The development of a contra-rotating
propeller system for large ships
*K Takekuma, DEng,
'41-Sasajima, DEng, *K Saki, IS Nakamura, -j-K Yonekura, tT Ohta,
tY Ujiie and tH Ohira
*Nagasaki Research and Development Centre and I-Nagasalci Shipyard and Machinery Works, Mitsubishi Heavy Industries Ltd
SYNOPSIS
Tomeet with the recent strong demands for energy saving, a contra-rotating propeller (CRP) system was developed by Mitsubishi Heavy Industries Ltd (MHI), which can be applied to large ships. The idea of a CRP is not new and originated about 150 years ago. Since then attempts have been made to apply this system in practice. Even though it was successful in limited cases, such as for torpedoes, commercial application was not realised. Technical problems which stand in the way of the practical application of a CRP system for large merchant ships were: (1) to develop
a
reliable bearing and sealing system for contra-rotating shafts; (2) to develop a compact reversing gear system; (3) to develop an optimum design method and a performance evaluation method of a CRP system. These problems were attacked by MHI recently and put into practice by utilising the oil lubricated plain bearing system, lip Seals, and a star type epicyclic gear system. One
of
the unique aspects is the introductionof
the asynchronous CRP (ACRP) system, where the shaft speed of the forward propeller is lower than that of the aft propeller, by which means a large energy saving was achieved. This system, after one-year's full scale stand tests, was fitted to the car carrier and since then the ship has been put into commercial operation without any troubles and with an energy saving of more than 13%. In this paper, the basic idea of the MN! CRP system, results of stand tests on bearing, seal and shaft alignment, and applicationof
the MN! CRP system to the full scale ship are reported, together with the resultsof
full scale measure-ments and operation experiences.INTRODUCTION
Recent strong demands for energy saving in the operation of ships encouraged many researchers and designers to develop new devices, such as reaction fins (or pre-swirl fins),' the Grim vane wheel,' flow control fins or ducts,' and the contra-rotating propeller (CRP),4 other than applying large-diameter slow-running propellers.'
Among them, the CRP has been recognised as one of the most powerful and understandable methods to improve the propulsive efficiency, because basically the CRP recovers the rotational velocity component left behind the propeller.
The idea of the CRP, originated by J Ericsson's British patent,6 granted in 1836, featured the improvement in propul-sive efficiency without changing the vessel's draught. A CRP was experimentally installed in an Italian training ship, the Christoforo Colombo in 1934,7 and CRPs have been employed in torpedoes for a long time for the purpose of torque balance. Roller type contra-rotating bearings are employed for the torpedoes.
A CRP system was also reported to have been installed experimentally in the American nuclear-powered submarine Jack, commissioned in March 1967,8 as a sophisticated device to reduce operating noise. In this case, the CRP was driven by two 15 000 hp steam turbines. To accommodate the large turbines, the machinery space was lengthened by 10 ft and the stern structure was lengthened by 7 ft to mount the two
propellers. The propellers were smaller than those of other submarines of this class. Although a 10% energy saving was achieved in the case of Jack, the CRP system was eventually replaced by a conventional propeller (CP). It has been reported
Trans IMarE, Vol 102, pp 141-151
Katsuyoshi Takekuma is Deputy General Manager of the Nagasaki Research and Development Center (NRDC), Mitsubishi Heavy Industries Ltd (MHI). He ob-tained a Bachelors Degree in Naval Architecture at the University of Tokyo in 1960 and a Masters Degree from
the same university in 1963. After joining MH I, he worked
for a short time in the Design Section of Yokohama Shipyards, but in 1976 he moved to NRDC. His major field is research into ship hydrodynamics, especially in the wave making resistance of ships. Later he was also involved in the research of ice problems relating to ships and marine structures. After taking the position of Man-ager of the. Ship Hydrodynamics Laboratory in 1984, he obtained a Doctors Degree in Naval Architecture from
Kyushu University in 1986 on the topic of hull form design
of ships. He was promoted to his present position in
1989.
that a tapered-land bearing,' a type of plain bearing, was
employed in this case.
Recently there have also been commercial CRP systems for outboard engines of motor boats.' Roller type contra-rotating
bearings and bevel gears are used in this case.
