NEDERLANDS SCHEEPSSTUDIEGENTRUM TNO
NETHERLANDS SHIP RESEARCH CENTRE TNO
ENG INEERING DEPARTMENT LEEGHWATERSTRAAT 5, DELFT
CONTRAROTATING PROPELLER PROPULSION
PART I
TERN GEAR, LINE S,HAFT SYSTEM AND ENGINE ROOM ARRANGEMENT FOR
DRIVING CONTRAROTATING
OPELLERS(VOORTSTU WING DOOR TEGENGESTELD DRAAIENDE SCHROEVEN)
DEELI
(ONDERSTEUNING, D ORVOERING EN CONSTRUCTIE VAN HET ASSYSTEEM VOOR EEN TEGENGESTELD DRAAIENDE VOORTSTUWINGSIÑST)L ATIE MET
BIJBEHORENDE INRICHTING VAN DE MACHINEKAMER)
by
IR. A. DE VOS
(Manager Machinery Division N.y. Koninklijke Máatschappij ,,De Scheide", Vlissingen)
IR. J. VAN HAASmRT IR. C. KAPSENBERG
IR. J. D. RUYS
IR C. C. SÑÈs
PROF. IR. W. VINI IR. A. DE Vos
IR. A. DE Moor (ex officio)
Gedurende de laatste jaren valt een opmerkelijke toeñame te
bespeuren in sneiheid, grootte en bet per schroefas geinstalleerde veimogen van het zeegaande koopvaardijschip.
Als gevoig van de 1iirmee gepaard gáande economische druk is het voor de reder van belang bet schip als vervoerseenheid te othnaliseren. Hierdoor is het gehele voortstuwihgssysteem, dat opgebouwd gedacht kan worden uit de voortstuwer, machine-installatie, verlflogeflstraflsnhissie en in zekere zin de scheeps-vorm, in het niiddëlpunt van de belangstellihg komen te staan. Voor wat betreft de machine-installatie mag worden gesteld dat een toenómende tendens van bet nuttig effect valt waar te nemen bu toenemende vermogens.
De toepasbaarheid van hoog belaste conventiònele
voort-stuwers wordt begrensd door fysische verschijnselen als cavitatie, trillingen als gevoig van fluctuerende askrachten en momenten en op de scheepshuid gemnduceerde hydrodynamische drukken. Mädel experimented onderzoek heeft reeds aangetoond dat een niet onaanzienlijke toename van het nuttig effect mag worden verwacht b toepassing van op twee co-axiale assen bevestigde schroeven, die in tegengestelde zin draaien. Aan deze winst mag niet voorbijgegaan worden, tenzij de pogingen orn een dergelijk systeern te realiseren zullen leiden tot compromissen elders in het systeem of tot onaanvaardbare constructieve complicaties
Orndat nst een te verwachten rendementswinst een std tegengesteld draaiende schroeven mogelijk andere veelbelovende
eigeñschappen zou kunnen bezitten, werd besloten tot de uit-voering van een researchprogramma, dat beoogde de diverse
aspecten van dit voortstuwingssysteem te bestuderen
Voor wat betreft bet werktuigbouwkundige aspect werd een compleet ontwerp van de ondersteuñing en doorvoering van de assdi, bevestiging van dó schroeven, askoppeling, c.d. gemaakt. Dit ontwerp werd uitgewerkt omdat gemeend werd dat de
be-staande concepties op verschillende punten konden wOrden
verbeterd.
Ht door de N.y. Koninkhjke Maatschappij Dc Scheide" in
Vlissingen vervaardigde ontwerp worth in het onderhavige rap-port gepresenteerd.
Dd classificatiebureaus lloyd's Register of Shipping, Bureau Ventas en Det Norske Ventas werden uitgenodigd het ontwerp te bestu4eren.
Na voltooiing van dit algemene ontwerp weid bet nuttig
ge-oordeeld voor een dubbelschroef contäinerschip, dat voor
re-kenitig van de Koninklijke Nediloyd N.y. in aanbouw was, een
schaduwontwerp te maken met het dod na te gaan op welke
puntei aanpassing van zowel het voortstuwingssysteem als bet schip noodzakelijk was.
Het schaduwontwerp is tevens in dit rapport opgenomen.
Daarnaast werd voor dit schip een uitvoerig vergeijkend
onderzoek gedaan naar de hydrodynamische aspecten. Voon de resultáten van dit onderzoek zu verwezen naar rapport no. 168 S.
Het; zal duideijk zijn dat, alvorens tot de toepassing van een
tegengesteld draaiende voortstuwingsinstallatiè besloten zal
worded een hoge mate van zekórheid dient te bestaan inzakede betrouwbaarheid van het systeem.
Ter bepaling van het dynaniisch gedrag van bet assysteem werd
een bórókeriingsmethode opgesteld die in rapport no. 167 M
wordt epresenteerd.
Het 'onderhavige rapport besluit met enkele aaiibevelingeñ
voor bet - verrichten van ñader onderzoek naar het gedrag van enige, de bedrijfszekerheid bepalende, a.specten.
NEDERLANDS SCHEÈPSSTÚDIECENTRLJM TNO
Over the past few years merchant ships have displayed a marked increase in speed, size and installed power per shaft. Also because of today's economic pressures,.shipowners are forced to optimize
the ship as a tÑnsportation unit. Thereforê interest focuses on
the propulsion system that can be thought to consist of propeller, engine installation, power transmission and, in some way, the
very shape of the ship. As for priie movers, a rising trend in
efficiency is observed as their capacity levels incréase.
.The application of highly loaded conventional propellers is
limited by such physil prOperties as cavitation, vibrations due
to fluctuating forces and moments, and hall-induced
hydro-dynamic pressures.
