DATA OUTPUT FROM REGISTER
SHIP TYPE 310 310 310 310 310
LENGTH LPP M 338.0 310.0 320.0 320.0 310.9
BLOCK COEFFICIENT .830 .847 .849 .837 .839
LENGTH/DISPL 1/3 4.891 5.083 5.156 5.506 5.006
BREADTH/DRAFT 3.006 2.567 2.216 2.616 2.357
SPECIAL HULL FORM 120 120 110 120 0
NO OF PROPELLERS 1 1 1 1 1
SHAFT HP VS STANDARD 98 .98 98 .99 98
STANDARD SHIP SERIES T2 T2 12 12 12
SHIP TYPE 310 310 310 310 310
LENGTH LPP M 310.9 310.9 329.2 329.2 321.7
BLOCK COEFFICIENT .8444 .809 .833 .816 .791
LENGTH/DISPL 1/3 4.995 5.945 5.110 5.540 5.967
BREADTH/DRAFT 2.357 4.058 2.591 3.455 4.851
SPECIAL HULL FORM 110 110 110 120 120
NO OF PROPELLERS 1 1 1 1 1
SHAFT HP VS STANDARD 1.00 1.01 1.00 .99 .99
lo
NO OF SHIPS 50 25 O 0.82 0 90 0.98 1.06 1 14 0.860.94
1.02 1.10 1.18 ENGINE SHP VS STANDARD NO OF SHIPS 25 o0.82
0.90
0.98 1.06 1.140.86
0.94
1.02 1.10 1.18ENGINE SI-tP VS STANDARD
Fig 6 Distribution of engine SI-IP vs standard.
Twin-screw ships L > 300 in Fig 7 Distribution of engine SI-IP vs standard. Twin-screw ships L > 300 in. Conven-tional bossingsNO OF SHIPS 25 Fig 8 Distribution of engine O I I SHP vs standard. Twin- 0.82 0.90 0 98 1.06 1.14 screw ships L > 300 m. 0.86
0.94
1.02 1.10 1.18 Twin-skeg afterbody ENGINE SHP VS STANDARDThe presented distribution comprises all twin-screw ships, the bulk of which have the conventional appearance, i e V-shaped afterbody, open or closed bossings with the propellers located on each side of a centerline deadwood, and one or two rudders. The result of the analysis for this class of twin-screw ships is pre-sented in Fig 7.
Other types of twin-screw arrangements are contra-rotating pro-pellers, overlapping propellers and the twin-skeg arrangement.
(For a closer description of the twin-skeg afterbody, see Appen-dix II.) Most of the results available at SSPA for unconventional twin-screw propulsion are from tests with the last-mentioned type. Therefore a special analysis has been performed for these ships, Fig 8. The tests are rather few in number and the approach to the different hull forms has sometimes been vague. In spite of this, the assembled material as it appears from Fig 8 is highly interes-ting. It can be assumed from this part of the investigation that the twin-skeg arrangement is more favourable than the conventio-nal twin-screw design regarding propulsion and also better than the normal single-screw arrangement.
In this connection we must be aware that this discussion only concerns one sector of marine propulsion, namely required engine shaft horse power. Propeller cavitation and risk of vibrations are not considered. However, for the twin-screw twin-skeg ships a number of cavitation tunnel tests have been performed, which have also included measurement of propeller-induced hull pressure variations. None of these tests have given indications that the
12
0.9 1 1.1
ENGINE SHP VS STANDARD
Table 3
Engine SEP vs standard. Mean value and standard deviation
Fig 9
Engine SEP vs standard. Mean value and standard
deviation
twin-skeg is, in this respect, inferior to other designs. (See
also Appendix II.)
4. CONCLUSIONS
A summing-up of the results from the investigation is presented as Table 3 and Fig 9.
The mean value of engine SEP vs standard is 1.02 for all single-screw ships and 1.01 for all twin-single-screw ships. The difference is hardly significant. This can also be said of the mean value 1.03 for the twin-screw with conventional bossings. The standard de-viations are too large. However, the last-mentioned figure may also be looked upon as a correction of the tendency obtained from the systematical series, see Fig 4 and accompanying text.