However, there has been no precedent for using CRPs for conventional large-sized merchant ships up to the present time. The reasons for this are considered to be:
compact contra-rotating bearings with long bearing life and easy maintenance had not been developed;
no small and simple contra-rotating mechanism had been developed;
K Talcelauna etcd Rear propeller Special plainilbearIng Hull Inner shalt Front Propeller Outer shalt
Single .stagetstat type CO ntra-rotatinuptanetary gear
3.. a CRP system with sufficiently high energy saving merits to stimulate economic interest had not been developed. To overcome these difficulties, the study of the CRP system with plain bearings startedat Millin 1983.
After 5 years of basic study with model: tests and full scale stand tests, finally in 1988, the MHI type CRP system was
suc-cessfully installed in a 4200 CT car carrier." After l'/2 years
commercial operation of the ship, the reliability of the system was proven.
In this paper, featuresoftheMHI CRPsystem; development of a contra-rotating plain bearing and gearing system,
hydro-Fig ttofftfa-Wkating propeller shafting for Toyofull No 5
Table Energy say ng of CRP studied
dynamic design considerationsoftheMHI CRPsystem, and full scale application experiences are reported and discussed.,
FEATURES OF THE MHI CRP SYSTEM
The CRP system developed by MHI has the following:
technical features:
.1. A special plain bearing is employed forthe contra-rotating bearings. This bearing is a type of hydrostatically
Case"- A B il C
Research
WI
MARAD ISTAL-LAVAL NSMB DTNSRDC SRI SRliInstitute (DTNSRDC) (SSPA)
Reference Not' I
0)
I (13) (14) 1 (15) 1
06)
(17) (18)J
Ship type Tanker Container Container Tanker Cargo Tanker Container Container
ship ship ship ship ship
Hull LPP (m) 222.00 1 228.80 156.00 195.07 158.50 ; 278.90 200:00 200.00 B (m) 44.00 28.04 23.64 25.91 I 22.40 I 40.23 29.00 29.00 D (m) 12,19 8.23 8.75 1 10.33 8.84 14.96 I 10.50 10.50 II. (t) 96 969 30 060 19 500 42 905 19 023 136 130 1
-
-"Hp (ps) 17 000 30 000 25 000 16 000 16 000 43 000-
34 900 Engine ' (MCR)I IN (rev/min) CP 90 128 128 120 120 105 104 103 CRP 70 90 1 84 120 120 105 104 82 1 'CP ID (m 7.1 .2 5 6.4 1 5 6.0 4 6.4 4 6.0 4 7.6 511 7.4 1 5 7.4 5 1 CRP R/DA (m) 7.1/6.62 ,, ZIZ, 4/5 6.4/6.1 5/4 6.0/- i 4/4 I' 5.2/5.0 4/5 5.7/4.9 4/5 6.9/6.4 il 4/5 6.6/6.3 4/5 17.18/6.8911 7.18/6.89' 4/5 4/5I 11Condition Full 'Ballast Design
Loaded Loaded Loaded Design Design' Design ,Evaluation, t
Propeller
,no tok 7(a)
11 (%) 12(b) 7(b) 13(b) 8(a) I 12(b) -4(b) , c 4(b) 3(b) 6(b)
-4(c) , -4(a) 3(b) 6(b) 7(b): 7(b) 6(b)' 'Condition Ballast Loaded Ballast
efficiency no (°/0) 7(a) I
li: Propulsive q (%) 14(b)i
efficiency
-13(b) L I -10(b) 8(b) I 1A: Diameter of forward propeller of CRP equal to CP ; B: Shaft speed of CRP is equal to that of CP; C: Shaft speed' or diameter of forward
propeller of CRP is arbitrarily chosen
t References are shown at the end of this paper
t (a); Kt/J =constant, (b); Vs=constant, ( c); HP =constant
Main engine
Cuter shaft
ueoge shaped channei
Irarr shaft
Floating WW1
Tapered-land bear ins Floating bushtype bearing lubrication oil
Oil supply hole Hydrostatical I y lubricated bearing
lubrication oil
Fig 2: Plain bearings for CRP Table II: Candidate plain bearings
Table ill: Principal particulars of full scale stand tester
Table IV: Monitoring items of 1800h endurance tests by full scale stand tester
Trans 1MarE, Vol 102, pp 141-151
lubricated plain bearing with a white metal lining designed for long bearing life, easy inspection, to be free from maintenance, and to be compact in
size like a conventional stern tube bearing.