Model experiments have indicated the magnitude of gains in efficiency when applying a pair of propellers attached to coaxial
shafts that rotate in opposite direction. These gains cannot be
overlooked unless efforts to realize them call for drastic com-promises elsewhere in the system or lead to unacceptable con-structional complications.
Because, apart from gains in propulsive efficiency, contra-rotating propulsion promises other attractive fóatures it was
decidedto carry out a research programme on vanòus aspects of such a propulsión system.
Accordingly, a complete design of stem gear, propeller mountings, shalt couplings etc. was made because it was held that existing designs could be improved on some vital points.
This general design as made by the N.y. Koninklijke
Maat-schappij "De ScheIde" is presented in the present report. The classification societies Lloyd's Register of Shipping, Bureau
Ventas and Det Norske Ventas were invited to comment on the
design.
After completion it was thought worthwhile to investigate whether this design could be madó suitable for a twin-screw
container ship that was under construction when the pertinent iiivestigatiöns were carried out. This study resulted in a "shadow-design"; it indicates where the contrarotating propulsion system
as well as the ship's stem had to be adapted. The
"shadow-design" is also presented in this report.
The hydrodynarnic and propulsive properties of a contra-rotating propulsion system were eatensively investigated and compared with the twin'.scre' arrangement of this particular
ship. These rósults are published in Report No. 168 S.
Before a contrarotating propulsion system can be brought
into actùal servicó, a high degree of reliability must be present concerning the operatio±ial behaviour of the system.
For determining the dynamic properties of the shafting, a
calculation method was dóvised; it is presented in Report No.
167 M.
The underlying report, finally, gives recommendations for
future work to be carried out concerning the behaviour of some major components.
CONTENTS
3 Shadow-design for a containership for the Koninklljke Nediloyd N.Y.
li
3.1 Introdùctión - 11
3.2 Data of the twin-screw ship
il
3.3 Data of the CRP-ship - 12
3.4 Coñstructional details 12
3.4.1 Support 12
3.4.2 Shafting 12
3.4.3 Engine room plant 12
3.4.4 the gearbox - 12 4 Future work 13 5 Acknowledgements 13 References . - 13 2 page Summary 5 IntrodUction 5 GeneraI design 6 2.1 Data - . 6 2.2 Propellers 7 2.2A Support . 7 2.2.2 Bearings 7 2.2.3 Fixing
..
.i
2.24 Seals 8 12.5 Lubrication 8 2.3 Shafting 9 2.3.1 Support . . 92.3.2. The contra-rotating bearing for the innershaft - 9
2.13 Lubricating oil system 9
2 3 4 Seal between shafting and gearbox lubricating oil 10
2.4 The gearbox
..,
.. 102.5 Loads and Stresses . . . 10
CONTRAROTATING PROPELLER PROPULSION
PART I
STERN GEAR, LINE SHAFT SYSTEM AND ENGINE ROOM ARRANGEMENT FÖR ÒRIVING CONTRAROTATING PROPELLERS
by Ir. A. DE VOS
Summary
A design relating to stern gear, lihe shaft system, and the fitting of a pair of contrarotating propellers to their respective tail shafts is proposed.
The general design is further discussed in connection with its application to a fast container ship under construction at the time the work was carried out.
i
introductionDüring recent years a veritable revolution has been taking place in the transport of both raw materials nd fimshed products, and in this connection one thinks particularly of container traffic Economically speaking,
container traffic only yields optimal results When traisport is rapid and loading and discharge times
short.
In the transport system as a whôle; therefore,
shipment must be made in vessels of advanced esign equipped with very high capacity propulsion units to ensure compliance with the rapid turnround require ments of the containers.
Inview of the high investment and operating costs of vessels fitted out exclusively for carrying containers there is everything to be said for trying to find new or
adapted constructions which may help to reduce
expeiditure.
In this connection one may take it that the type of
propeller decided upon is
particularly important becaúse this partly determines the size and natUre of the prithmover selected. In view of the high power of the machinery currently being installed in container ships,a choice can be made in principle between the
fol-lowing systems: twin propeller unit overlapping propellers
C. triple propeller unit
d. contrarotatiñg propellers
Although contrarotating propellers have long been
employed fór propulsion (mainly for torpedoes), up to 1960 their use was restricted to low power units in tended for short operational periods.
Since experimental research has shown that
pro-pulsionby means of contrarotating propellets can lead
to not inconsiderable output gains compared with
results obtained with more conventional propellers, various propulsion units have been designed incor-porating two propellers tUrning in opposite directions.
T. W. Steele gives an account of one such design [11*
He makes reference to Ingvar Jung's welhknown
publication [2] which describes many different ways of achieving the ultimate effect of driving tandem pro-pellers in opposite directions.
T. W. Steele points out that contrarotating propellers are now in use in United States submarines and one experimental surface vessel. On the basis of experience gained with these ships Steele describes his design, a conspicuous feature of which is the use of a planetary second reduction gear. This may well be regarded as advanced fot mercantile marine purposes since the only experience with a gear like this, transmitting
con-siderable power, has been obtained on board the
"Tinimerman" of the U. S. Navy.
Steele further discusses what he has in mind for the arrangement of the coaxial propeller shafts. He rightly emphasizes the importance of proper alignment of the shafting in view of the big overhanging moment of the aft propeller. The bearing he describes between the two contrarotating shafts does not seem to have caused any difficulty during testing but it is not apparent from
the arrangement of the coaxial shafting how the
propeller shaft bearings can be inspected without com-pletely dismantling the inner and outerpropeller shafts. SKF couplings are used for coupling the inner shafts and are rendered accessible by splitting the outer shaft longitudinally over a sufficient length.