The twin-screw twin-skeg ships have an SEP-relation to standard as low as 0.96, which means that by adopting this propulsion principle a reduction in required SHP of well over 5 per cent can be expected when keeping ship main dimensions, hull fullness and propeller RPM constant. PROPULSIVE ARR SHP VS MEAN VALUE STAND STAND DEV S SINGLE-SCREW TWIN-SCREW CONV BOSS TWIN-SKEG 1.02 1.01 1.03 0.96 0.07 0.09 0.08 0.09 SINGLE- . CREW S S
TWIN-SC.
L SHIPS TWIN-SC:1 NV BOSS TWIN-SC 1*L...i"J11*I
Complementary conclusions are given in the following Appendix I, where, as an application of the above results, two recently
de-livered large ships are compared. Appendix II deals with the flow around the afterbodies of a single-screw ship and a twin-skeg ship respectively.
APPENDIX I
COMPARISON OF THE PROPULSIVE QUALITIES OF TWO LARGE ORE-OIL CARRIERS WITH SINGLE- AND TWIN-SCREW PROPULSION RESPECTIVELY
The investigation accounted for in the foregoing indicated cer-tain advantages of the twin-screw twin-skeg arrangement. It must be kept in mind, however, that such a characterization is valid only from a relatively limited point of view, namely the hydro-dynamicist's. Other scales of measure may be used in order to judge a ship's propulsive qualities, both technical and purely economic. The following comparison is an attempt to see the prob-lem using also a couple of other simple criteria.
The ships to be compared are two large ore/àil carriers recently delivered to Swedish owners. Ship A is the "Svealand" of 278 800 TDW, owned by Broströms, Göteborg. Ship B is the "Tarfala" of 269 980 TDW, owned by Gränges, Stockholm. "Svealand" has single-screw propulsion and a conventional afterbody hull form, while
"Tarfala" represents the twin-screw twin-skeg alternative. Hull and main engine particulars are to be found in Table 4. The after-body lines of "Svealand" have been designed according to the same principles as SSPA standard single-screw U-form, Fig 10. The SSPA standard twin-screw twin-skeg hull form, Fig 11, may represent
"Tarfala's" afterbody lines.
Table 4
Ships A and B, hull and engine main particulars
14
SHIP A B
MAX DRAFT CORRESP DWT
DRAFT AT MODEL TESTS
CORRESP DWT
CORRESP DISPLACEMENT LENGTH BEN PERP
BEAM
FOREBODY HULL FORM
AFTERBODY HULL FORM PROPELLER ARRANGEMENT MACHINERY RATE OF REVS PROPELLER TYPE (MOULDED) TONS A TONS A M 1016 KG M 1016 KG M 3 M M SHP RPM 21.85 278800 21.67 275550 312200 321.6 54.55 BULB CONV 1-SCREW 41000 106 4-BL FP 21.99 269980 21.80 266215 304300 320.0 52.00 BULB 2-SKEG 2-SCREW 2x20000 114 4-BL CP
Fig 10
SSPA standard single-screw tankers, U-formed afterbody
- 1/2
n"
t%_liii
15I
3VIII
AU o j,2 o Fig 1130 000 16 KNOTS ENGINE SHP 50 000 16 ENGINE SHP 45 000 40000 35 000 45 000 40000 35 000 16.5 17 Fig 12
Enqine SEP for ships A and B, direct comparison
Fig 13
Engine SHP for stan-dard (single-screw) ships corresponding to ships A and B
The analysis is carried out on the basis of the model test re-suits obtained at drafts indicated in Table 4. This limitation of the analysis is not severe, as both ships seem to have fulfilled the performance predictions to the same degree.