2. A star type epicyclic gear system is employed as the reversing and reduc-ing mechanism of the rotatreduc-ing forward propeller, while the aft propeller is directly connected to the main engine. This gear system has been widely used in conventional ships and its reliabil-ity in operation is widely acknowledged.
An asynchronous CRP (ACRP) system is employed. The existing concept of CRP systems has been the synchro-nous contra-rotating system (SCRP), ie with concentric dual shafts rotating in opposite directions at equal speeds. However, in the case of the ACRP system, concentric dual shafts rotate in opposite directions at different speeds. The ACRP system can achieve higher energy saving than the SCRP system by lowering the revolutions of the forward propeller.
Figure 1 shows the MHI CRP system designed for Toyofuji No 5.
Regarding the hydrodynamic design of CRP, the first complete design method was reported by Morgan,12 in 1960. Since then, evaluations of energy saving by model tests were conducted by several people (see Table I). But all of those studies were for the SCRP system and no design methods existed for the ACRP. Thus, a model test based approach was used in designing ACRPs.
DEVELOPMENT OF CONTRA-ROTATING
BEARING AND GEAR SYSTEMS
Selection of bearing type
In the very early stage of the development, three types of bearing were evaluated from the view points of loading capa-bility, toughness to unsteady loading, operational life, and ease of maintenance and manufacture. From the beginning the
roller or ball bearing type was thought to be not feasible
practically due to the following reasons*:
limited fatigue life of roller bearings due to metal contact; weakness of the sealing systems (intrusion of sea water); estimated rather large boss diameter for the forward
propeller (outer shaft).
Candidates were tapered-land, floating bush type and hy-drostatically lubricated bearings (which is MHI's new idea), since theoretically hydrodynamic lubrication due to wedge action can not be expected if the inner shaft and outer shaft are rotating in opposite directions at the same shaft speed.
Figure 2 shows the schematic drawings of candidate bear-ings. Table II summarises the comparison of three types of candidate bearings using data already existing, calculated data and data obtained by small scale mode tests. Finally, a hy-drostatically lubricated bearing was chosen and a full scale
stand test was planned.
Recently Ishikawajima-Harima Heavy Industries Ltd, developed a CRP system with roller bearings and installed iton a 36 000
dwt cargo ship."
Type Tapered land
bearing Floating bush bearing Hydrostatically lubricated bearing Mechanism Hydrodynamical lubrication Hydrostatical
lubrication Loading
capacity Very small Small Large
Remarks , X Difficult to keep accuracy in manufacturing X Problem in keeping strength of floating bush X Difficult to manufacture floating bush X One side touch or one point touch
of the inner shaft to the bearing
0 Flexible for loading change
Items Detail
Dimension of tester L = 15.5m, H = 3.3m (from the ground level to the shaft centre) Inner shaft Outer shaft Diameter of shaft
Shaft speed - 94 rev/min510 mm
790 mm
- 67 rev/min
Loading capacity Dynamic load
18.4 ±4.6t at 6.2 Hz Static load 28t Seal Lip material Circumferential velocity Fluoric rubber - 4.7m/s
Items Check Measuring items
Hydrostatically lubricated bearings (inner and outer shafts) Loading capability
1. Oil film thickness
and shaft inclined angle 2. Erosion or abrasion of white metal Lubrication oil 1. Amount of lubrication oil supply 2. Temperature of oil
Sealing Sealing capabilityunder dynamic
loading
1. Abrasion of seal lip
and liner 2. Temperature of
seal lip
3. Pressure at lip seals
Ojcar shaft Inner c:.ft siva hirer slatt
3..
-L l I ' [I IK Takekuma et al
Inner Shaft (510mrndia.)