According to Laskey and Gruber [3], howevet, such a consttuction cannot be reliably executed. They state
* Numbers in parentheses designate References at the end of the report.
in their publication that this has been proved by full-size tests carried out by Ishikawajima-Harima Heavy IndustriesCo. Ltd. on a split hollow shaft 7' in length. Steele uses a conventional means of securing the pro-pellers to the conical ends of the shafts by propeller nuts. Laskey and Gruber press the propeller hubs on to the cones hydraulically, but re-introduce keys.
In both constructions, of course, the overhanging moment of the aft propeller is considerable, and no mention is made of any constructional measures being taken to counteract this.
Opinions still differ as to the design of the propeller shaft seals. It is hoped that lip seals, more or less of the type much used at present in conventional ships, can be modified for use with contrarotating propellers. As, however, the circumferential speed of the stainless steel sleeve on the propeller shaft is so much higher than is now customary, it is doubtful whether this is in fact possible.
In Stal-Laval's Technical Information Letter [4],
Hilding Hillander deals particularly with the shafting
fór contrarotating propellers. He describes a metal
bearing specially designed for the inner shaft that has been successfully tested at Kockums, in Malmø. The
drawings of the propeller shaft bearings are only
schematic, and it does not look as if the bearings can be easily inspected.
The propellers themselves are mounted on cones, 12 for the aft propeller and i : 20 for the forward propeller. A key is used fOr the aft propeller. No special measures are taken to counteract the bigoverhanging moment of the two propellers.
Sterling A. Fielding [5] suggests a solution for the problem of dismantling the shafting of contrarotating
propellers. By arranging for inboard withdrawal of
the inner and outer shafts a single rudder can be used. The use of special couplings avoids the necessity for horizontally-split sections in the outer shaft.
This brief summary of some publications relating to contrarotating propellers shows that a few new and attractive designs have indeed been proposed. Never-theless, sufficient effort has nOt always been made to arrive at practical constructional solutions. In particular there seems to be a need för further evaluation of the manner in which the weight of the propellers should be transmitted to the ship, the accessibility for inspection and maintenance of the bearings, and the way in which the bearings should be protected against the effect of seawater.
It seemed definitely worthwhile, therefore to try to find a new design incorporating constructional solutions for the above mentioned items.
On completion of the design, its possible
applic-tions were investigated by projecting it in that of the
container ship already under construction for the
Koninklijke Nedlloyd N.Y.
The general design and the shadow design for ths ship, both of which are presented in this report, were
executed by N.Y. Koninldijke Maatschappij "Ire
Schelde" (Royal Schelde), specialists being consulted on certain points.
The design was discussed with experts from tie
Classification Societies of Lloyd's Register of Shippiig, Bureau Yeritas, and Det Norske Ventas.
2 General design
The general design is based on that of a fast container ship driven by contrarotating propellers.
The Netherlands Ship Model Basin in Wageninen (NSMB) made a rough calculation of some dat of
vItal importance to the designer on the basis of/the
2 x 35,000 SHP selected. 2.1 Data
The main dimensions of the ship are:
= 550ft
B = 73ft
D = 30ft
A = 20,000 tons (metric) Horsepower installed: 2 x 35,000 thp Shaft revolutions: 140rpmDiameter forward propeller: 6,5,00 mm
Number of forward propeller blades: 4
Diameter aft propeller: 6,000 mm
Number of aft propeller blades: 5
Speed:
approx 2 kn.
It is estimated that when sailing ahead the thrus of the
aft propeller will be about 120 tons and that 'of the
forward propeller about 130 tons.
It is provisionally assumed that, given identical
rpm, the total power is equally distributed betwen the two propellers. W. H. Budd [61, however, poiits out
that "fine" ships show the highest output when the
capacity of the forward propeller is less than thdt of the
aft propeller at the same rpm. Further tests (will be necessary to determine the most suitable distribitionof
power and what the difference in the number
of revolutions will have to be to attain maximum efficien-cy. The differences will not be great, however, nd will have no influence upon the design of the constructions described below.2.2 Propellers 2.2.1 Süpport (fig. 1)
The use of increasingly heavy propellers means that in çonventional constructions the overhanging moment
of the propeller becomes bigger and bigger. To
overcome this difficulty, the stern tube is not bored parallel to the base line but at an angle of inclination corresponding as closely as possible to the elastic line of the shafting. The Classiflcatiòn Societies all have a còmputer programme at their disposal which enables them to make a rapid calculation of this angle and at the same time to determine the most favourable posi-tion for the bearings, both fore-and-aft and vertically. With contrarotating propellers in a tandem
arrange-ment, however, no such measures can be taken in
respect of the aft propeller.
Attempts have therefore been made to solve the
problem by eliminating the big overhanging moment altogether or at least reducing it considerably. As the figure shows, this has been achieved by fitting a forged steel supporting bush which has a very high moment of resistance to bending and which, because it is excep-tionally firmly fixed in the boss of the stern frame, has
become, as it were, part of the ship itself.*
The forward propeller is driven at the back by the outer shaft, and for bearing purposes is fitted with a
pressed-in steel bush lined with white metal. The
supporting bush has suitable longitudinal oil grooves. Clearly the weight of the propeller is borne on the spot an4 the outer shaft is therefore not subject to bending.
Foi this reason, discussions with the Classification
Societies have resulted in their agreeing to consider the outêr shaft as an intermediate shaft in this construction andto dispense with the fórmula governing the dia-metèr of the (conventional) propeller shaft in this case.
The same principle has been followed for the aft
proeller. Because of the big moment of resistance of the thickened outer shaft at that point, practically all
the inner shaft has to do is to transmit the torque.