A direct comparison of the SHP-curves, Fig 12, gives the
im-pression that the two ships have the same relative power consump-tion, as the difference in required SEP is of the same order as the difference in displacement (or deadweight). However, it will be shown below that this is by far a too hasty conclusion. Ship B has a significantly higher block coefficient, while the slender-ness (L/V'/) is about the same for both ships. This means a con-siderably higher required SUP for the "standard ship"
correspon-_.-::: SFIIP Al SHIP B STANDARD SHIP CORRESP TO SHP B \CORRESP STANDARD SHIP TO SHIP A 16 KNOTS 16.5 17
Fig 14 Ships A and B. Comparison of SEP with regard to deviations in DWT and displacement ENGINE SHP 50 000 40000 30 000 20 000 16.5 KNOTS DIFF= 1.5% DIFF= 5% 10 000 SHIP SHIP A Al A2 B MODIFICATIONS OF SHIP A TOWARDS SHIP B
ding to ship B, Fig 13. The 'standard ship' against which every studied ship is compared, has the same main dimensions, hull full-ness, propeller RPM and propeller blade area as the topical ship. Further, the standard ship has, in this case, hull form and pro-peller arrangement according to 'SSPA tanker series T2" (V-formed
afterbody, single-screw, U-formed forebody without bulbous bow)
The continued comparison, at a constant speed of 16.5 knots, is carried out as a stepwise modification of ship A towards ship B. The modifications are supposed to be applied to ship A at the de-sign stage without strictly practical considerations. Thus, the first modification of ship A, a reduction of about 8 000 cubic metres of displacement to agree with ship B, is carried out by reducing the hull fullness, keeping length, breadth and draft constant. According to tanker series T2 this means a 5.5% reduc-tion of required SEP for ship A, see Fig 14. Modificareduc-tion to the saine deadweight means a slight further displacement reduction due to the superior light ship weight figures for ship A. As is seen in Fig 14, the final result is that ship A requires 5% less SHP at the same deadweight provided original hull main dimensions.
18 ENGINE SHP 50 000 40000 16.5 KNOTS )IFF= 10% SHIP SHIP A
]j3
A4, B MODIFICATIONS OF SHIP A TOWARDS SHIP B Fig 15 Ships A and B. Comparison of SHP with regard to deviations in dis-placement, hull main dimensions and propeller RPMThe above kind of conclusion may be satisfactory to the ship-owners as the comparison is based on transport capacity. However, there are some objections from the ship hydrodynamicist he will want to know how much of the total difference in required SHP
that is entirely due to hull form respectively propeller arrange-ment. In other words, the influence from diverging main dimen-sions and propeller RPM must be estimated and separated from the judgement. It is essential to have the hull form (and the pro-peller arrangement) judged independently so that a certain hull shape can be applied to new hull main dimensions. Again using tanker series T2, ship A will suffer an 11% increase of required SHP if given the same main dimensions as ship B, see Fig 15. On the basis of constant displacement this means that ship A re-quires 7% higher SHP provided the same hull main dimensions as ship B.
Ship A takes advantage also of a lower propeller RPM than that of ship B. The complete comparison includes an additional correction due to deviation in propeller RPM. If this correction is applied, which sometimes can be discussed from engine installation points
DIFF= 7% 30 000
o
o
I-L) L)o
L) L) 20 000I-
Ó L)Z
L) I-4o
<
<
-Jo
o
V) (f) V) I-I J -J 10 000-Fig 16
Engine SHP vs standard, positions of ships A and B in the assembled material. (See also Fig 9)
0.9 1 1.1
ENGINE SHP VS STANDARD
of view, the above difference will rise from 7 to 10%.
The result of the present analysis is in good agreement with the earlier general conclusions regarding the propulsive properties of twin-screw twin-skeg ships in relation to conventional
single-screw ships (Fig 9) . It is also clear that the two exemplified vessels hold favourable positions within their respective classes regarding required SliP, Fig 16.
The following concluding remarks can be made:
Both ships are well designed from the hydrodynamical point of view considering the fixed choice of propeller arrangement for both of the ships.
Ship A is specially favoured by the good choice of main dimen-sions (and the relation between them) , a moderate hull f ull-ness and a low propeller RPM.