Pressurized Water Jacket for Aft Inner Shaft Seal
The part containing Aft Inner Shaft Bearing Pressurized Water Jacket for Aft Outer Shaft Seal
The part containing Aft Outer Shaft Bearing
Bearing Housing Fore Outer Shaft Seal
AN. 11173 i , . - ,71.171.141S3R. \ x. . - .. :.,.. '.4-_ . .., m.
a
5r. Nfirb" Outer Shaft (180mmdia.) 5Loading Plunger(Simulating fore Propeller)
Vibrator (Simulating Aft Propeller)
Fig 3: General view of full scale stand tester for contra-rotating shafting
Fig 4: influence of bearing load on relative inclined angle and oil film thickness of inner shaft bearing
Selection of sealing system
Another important technology relating to the contra-rotat-ing shaftcontra-rotat-ing is the selection of a lubrication oil sealcontra-rotat-ing system.
Table V: Principal particulars of ToyolujI No 5
Kind of ship
Gross tonnage
Ship's length
(between perpendiculars) Ship's width (mould) Full loaded draught (mould)
Main engine
Maximum output (before modification) (after modification)
Car nne.IvV11., le<1.11 /WIrrtes
4176.89t 110m 19.1m 6.05m Mitsubishi-6 UET 52/900 6000 hp X 198 rev/min 5400 hp X 185 rev/min
Table VI: Principal particulars of CRP for Toyofujl No 5
Among seals used for marine application, such as lip seal,
mechanical seal and segment seal, the lip seal is the most
popular for stern tube bearings.
But., in the case of the CRP, a large deflection of the seal lips is expected due to the combination of shaft forces of both pro-pellers. So, small scale model tests, with 150 mm diameter inner shaft, were conducted to check if the hydrodynamical lubrication between seal lips and shaft liners is kept or not, if the response of the seal lips to the eccentric movement of the inner shaft is enough or not, and if the abrasion of the seal lips
CRP CP Forward propeller Aft propeller D (m) 3.7
as
3.7 P 0.811 1.000 0.805 Ao/Ad 0.70 0.42 0.37 d/D 0.187 0.195 0.178 Z 5 4 3Voyaging speed at normal output 15.6 kn
Driving gear for Inner Shaft Driving gear for
Outer Shaft
TO 14(m)
4
Ship_ with CP Engine Resistance 4 Self-propulsion factors (SPF) Ship_mith CRP Engine Estimation of S.P.F. Principal Dimension of CRP Va, Detail design of each propeller for CRP Powering (matcling of BHP and N)
-3--Final Evaluation Propeller chart Stern arrangment Expected Energy saving optimization dataFig 5: Flow chart for design of ACRP
O.4 0. 5 0.6
Fig 6: Example of effect of thrust loading on energy
saving
or shaft liners is within an acceptable level or not at relatively high speed operation. The conclusion was that the lip seal is
Trans IMarE, Vol 102, pp 141-151
feasible even for such severe operating
conditions, because the supplied pressure level of the lubrication oil was reasonable and no special temperature rise at the lip seal was recorded.
Development of reversing and
reduction gear system
Development of a reversing and reduc-tion gear system was done basically in re-spect of reliability, compactness and manu-facturing cost. In the Mal CRP system, a star type epicyclic gear system was chosen. Compactness was especially achieved by adopting the ACRP system, the mechanism of which is shown in Fig
Stand tests for contra-rotating
shafting system
The most important aspect of applying a CRP to full scale is reliability to ensure the time scheduled operation of the ship.
Even though the hydrostatically lubri-cated bearings and lip seals were chosen to be the most suitable, it was thought neces-sary to check the system, including the effect of shaft alignment, under the same condi-tions as the ship operation. Assuming a bulk carrier of about 60 000 dwt, the
contra-rotating shafting system was designed and manufactured as the full scale stand tester as shown in Fig 3. Its principal particulars are shown in Table III.
To simulate the full scale ship conditions, the tester was equipped with such devices as a loading plunger at the aft end of the outer
shaft to simulate the bearing load of the forward propeller, a vibrator at the aft end of the inner shaftto
simulate dynamic loading due to the aft propeller, and a pres-surised water jacket at the aft part of the seal system for
imposing sea water pressure corresponding to the draught of the ship.
Two kinds of test were conducted as follows:
Measurements of bearing characteristics, such as the bearing oil film thickness, oil film pressure distribution, bearing temperature, and lip seal temperature.
Endurance test of the system up to 1800h.