Thus it may also be considered as an intermediate shaft. This, in fact, is just What has happened in the case of the design for the 2 x 35,000 shp propellers.
The outer shaft betweeen the two propellers is notj
requi-ed to transmit any torque and, when turning,
only a short length of it is subject to bending because of the weight of the aft propeller.
Thé diameter of the boss in the stern frame has had to be increased from 1,800 mm to 2,000 mm so that a large enough supporting bush could be fitted.
There i.s no reason to assume that this slight
modifi-cation will have any noticeably adverse effect on the resistance of the ship or on the propulsion efficiency.
2.2.2 Bearings (fig. 1)
As already stated, the forward propeller is fitted with a pressed-in steel bush.
This is also true of the aft propeller.
The latter, however, does not
revolve round a
stationary supporting bush
as does the forward
propeller but round the outer shaft which is turning in the opposite direction. Contrarotaton is therefore set up.
Theoretically speaking, a contrarotating bearing
yields zero load capacity. However, this only applies to perfectly circular bearings. Any bearing surface which has a centre of curvatute not coincident with its centre of rotation generates hydrodynamic forces.
Extensive tests have been carried out by O. Pinkus[7] with a 6" shaft supported by a contrarotating bearing. These tests showed that small, unavoIdable, defor-mations and dynamic loads ensure that contrarotating bearings of various shapes are capable of supporting a
load of at least 1,500 psi. Only the contrarotating
bearing that had no axial oil grooves failed.
Floating pads and a floating ring between shaft and bearing were also tested, likewise with success. The clearances used seem to be normal. On the basis of these tests, which are frequently mentioned in technical literature, it was decided to incorporate two floating rings in the design of the contrarotating bearing of the aft propeller. These rings are omnilaterally lined with white metal and have 4 axial oil grooves on both the inside and outside. The angle at the centre between the 8 oil grooves is 45°. If these grooves are made during the fiia1 machining, slight deformation of the circular ring occurs, which helps to produce squeezing of the oil film during contrarotation. The floating ring will turn slowly with the outer shaft. The number of revolutions depends on the oil pressure and on the bearing load. This has been demonstrated by Pinkus's tests.
If the number of floating ring revolutions is no more than 20% of the number of contrarotating revolutions
then - as stated in the literature - there is really no
contrarotation problem.
In the present design the floating ring is split in two. This means that in consequence of a slightly diffçrent oil pressure or a deviating load, for instance, the front ring can rotate with a slightly different number of revolu-tions from the back ring. In principle, more than two rings could be used.
2.2.3 Fixing (fig. 1) * Ston Manganese Marine Limited has meanwhile obtained a
securing the propellers to the two shafts (at the back, of course) since this would mean that the shafts could only be introduced from the rear. Moreover, inspection of the propeller bearings would only be possible after both shafts had been withdrawn backwards. This would involve considerable work. The solution has now been found in lengthening the back of the propeller hubs. In each of the elongated hubs a steel bushing has béen pressed in at a pre-calculated pressure. Propeller and
bushing are now secured to the propeller shaft by
means of SKF hydraulic couplings, for which 1: 30
conical bushes are used. At the point where the
forward propeller is fixed, the hollow outer shaft is reinforced by pressing in a bushing of high quality steel. The propeller shafts need never be withdrawn
because the shafts themselves are not supported by bearings and there are therefore no bearing surfaces to be inspected during the 4 or 5-yearly survey. This may well be regarded as one of the main advantages of this design over all the others proposed, since withdrawing
and reassembling a set of coaxial propeller shafts
involves a great deal of work and many extra days in dry dock.
Inspection of the propeller bearings is effected by loosening the propellers hydraulically according to the usual SKF principle and then removing them.
In due course, some auxiliary tools can be designed to supplement the standard SKF tools.
A white metal lined steel bush for the forward
propeller and a set of floating rings for the aft propeller can be carried as spares.
2.2.4 Seals (fig. i and 2)
In deciding on the seals, a primary consideraticn was that they had to be of split construction, and the high circumferential speed of 21.1 rn/sec. was also taken into account.
The type selected has alEeady been used on water turbines for shaft thicknesses of 400 to 1,800 mm and for the oil lubricated thrust blocks of Voith-Schneider propellers, which have shaft thicknesses of 1,380 and
1,520 mm.
Figure 2 shows the design construction and the
materials to be employed. The seal between the two propellers has seawater behind it and lúbricating oil in
froht. Each seal consists of 16 segments which are
pressed on to the sleeve and against the vertical sealing surfaces of the (split, and thus removable) casing by
springs around the circumference. The casing is provided with the necessary driving pins.
Although the aim is complete separation of seawater and oil, the seals themselves need not be absolutely tight. Sorne leakage is permissible since this drains off
through 4 radial drihings in the casing. There are 4
stainless steel pipes cast into the hub of the aft pro1 peller for draining the leakage mixture back into the ship. At the seal between the forward propeller and th? ship itself, the circumferential speeds are much Iower. The same type of seal has been used, however, but th. casing in which the seatiñg elements are housed is nov in two parts, each of which is also of split construction, of course. The outer seal keeps out the seawater, the inner the lubricating oil.
Leakage from both sides collects in the large spae between the two casings which also contains.the leakage mixture from the seals between the two propellers. Several holes of ample size in the stern frame allow the leakage to flow easily back into the ship where it is collected in a drain tank.
Lubricating oil can be recovered by centrifuging.