Ship B is specially favoured by the twin-screw drive in combi-nation with a good alternative of twin-skeg afterbody.
By use of the twin-screw twin-skeg principle, relatively lower required SHP can be gained than for conventional single-screw propulsion. This is confirmed by the assembled statistics at SSPA as well as by the topical comparison. However, experience is up to now limited to large and full ships.
19 SINGLE CREW bHIP A fl4INSC1.
TWINSCEW
PL SHIPS LV BOSSTWIN CEW
INSKEGAPPENDIX II
STUDIES OF THE FLOW AROUND THE AFTERBODY OF A TWIN-SKEG SHIP MODEL
The twin-skeg afterbody hull form is not new. The principle has long been used, mostly for small, but also for medium-sized twin-screw ships. The main purpose of the skegs has been to carry the shafts to the wing propellers. The advantage of such an arrange-ment is among other things a better support of the shaft bearings and the possibility of inboard inspection of the shaft line. For these ships the port and starboard skegs are relatively thin as the shafts only are to be accomodated.
However, some hydrodynamical disadvantages have been coupled to the twin-skeg arrangement, which have possibly for many years limited the use of this alternative to relatively small and low-powered ships. In Harold Saunders' book "Hydrodynamics in Ship Design" [6] he states:
"...On ships with offset or wing propellers carried by offset skegs it sometimes happens that the inward component of flow in way of the skeg is much greater than can be accomodated by the
tapered planform of the skeg or by its convergence inward and aft. As a result, separation occurs abreast the after end of the skeg, on its inner side. It is accompanied by a large irregular vortex with its axis generally vertical. A part of the active portion of this vortex may lie within the swept volume of the propeller. In this case the almost disastrous effect of the propeller in setting up local hull vibrations can well be imagined. Even though the vertical-core vortex may lie outside the swept volume, water is found to rush across this volume, from the outside toward the in-side of the skeg, almost at right angles to the aperture and par-allel to the plane of the propeller disc. Tremendous variations in the forces and moments developed at the upper and lower blade positions generate equally tremendous variations in the ship..
The idea of using the twin-skeg principle also for large ships, VLCCS and ULCCS, was taken up in 1968 by the European diesel engine manufacturer Sulzer. The background was that the length of the machinery space became almost unacceptable when two slow-running diesel engines were installed in a large ship with
ventional afterbody and bossings. The competing steam turbine plant could offer, firstly, the demanded total SHP on one shaft and, secondly, a reasonable engine room length with regard to cargo cubic capacity strived for. If the aft portion of the diesel engines could be pushed backwards into two skegs (of rela-tively high fullness) built outside the normal afterbody hull, the forward engine room bulkhead could be located as favourably as for the steam turbine alternative. Sulzer carried out some preliminary model tests with such a hull form [7]. The results indicated, however, bad figures for resistance and propulsion. Closer studies of the flow around the afterbody were apparently not carried out.
A few years ago new hydrodynamical investigations for large twin-skeg (or better: twin-gondola) ships were started, this time with the aim to really build ships of this design. Extensive model tests were carried out, among others, at the Swedish model basin, SSPA, which resulted in the construction of a series of large twin-skeg ore-oil carriers, among these the 'Tarfala" mentioned in Appendix I.
Further studies of twin-screw twin-skeg ships have been performed at SSPA for the account of Uddevallavarvet AB, Sweden. These ships are of ULCC-size with wide beam and restricted draft, char-acteristics which seem to justify the adoption of the twin-skeg principle. Some results are presented below from flow studies in SSPA's large cavitation tunnel. For comparison results are also given from corresponding te5ts with a single-screw 300 000 TDW tanker with U-shaped afterbody. The last-mentioned investigation is included in a research programme on propulsion of large tan-kers for the Swedish Ship Research Foundation.