Results of the full scale stand tests for Mare summarised as follows:
Loading capability of the contra-rotating shafting was found to be satisfactory. Figure 4 shows the relationship between the shaft loading and the minimum oil film
thickness of the contra-rotating bearing. In this case a static load, corresponding to the weight of the propeller and an additional dynamic load of about 25% of thepro-peller thrust, was applied vertically by using a hydraulic system. In this case more than 50 tun thick oil film was produced.
Pressure distribution along the bearing was found to be similar to that estimated by calculation.
Amount of lubrication oil supply under the vibratory vertical load was found to be within the estimated value. As to the endurance tests for 1800h, the items listed in Table IV were checked. The results are summarisedas follows:
-, D,44 8 A B 3 4 / 3 / 5 10.5/10.010.0/ 9.7 67/6750/50 C 4 / 3 10.5/10.0 50/67 D 4 / 3 10.5/ 9.7 50/67 10. 12 14 Data base Correlation data OF/D., Np/N. S.P.F)c. etc f(Tx/T) 2. A,/N
K Takekuma et al
Table VIE Principal particulars of contra-rotating shafting
DHP (PS) 3500
3000 Optimum
20 30 40 50
TA : Thrust of Aft propeller
T I Thrust of Total VS: 15.6kn NE: 138.75rpm NA: 185rpm 60 70 TA/T(%)
Fig 7: Optimum thrust loading ratio for Toyofuji No 5
I. Thickness of oil film and temperature for inner and outer
bearing were kept almost the same as those measured in the short time test, and it was concluded that this type of plain bearing is feasible. Also it was confirmed that no erosion of white metal and no choking at lubrication oil supply holes occurred.
Temperature of seal lip was stable at about 60-70°C
throughout test duration.
Abrasion of forward seal at the inner and outer shafts was almost at the same level as on the conventional ship. No intrusion of surrounding sea water into the bearing
part was found, even though a little oil/sea water mixture was found in between the seal lips.
Thus full scale stand tests gave MHI confidence that the hydrostatically lubricated bearing and seal system is feasible.
DESIGN OF CRP AND ITS EVALUATION
BY MODEL TESTS
Existing data of energy savings
The energy saving possibilities of the SCRP, which have
Fig 8: CRP fitted to Toyofuji No 5
appeared in various publications, are between 3 and 14% in comparison with the CP, varying widely with the design and operation conditions of the CRP, as summarised in Table I.
Where the SCRP is designed to rotate at the same shaft
speed as the CP (case B in Table I), the tanker studied at NS MB, for example, gave a 4-8% energy saving in compari-son with a CP at the same speed. If the SCRP is designed so that the diameter of the forward propeller is kept the same as the CP (case A in Table I), energy saving will increase by up to 12-14%, as for example in the case of a tanker studied by MI-I, since in this case the merit of reduction of the shaft speed of the propeller can further be added.
CP CRP
Inner shaft Outer shaft
Propeller shaft 370mm dia 360/90mm dia 540/430mm dia
Stern tube bearing
Type Hydrodynamical lubrication Hydrostatical lubrication Hydrodynamical lubrication Length 400F/1 000Amm 250F/625"mm 2501/850Amm
Material White metal Lube oil SAE No 30, 40°C
Seal Lip No 400 2F/3A No 380 2F/4A No 600 2F/3A
pear
Star type epicyclic gears Reduction gear ratio =1.333:II Number of star gear:9
Power 2700hp
Elastic coupling Geislinger coupling
-I U
-Main Engine-Horse Power (PS) 6,000 5,500 5,000 4,500 4,000
'fc,y(..,I tij i
No 5
Fig 9: Ship speed-horse power curve at sea trial
Fig 10: Resultsof turningtest
In order to make the CRP system an effective energy saving device which will be economically attractive for merchant shipowners, it is necessary to develop a CRP system not only with higher energy savings but also without excessive modifi-cation of the main engine, engine room, and aft part of the ship in order to install it. From this point of view MHI's ACRP is
effective, since it has the advantage of using a compact
reversing and reduction gear system.