As far as is known, the proposed arrangement
provides the best insurance against penetration ofseawater into the lubrication systern
2.2.5 Lubrication (fig. 1)
The lubricating oil is fed in under pressure between the outer shaft and the supporting bush
The oil then flOws astern and part of it rapdly
reaches the bearing of the forward propeller. 'rhe supporting bush of this propeller is fitted with the axial oil grooves needed to ensure that the beajngsurface receives enough oil for lubrication and cooing After passing the bearing of the forward propeE1er, the oil flows into a space sealed off by one of the seals described in 22.4. That oil can then flow back thrugh some large holes in the stern frame to a lubricating oil drain tank in the ship. The rest of the oil flows thrugh some radial holes in the outer shaft to the space betveen the two shafts. These radial holes have been made in
thr L:..1.___ÁL1U.ae11eu parI. 01 L11_L .t..outer ShalL.
The oil that flows into the space between the two coaxial shafts divides into two streams. The one that flows astern attends to the lubrication and cooiihg of the contrarotating bearing
of the
aftproeller.
Emerging frOm this bearing, the oil flows into a pace which is protected from seawater by the contrarotating seal described in 2.2.4., This oil can now pass though the hub of thé forward propeller via 4 large stainless steel pipes that are cast into the propeller hub. I then enters the same space as the oil.comingfrom th bear-ing of the forward propeller and can flow off in the same way to the lubricating oil drain tank in the ship. The stream of oil that flows forward into thej space
between the shafts attends to the lubrication and
cooling of the contrarotating beariñgs in the sIafting as described in 2.3.3.
2.3 Shafting
2.3.1 Support (fig. 3)
The back of the outer shaft is suspended at the sup
porting bush in the steth frame by means of the
coupling of the propeller hub and the propeller hub
itself.
The front of the outer shaft is supported by bearings iñ the gearbox. In between these two supports, two conventional force lubricated tunnel blocks are placed. Moreover, one radial bearing of the
conven-tional thrust block is used as support for the outer
shaft.
whole shafting is so designed that the bearings
of the inner shaft do not coincide with those of the
outer shaft, as is mostly the case in other designs This wôuld have made inspection of the inner shaft bearings practically impossible.
In the present design, the contrarotating bearing of thè inner shaft is housed in three very short split sections of the outer shaft. These are the only sections that are
split and they are so short and of such rigid
con-strúction that no danger arises of the kind indicated by Laskey and Gruber [3]. SKF flange couplings with I ::30 conical bushes are used for coupling the sections of the iñner shaft
At two points where this was feasible, an SKF flange
coupling was connected to a forged flange so as to
reduce the length of the "bridge" sections of the outer shaft as much as possible. These "bridge" sections are not split horizontally. It are the big flanged plates which coniiect the "bridge" sections to the normal outer shaft by means of reamed coupling bolts that are split. This
construction, which is not unknown, can be very
relibly executed.
2.3.2 The contrarotating bearing for the inner shaft
(fig. 4)
The figure gives details of this contrarotating bearing and illustrates the very rigid construction of the split section of the outer shaft, which is only 420 mm in length.
Instead of floating rings as in the aft propeller,
floating pads are employed. According to Pinkus [7],
the rèsuits are just as good. Loose pads are indeed
essential here siñce inspection and replacethent would otherwise not be possible without dismantling much of the entire shafting.
The pads, lined inside and out with white metal, are loosely connected to ofle another by two rings, one at the frönt and one at the back. There are holes in the rings with pins fitting loosely through them. Each pad has one pin screwed in at the front and one at the back. The rings are split and are assembled With the joints at
an angle of 900 to each other to prevent the mechanism from falling apart when the split section of the outer
shaft is dismantled. A rethining ring, aÏso split of
course, holds the pads in line when they start turning slowly with the outer shaft.
Fig. illustrates a way of connecting the pads to one another and preventing possible skewing.
Double T=shaped links are used to make the pads into a chain. The chain is not closed since this does not seem necessary.
One may ask whether a wa cannot be found to
eliminate contrarotation entirely. One possibility
would be to drive the floating pads by means of an
epicyclic gear train. Fig. 6 shows how this could be
done. The pads still rotate with the outer shaft at
10.45 rpm, which is well under the value of 20% of 140 rpm so that there would be no question of real contrarotation.
The pads can be kept absolutely still by fixing two pinions to the driving pins of the pad in question so that one pinion meshes with the external teeth of the inner shaft and the other with the internal teeth of the outer shaft.*
There is a limit to the possibilities, however.
In the present design the ratios are such that the
pinions rotate either with the inner or With the outer shaft.
Proceeding on the basis of pifions with an even
number of teeth, if 4 pinions are used the number of teeth for the gear must be divisible by 4. This means 144 or 148 teeth for the internal teeth of the outer shaft, meshing with a second pinion with 10 or 14 teeth. The result is that the pinions turn with the inner shaft at
11.3 rpm, which is slightly less favourable than the solution shoWn in fig. 6. The pinions themselves rotate here at 3,752 rpm, which is very high. It is questionable,
therefore, whether this idea will ever be put into
practice. In view of the elastic behaviour of the
shafting, the length of the contrarotating bearing has been kept short (1 - d), which is regarded as favour-able.
2.3.3 Lubricating oil system (fig. 3)
A flexible connection into which the lubricating oil is pressure fed has been introduced between the aftermost bearing block and the aft peak bulkhead From there, the oil flows aft and lubricates the propeller bearings. Part of the oil returns through the space between the inner and the outer shaft as described iñ 2.2.5. This oil flows between the pads of the aftermost contrarotating
bearing and lubricates them at the same time. Two
* The Glacier Metal Co. Ltd. has meanwhile applied for a patent on this idea.
other contrarotating bearings are lubricated by the oil flowing forward in the space between the inner and the
outer shaft. Ultimately the oil reaches a large oil
discharge compartment and can flow off from there to a lubricating oil drain tank.