Let us start by looking at the flow around the single-screw after-body, when being tested in the cavitation tunnel. The visualiza-tion of the flow is attained a) by air ejecvisualiza-tion from pressure holes at the stern, see photograph, Fig 17 and b) by application of "tufts" on the afterbody hull surface, Figs 18 and 19. The two manners of presentation complement each other and we can observe the following: Due to the afterbody hull form and fullness a bilge vortex is created by the bilge flow and the downward flow ahead of the propeller. This vortex is not necessarily unhealthy
22 Fig 17 Result of air ejection test. Single-screw ship Fig 18 Flow observations with tufts. Single-screw ship. Without propeller
for the propeller operation. The problem is to predict from model tests the strength, location and stability of the vortex in full scale.
The area of separated flow above the propeller, revealed by the tufts, is in this case not large enough to influence the propel-ler performance. It is interesting to note, however, that the
Fig 19 Flow observations with tufts. Single-screw ship. With propeller Fig 20 Result of air ejection test. Twin-skeg ship 1/2
propeller action somewhat increases the separation area. Similar tests with other single-screw ships, not so well designed near the stern frame, have indicated separated areas in this region of such large extent that so called stern fins above the propel-ler were necessary.
The air-bubble flow test for the twin-skeg ship, Fig 20, gives indications of separated flow between the rudders, which was also confirmed by the tests with tufts, view from below, Figs 21
24 R ZO'IE.
-CL Fig 21 Flow observations with tufts. Twin-skeg ship. Without propeller. View from below
Fig 22
Flow observations with tufts.
Twin-skeg ship. With propeller. View from below
and 22. Also in this case the area of separation is increased by the working propeller. At the air-bubble test it could further be seen that some of the air paths outside the bilges were twisted, indicating weak bilge vortices. A sign of a possible bilge vor-tex can also be observed at the tests with tufts, side view, Figs 23 and 24. Inside the skegs no twisting of air paths was observed.
Despite the above observations, the overall impression of the topical twin-skeg design at tests in cavitation tunnel is favour-able. In addition to the flow test results, reported above, this judgement is based on hull pressure fluctuation measurements and cavitation erosion tests. All tests were carried out with a large ship model mounted in SSPA cavitation tunnel No 2.
Fig 23
Flow observations with tufts.
Twin-skeg ship. Without propeller. View from SB side
Fig 24
Flow observations with tufts.
Twin-skeg ship. With propeller. View from SB side
W LWL 10 LWL
ACKNOWLEDGMENT
The author wishes to thank Dr Hans Edstrand, Director General of SSPA, for having been given the opportunity to carry out this work and for the attention he has paid to it. Thanks are also due to Mr Hasse Olofsson and Mr Kjell Svensson for their technical assistance and, further, to Mrs Barbara Karsberg, who has revised the English and typed the manuscript and to Mrs Gunilla Bokedal, who has prepared the figures.
REFE RENCES
Johnsson, C-A and Söntvedt, T: Propeller Excitation and Response of 230 000 TDW tankers. SSPA Publication No 70, Det norske Ventas Report No 79, 1972
Dyne, G: A Study of the Scale Effect on Wake, Propeller Cavitation and Vibratory Pressure at Hull of Two Tanker Models. Trans SNAME, 1974
Williams, A: Comparative Investigations of Single- and Twin-Screw Propulsion of Merchant Ships. SSPA General Report No 21, 1967 (in Swedish)
Williams, A: Triple-Screw Propulsion of Container Ships. Shipping World and Shipbuilder, October 1971
Lindgren, H: Ship-Model Correlation Method Based on Theore-tical Considerations. Trans 13th ITTC Performance Committee Report, Berlin/Hamburg, September 1972
Saunders, H: Hydrodynamics in Ship Design. Vol I, p 495-496. SNAME 1957
Marine Engineer and Naval Architect, September 1968, p 355-358: Sulzer Proposals for High-Powered Motorships
CONTENTS
Summary i
i. Introduction 2
Some results from systematical tests 3 Analysis of collected propulsive data for large ships 7
Conclusions 12 Appendix I 14 Appendix II 20 Acknowledgment 26 References 26 27