Design of ACRP
Since all the existing design methods of the CRP is for a
SCRP, the authors' company started to build up a design method for the ACRP based on that of conventional propellers and model test data. A general design flow chart thus builtup is shown in Fig 5. Self-propulsion factors for forward and aft propellers were analysed separately from model test data of several combinations of forward and aft propellers. Then, systematic calculation was made to study the relationship between energy saving and other factors such as Dr/D, Z/Z,
Trans 1MarE, Vol 102, pp 141-151
C RP
STEADY ASTERN REVOL
SHIP'S HEADING "62.0 DEC. ,'
ASTERN SPEED 10 KN. '
0' -J STEADY SHIPS ASTERN SPEED
SHIP'S HEADING P15.0 DEC. ASTERN SPEED 10 KN.
10 40 TIME OF ASTERNORDERED
ROWSHEADING 0.0 DEC. AHEAD SPEED 17.0 KN. RUDDER AMIDSHIPS
SHIP'S HEADING '10.0 DEC. STOP DISTANCE 1.511 HI
P SHIP'S HEADING '26.0 DEC. STOP DISTANCE 1,511 It
MAIN ENGINE STOP AND ASTERN REV. START
2'
Fig 11: Results of crash stop astern test
and NINA. But the important point is how to choose the
optimum loading ratio of forward and aft propellers. So the self-propulsion tests were also conducted by using CRP models with an adjustable pitch angle. Figure 6 shows an example of such model test results and it can be said that the optimum con-dition exists around T,/T = 0.55 in this case.
FULL SCALE APPLICATION OF THE
SYSTEM
Subject ship
A Mitsubishi-type CRP system was retrofitted to the Toyofuji No 5, a 4200CTcar and a CKD carrier, as the first application of a CRP to a merchant ship in the world. This ship, owned by Toyofuji Kaiun Kaisha Ltd, was delivered from the Nagasaki Shipyard and Machinery Works of MHI in 1979. Table V shows the general specifications of the ship.
The CRP
Design of CRP
The design of the CRP for Toyofuji No 5 is based on the method mentioned before. Since more than 10% energy saving was expected, the main engine output was reduced by 10%. The diameter of the aft propeller was chosen to be the same as that of the conventional propeller, since no vibration problems induced by the propeller were reported for the CP.
Another feature of the CRP fitted to Toyofuji No 5 is that the shaft speed of the forward propeller was chosen so that the blade frequency of the forward and aft propellers is the same.
Measured Standard --Correa cd Analysis C P '440, For Wave
op
ii
i
/
pr.
/
CRP,
,
/'
_
El
c>vc.luii N. 5 IN) 300 200 10C 0 ,..'..09f..7--.1...,>CRP CP 1 I 90 I 160' a 1"' 210' 360' ,,c' V c,360'\
2 0', N411.110.r
`1411.1111.' MN° 300 200 100 {D., 100 200 as 400t,11SHIP'S CONDI ION : BALLAS1
SHIP'S SPEED 17KN HOOD ANGLE: 4- 35:
CRASH STOP ASTERN TEST
FULL.. MEAD FULL ASTERN
BALLAST COMMON ) SCAIE
a ttilFil 11,5 18 Ship Speed (Mn) 16 16.5(16.1) 11 0 8'
/
I
CP Model Test CRP = 20° (Aft Propeller) (n 1 1.0 1 co 4a) =340 (Fwd Propeller) CRP
Full Scale Observation
FKO SIDE
HA1N-124
AF1 SIDE
AhIM-159
TEST NO. TY5 17G NJ = 1I19 RPH
tO 13 CA-2601 0J R0. I. 120 5 150 6 160 7 210
Fig 13: Trace of Inner shaft centre movement in bearing
This simplified the check of the vibration response of struc-tures to propeller exciting forces. The principal particularsof the CRP for Toyofuji No 5 arc shown in Table VI and the principal particulars of the contra-rotating shafting are shown
in Table VII.
Model tests on CRP
Model tests were conducted by using a 1/15.7143 model in K Takekuma et al
rl-ray IL.Cavitation patterns; a 11d1 1.17111;11110119 ttrigiriu FrIcIXIMUFF1 output
a towing tank, cavitation tunnel and !seakeeping basin, where the following items were studied:
propulsive performance in normal operation; wake measurements behind propellers;
check of cavitation characteristics and pressure fluctua-tion:
manouevring characteristics.
The results are discussed later in comparison with full scale data.