2.3.4 Seal between shafting and gearbox lubricating
oil (fig. 7)
The Ïubricating oil which lubricates the foremost con trarotating bearing and then flows to the lubricating oil dischargecompartment is sealed off from the splash oil spreading along the coaxial shafts from the gearbox. This is achieved by two sealing elements described in 2.2.4, also housed in a split casing.
There is no reason to expect oil pressure on either side of the seal, so leakage there, if any, can only be
negligible.
2.4 The gearbox (fig. 8)
Proceeding on the assumption that in this case the
shafting is driven by a steam turbine, double reduction
gearing of normal design is projected. Two extra
wheels are fitted in the casing to reverse the direction of rotation of the large forward wheel in respect of that of the large aft wheel.
The arrangemeni of the wheels is shown in fig. 9.
The thrust block for the inner shaft is built into the
gearbox. Quillshafts ensure even distribution of power. 2.5 Loads and stresses
One may ask whether the proposed construction might not lead to excessive loads and material stresses but, as far as is known, this is not the case.
As stated in 2.2.1, the Classification Societies have
agreed in principle to consider the shafts as
inter-mediate shafts and the design has been worked out on this basis, using steel with a tensile strength of 44-52
kg/mm2
Since a short length of the outer shaft is subject to
bending because of the weight of the àft propeller,
the Classification formula for determining the thick-ness of the propeller shaft has been applied and the moment of resistance is considerably greater than
required. The dimensions of the supporting bush are
such that if the total weight of the propellers and
shafting is applied to the end as concentrated load, the
maximum bending stress amounts to less than 200
kg/cm2. When the bush is perfectly clamped, deflection is no more than a few hundredths of a millimetre.
Thanks to the large diameter, the forward propeller bearing has a specific bearing load of not more than 7 kg/cm2 and that of the aft propeller bearing is even less.
Much attention has been paid to calculating the SKF coupling of the propellers.
Safety factor 3 required by Lloyd's Register in
respect Of maximum torque and maximum thrust hak been achieved, which allows stress in the outer shat equivalent to 60% of the 0.2% yield point. In this case, therefore, a stress of 0.60 x 22 = 13.2 kg/mm2.
The stresses in the couplings, like those in the shaft, have been calculated according to the energy criterión of Huber-Hencky-von Mises.
For thecouplingsand in th high quality steel bush the outer shaft, the permissible stress is equivalent to 80% of the 0.2% yield point of the material in question.
The coefficient of friction for bronze on steel hs
been taken as 0.17. This is in line with the
recommenda-tions of Lloyd's Register of Shipping and with tets
carried out by Lips N.Y. It will be possible to attain this coefficIent by making the contact surfaces betwdn propeller and sleeve absolutely "dry".
I
It is not yet known how the propeller couplings ,ill
behave after they have been dismantled several tinies. The safety factor of 3 that has been mentioned is vlid
up to 35°C. At 0°C the permissible stress is flot
exceeded. The material chosen for the propellers is a quality which has the 0.2% yield point at 32 kg/mm2 whereas the actual stress is 27 kg/mm2.
Neverthèless, the stresses in the propeller hubs inay be high at certain points. SKF has set up a compüter programme to investigate this further.
Muth attention will still have to be devoted to the lubrication of the contrarotating bearings. The oil has to travel a long distance between the coaxial cont1aro-tating shafts. All the heads and nuts of the coupling bolts will have to be screened to prevent the oil ¶rom foaming. Calculations will have to be made to dete9mne the most suitable oil pressure to overcome all resistances and how the oil is to be properly distributed. Clecks must be made to see that the temperature f the oil does not rise too much in the course of its travel.
Should this be so, then an extra feed point will have to be arranged, which will complicate what is at present an essentially simple lubricating system. The behafriour of the oil in the contrarotating bearings should certainly be investigated further. It is already known, námely, that the pads proposed can result in the oil flowing off
laterally so that a pressureless area can arise n the
centre of the pads.
Model tests by others have demonstrated that the
minimum pressure in the bearing over the entire
circumference of a pad must not fall below 0.5 k/cm2.
Further it appears possible that, when pahs are
used, vortices may form in the oil stream and obstruct the free flow of the oil.
i
behaviour of this shafting for contrarotating propellers,
particularly as the supporting points of inner and outer shaft do not cincide. A. W. van Beek 8] has already indiôated the direction this work should take. Tooth loàds for this essentially conventional gear transmission are not high
The helical teeth on the pinions and wheels are cut.
All the pinions and the primary wheels are then
hardened and ground, the secondary wheels ground onÏy. The pifions are of case hardening steel, BS type
En 36, and the wheels are' of Cr Ni Mo steel, BS
type En 28. In the proposed design the K factors are below the limits set by the Classification Societies.
2.6 Emergency measures (fig. 10)
It cannot be denied that the design under discussion còntäins several néW features which have not been süfliniently tested yet. This could mean that, for the time being, adoption of the proposed design involves a higher element of risk than a conventional installation,
If such a conventional installation consisted of a
doUble steam turbine plant for a twin screw ship, the "way back" - were the proposed installation with con-trarotating propellers to fail is practically impossible. The aft ship would have to be replaced by a new aft sectiòn for twin propellers, and the singlesteam turbine and gearbox would have to be replaced by a double set.
Befòre ordering a contrarotating installation (evefi
afterfurther studies and tests have been carried out) the shippwner will ask himself whether there is not a more
reasonable and economical solution should such a
"way back" - contrai-y to expectations - prove
necessary. To answer this question a design was made for rèplacing the two contrarotating propellers by a
siìigle one.
Fig. 10 illustrates this.
The new, and naturally much larger, propeller is slid on tO the supporting bush that is already there and, in
the manner described earlier, fitted with an SKF
coupling to a new shaft.
An alrea4y existing SKF coupling is fixed to the
front of this new shaft. The flange of this coupling is again connected to the aftermost, short, horizontally-split section of the existing outer shaft.
Inñer and outer shaft are now connected at that
point änd can function as 'a single shaft.
The existing coaxial shafting in the ship can be
retained in its entirety.'
The rpm of the new propeller remains 140.
The diameter of the propeller will, it i estimated, become
to 8 m. A propeller with a diameter of
7.20 m' which is still clear of the base line - has been drawn in provisionally.
The advantage of the higher efficienc' of the contra-rotating propellers is now lost of course.
It would perhaps have been a good idea to design a slightly larger propeller frame right at the start, with a view to clearance for a larger propeller. The unbroken line in fig. 10 represents this.
The gear casing itself need not be changed. What must be done is to replace the large fòrward wheel by a wheel of the same size as the large aft wheel. Two new secondary pinions are then necessary and the idler in the drive of the inner shaft must be remoied. This is clearly seen from fig. 9.
The pads in the thtust block in the gearbox' are
omitted.
The thrust block on the outer shaft should perhaps have been made heavier right at the beginning since this now 'has to absorb the entire thrust of 10,000 shp
If the following units are kept ready ashore - a new propeller and SKF coupling
- a new shaft
- a ñew large gear and two pinions
then, should conversion be necessary, it
can be
accomplithed 'in a relatively short time. Efficiency will still be higher than with a twin screw installation.
If inner and outer shafts are coupled together as described, the quillshafts in the gearing are of
ex-ceptional value.
3 Shadow-design for a container ship for the Koflinklijke Nediloyd N.y.
3.1 Introductiòn
On completion of the above design it was thought
sensible to compare the engine room arrangement of a
twin screw container ship with that of an identical
container ship equipped with contrarotating propellers. Since there is a twin screw container ship under
coñ-struction for Dutch account, it was decided that it
would be interesting to include this particular engine
room installation in the comparison, especially as
comparative studies are also being made in respect of the hydrodynamiç aspects of the ship.
3.2 Data of the twin screw ship
The main dimensions of the container ship are: Length between perpendiculars (La,,): 270 m
Breadth (B): 32.20 m
Draught (D): 10.00 m
Displacement (A): 52,291 tons (metric)
Horse power installed: 2 x 43,000 shp Shaft horse power (service): 2x38,600 shp
Speed: 27.5 kn
The rudder is so designed and fitted that the shaft
can also be introduced from the stern if the rudder ha not already been mounted. For the thrust block of th outer shaft, a centreline thrust block is now projected.
3.4.3 Engine room plant (fig 12)
In view of the reduction of power whiçh is possibe with the use of contrarotating propellers, two smaller
boilers are projected, namely two "ScheIde Combustion Engineering" boilers, each producing 125 tons of steaith per hour.
Thére is a steath turbine of the required power, the high and low pressure parts of whiöh driie the pinins
of a gearbox One large condenser is
projected. Otherwise, the installation is in principle unchanged. lo keep the (expensive) shafting as short as possible, the engine room has been placed two container lengths further aft and the gearbox as far aft as the shape of the ship permits. At the same time this represen1s a (small) gain in the number of containers that cani be carried. It does mean, however, that the connection between the turbine and the gearbox must be realized by fitting two fastrunning shafts.Although not customary, it ought to be possib! to execute this arrangement reliably. By ca1culating thç number of journal bearings, the critical whirl speds can be kept outside the speed area.
/
ShoUld the calculations reveal that a great many bearings are needed, this may be undesirable. ¡ The efficiency of a gearing is, after all, largely determined by
the losses sustained in the bearings of the rapidly
rotating pinions of the reduction gear. In that eveit, an alternative might be to split üp the gearbox into one
box for the first reduction and another for the second.
The two, long, fast-running shafts would thn be
replaced by 4 slower-running ones.
This version would certainly be more expensive and has not been taken any further here.
3.44 The gearbox (fig. 13)
The lower number of propeller revolutions means that The double helical teeth on the pinions and 'wheels are hobbed.,
The pinions are of Cr Ni Mo steel, BS type/En 25, the intermediate wheels are of similar material, BS type En 26, and the gearwheels are of carbon steól BS type
En 8. The K factors are below the Classification
limits.
In this version, too, the thrust blöck for inner shaft is still in the gearbox. Should the design be put into execution, the high thrust renders it advsable to a locked train gear must be chosen.
Speed of propellers: 134 rpm
Diameter of propellers: 6,300 mm Number of propeller blades: 5
As the data of the actual ship indicated above by *)
were not yet available at the time this report was
prepared for publication, data resulting from
propul-sion tests carried out by the NSMB are used. The
same extrapolation procedure has been used for both the twin-screw version. and the CRP version.
3.3 batäof the CRF ship
Based upon the results of the NSMB's propulsion tests
the following main dimensions were used for the
design:
Horse power installed: 2 x 38,000 slip Shaft horse power (service): 2 x 34,500 shp
Speed: 27.5 kn
Speed of propellers: 105 rpm
Diameter of forward propeller
(left-handed): 7,300 mm
Number of blades: 4
Diameter of aft propeller
(right-handed): 6,750 mth
Number of blades: 5
The body plan of the twin-screw ship has been adapted to that of a CRP ship.
3.4 Constructional details
The shadow design for the CRP container ship was
based entirely on the design described earlier for a
plant of 2 x 35,000 shp at 140 rpm. Since 105 rpm has
now been chosen however, the dimensions for the
plant of 2 x 38 000 shp are much increased Detailed calculations have not been made.