Selection of optimum condition
To achieve the highest energy saving, the optimum thrust loading ratio for Toyofuji No 5 was studied by changing the pitch of the forward and aft propellers in self-propulsion tests. Mean be seen in Fig 7, it was shown that the optimum condition can be obtained at TIT = 0.52.
Construction and fitting
Due to the limitation of time for retrofitting the CRP system for Toyofuji No 5, the stern part of the ship was replaced by the new one with the CRP and shafting system installed before-hand. Also the measuring sensors for oil pressure, clearance, oil temperature, etc were installed. It took 15 days to fit the new
stern part. The replacement was done in No 3 dock ofthe Nagasaki Shipyard and Machinery Works starting on 8 August
1988.
A special team was organised to install the stem part,
keeping the accuracy of shaft alignment. Figure 8 shows the CRP system fittedto Toyofuji No5.
Results of sea trial
Evaluation of energy saving
Figure 9 shows the results of the sea trial of ToyofujiNo5 with the CRP system, in comparison with the CP. About 16% energy saving was achieved.
Manouevring test
Figure 10 shows the results of the turning test in comparison with the CP installation. If we take into account the effect of wind, no difference is seen between both cases.
3 90 240 9 270 10 300 II 333 A
Fig 14: An example of long term monitoring records of temperature
Sn
foyo-Fuj i
No.Fig 15: Change of metal contents in stern tube lubrication oil
Figure 11 shows the results of crash stop astern tests. Almost the same results of time and distance until the ship stopped were obtained. A tracing of the ship after it started to move backwards was found to be almost straight in the case of the CRP, while in the case of the CP the ship made a turn towards port side due to the effect of rotational wake of the propeller.
Cavitation observation
Cavitation observation was conducted through the observa-tion windows at the bottom of the ship's hull by using a video camera. Since no cavitation observation tests were conducted for the CP, direct comparison could not be made. But, as shown in Fig 12. cavitation of the forward propeller was almost to the same extent as that for the CF and no cavitation was observed on the aft propeller. The pressure fluctuation level was esti-mated to be reduced a little.
Vibration and noise in engine room of ship
The vibration level was generally reduced, and also noise
Trans 1MarE, Vol 102, pp 141-151
level in the engine room was reported to be reduced, but it is not clear if the reduction of noise in the engine room is caused by replac-ing the CF by the CRP or by reduction of the engine power.
Emergency operation
An emergency situation for the inner shaft bearing would be the shut-down of the lubri-cant oil supply by accident or some other
unknown reason. But, according to the full scale stand tests, it was found that even if the oil supply was shut-down intentionally, the rotating inner shaft was still supported by the hydrodynamic action of the lubrication oil, even though the loading capability decreased. This means that by checking the conventional monitoring system, such as meters for the lu-brication oil supply and oil temperature, the
bearing system will be secured. Such an emergency case was tried in the sea trial and confirmed that the shafting system was completely safe.
Contra-rotating bearing
The following data were obtained regarding the contra-rotating bearing:
The oil film thickness decreased with increase of the shaft speed, but even in the turning condition the oil film thickness was kept consistent enough and no trouble was reported.
The behaviour of the inner shaft was considered to be normal, as shown in Fig 13, if we consider the shaft forces induced by the aft propeller.
The temperature range of the seal lip and bearing material of the inner shaft was kept normal even with an increase of propeller shaft speed.
The scatter of temperature among the three axes of the planetary gear was within 2°C. This means that the load-ing on each planetary gear was almost even.
Experience of long term operation
Toyofuji No 5 departed to Nagoya from Nagasaki on 29
80
40 ,
30
5fl
30 i
Seal lip temperature (CC)
Toyc>fu.ji
Nic.. 5
1 I
:-
. I..
[ I
6O'
I I 1Outer bearing temperature (°C)
Lubrication oil temperature (°C) 0 in
tout
. ,
!