The shadow design has been obtained more or less by extrapolation. Because the stresses calculated for the first design were low, it may be assumed that
after further study, of course - this is a practical
proposition.
3.4.1 Support (fig. Il)
This is no different from that of fig. i as described in 2.2. The same applies to bearings, fitting, seals, and lubrication The circumferential speeds on the seals are slightly lower here.
14.2 Shafting
This, too, is based entirely on that of the first design as described in 2.3
make the inner shaft longer and to fit a separate thrust block in front of the gearbox.
Fig. 12 shows that there is sufficient space for it.
4 Future wOrk
I. It is desirable to obtainan insight into the total cost of a ship equipped with contrarotating propellers as described in the foregoing pages and to compare it with that of a twin screw ship of the samç tonnage and speed.
It is desirable to find out which friction coeffióients can be used as a basis for the calculations, given the couplings used in the design, and to what extent they change after the couplings have heen t ken apart several times. The latter applies not so much to the couplings in the shafting as to those on the propellers.
It is desirable to investigate thé way the súggested seals react to the high circumferential speeds that
occur, especially as regards wear and tear and
leakage as a function of the running hours.
Actual service conditions will have to be simulated
r
as realistically as possible.
The result of such an enquiry could also be useful for the coñventional stern tube seal for ships where increasing power also means increasing circum ferential speeds and where the limit now seems to be in sight.
It is desirable to obtain an insight into the. problem of the oil transmission through the space between the coaxial shafts and into that of the lubrication of the contrarotating bearings.
lests with plastic models can help to find ways of ¿ounteracting instability and preventing vortices
fl orn developing.
5 Acknowledgements
The Netherlands Ship Research
Centre TNO
gratefully acknowledges the co-operation and assis-tance reveived from the N.y. Koniñkiijke
Maatschap-pij "De Scheide", Lloyd's
Register of Shipping,Bureau Ventas, Det Norske Ventas, the Glacier Metal Company Ltd., Gustav Húhn A.B. and S.K.F. Neder-land N.Y.. for the work reported herein.
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J.. M.. J. Journée 1972.
159 5 DASH computer program for Dynamic Analysis of Ship Hulls: S. Hylarides, 1971.
160 M Marine refrigeration engineering Part VII. Predicting the con-trol properties of water valves in marine refrigerating installations.
A. H. van derTak, 1971.
161 S Full-scale measurements of stresses in the bulkcarrier m.v. 'Ossendrecht'. ist Progress Report: General introduction and
information. Verification of the gaussian law for stress-response to waves. F. X. P. Soejadi, 1971.
162 S Motions and mooring forces of twin-hulled ship configurations. M. F. van Sluijs, 1971.
163 S Performance and propeller load fluctuations of a ship in waves. M. F. van Sluijs, 1972.
166 M Contrarotating propeller propulsión, Part I, Stern gear, line
shaft system and engine room arrangement for driving contra-rotating propellers. A. de Vos, 1972.
169 S Analysis of the resistance increase in waves of a fast cargo ship J. Gerritsma and W. Beukelman, 1972.
182 S Stress-analysis of a plane bulkhead subjected to a lateral load.
P. Meijers, 1972.
Communications
1 i C Investigations into the use of some shipbottom paints, based on scarcely sapoóifiable vehicles (Dutch). A. M. van Londen and
P. de Wolfe 1964.
12 C The pre-treatment of ship plates : The treatment of welded jointr
prior to painting (Dutch). A. M. van Londen and W. Muider,
1965.
13 C CorrOsion, ship bottom paints (Dutch). H. C. Ekama, I 966.
14 S Human reaction to shipboard vibration, a study of existing
literature (Dutch). W. ten Cate, 1966.
15 M Refrigerated containerized transport (Dutch). J. A. Knobbout,
1967.
16 S Measures to prevent sound and vibration annoyance aboard a seagoing passenger and carferry, fitted out with dieselengines
(Dutch). J. Biziten, J. H. Janssen, H. F. Steenhoek and L. A. S. Hageman. 1968.
17 S Guide for the specification, testing and inspection of glass reinforced polyester structures in shipbuilding (Dutch). G.
Hamm, 1968.
18 S An experimental simulator for the manoeuvring of surface ships. J. B. van den Brug and W. A. Wagenaar, 1969.
19 S The computer programmes system and the NALS language for numerical control for shipbuilding. H. le Grand, 1969.
20 S A case study on networkplanning in shipbuilding (Dutch). J. S. Folkers, H. J. de Ruiter, A. W. Ruys, 1970.
21 5 The effect of a contracted time-scale on the learning ability for manoeuvringoflarge,sbips(Dutch). C. L Truijens, W. A. Wage-naar, W. R. van Wjjk, 1970.
22 M An improved stern gear arrangement. C. Kapsenberg, 1970. 23 M Màrie refrigeration engineering. Part V (Dutch). A. H. van der
Talc, 1970.
24 M Marine refrigeration engineering. Part VI (Dutch). P. J. G. Goris and A. H. van der Talc, 1970.
25 S A second case study on the application of networks for pro-ductionplanning in shipbuilding. (Dutch). H. J. de Ruiter, H. Aartsen, W. G. Stapper and W. F. V. Vrisou van Eck, 1971. 26 S On optimum propellers with a duct of finite length. Part II.
C. A. Slijper and J. A. Sparenberg, 1971.
27 S Finite element and experimental stress analysis of models of shipdecks, provided with large openings (Dutch). A. W. van
Beck and J. Stapel, 1972.
29 S The equilibrium drift and rudder angles of a hopper dredger