1 L
1
Sept. Oct. Nov. Dec',
Jai-
Feb. March April May June July Aug.Sec). Oct. 'icy. Dec. Jan. etr. 'larch Aorll ,tay June July
1988 1989 40 Fe 20 5 1988 1989
K Takekuma et al
Fig 16: Surface condition of inner shaft inspected at
overhaul
Fig 17: Ship speed-horse power curves in operation
August 1988. After one year's operation, the ship came back to Nagasaki for inspection. During the voyage the ship expe-rienced no problems and data obtained during the voyage were stable. For example, the results of the temperature monitoring of the seal lip, outer bearing and lubrication oil during one
year's operation are shown in Fig 14. It is clear that the temperature was stable. Further, in Fig 15, the change of Fe and Sn in the lubrication oil analysed is shown. The data show clearly that no abrasion was going on in the inner shaft bearing and it was also confirmed by direct inspection of the white
metal surface of the inner shaft at the occasion of the overhaul after one year's operation, as shown in Fig 16.
Figure 17 shows the power curve obtained by analysing log book data and correcting for the effect of winds and waves. The data used were obtained from September 1986 to August 1988 (24 months) for the CP and from July 1988 to July 1989 (11 months) for the CRP. As shown, an average 13.5% energy saving was achieved. If we take into account the scatter of the data, energy saving was between 10 and 17%.
CONCLUSIONS
Anew CRP system was developed and successfully fitted to
the car and CKD carrier Toyofuji No 5. She is now in her second
year at sea without any problems. It is reported that the ship's crew is operating this ship with the MI-II CRP system in the
same way as a conventional ship.
Through these experiences, a great deal of confidence has been obtained in applying this MHI CRP system to larger vessels such as VLCCs or to high powered container ships.
In this paper there has been no discussion regarding the results of the basic research, the theoretical approaches to estimating propulsive performances of the CRP, based on the propeller lifting surface theory, and the dynamic response of propeller shafts and oil film in hydrostatically lubricated plain bearings, based on the lubrication theory and alignment analy-sis methods. Such analytical studies are important in designing the CRP system and are also carried out in parallel with the development of the ACRP. Technical papers on such topics will be presented at other occasions in the near future.
ACKNOWLEDGEMENTS
In developing the new ACRP system, Min co-operated with the Japan Foundation for Shipbuilding Advancement
(JAFSA) for three years. Retrofitting the CRP system to
Toyofuji No 5 was achieved through the kind co-operation of Toyofuji Kaiun Kaisha Ltd. MEI would like to express its heartfelt appreciation to J AFS A and the shipowner for their co-operation.
As to the new Mitsubishi CRP system, patents have been applied for in Japan, the USA, the EC, Canada, Spain and Korea.
REFERENCES
H Tanibayashi, 'Development of reaction fins for full ships', Mitsubishi Juko Giho, Vol 12, No 4 (1980).
0 Grim, 'Propeller und Lei trad' Jahrbuch der Schiffsbautech-nischen Gesellschaft, Band 60 (1966).
E J Stierman, 'The design of an energy saving, wake adapted duct'. Proceedings of the Third International Symposium of Practical Design of Ships and Mobile Units, Trondheim (1987).
M Nakanishi and T Ikeda, 'A study on performance of
contraro-tating propellers', Trans West Japan Society of Naval Archi-tects, No 70 (1985).
H Tanibayashi and K Takekuma, 'Hydrodynamic
perform-ances of various devices for the improvement of propulsive
performance of ships', Mitsubishi Heavy Industries Ltd,
Technical Review (October 1982).
J Ericsson, 'Propeller for steam navigation', AD 1836, No
F Rotundi, 'Trials of the training ship Christoforo Colombo' with two co-axial turning screws', RINA (1934).
'Permit' class submarines, Janes Fighting Ships 1986-87, p
701 (1987).
M Meier-Peter, 'Engineering aspects of contra-rotating
propul-sion system for seagoing merchant ships', Joint Meeting of
Schiffsbautechnische Gesellschaft ev and Associazione aria di Technica Navale (May 1972).
10, J Savikurki, 'Contra-rotating propellers', HANSA-Schiffahrt-Schiffbau-Hafen, 125 Jahrgang, No 2 (1988).
S Nakamura et al, 'World's first contra-rotating propeller
system successfully fitted to a merchant ship', The Motor Ship
111.11 International Marine Propulsion Conference and
Exhibi-tion Economy and Reliability, The Motor Ship, London
(1989).
W B Morgan, 'The design of contra-rotating propellers using Lerbs' theory', Trans SNAME, Vol 68 (1960).
A W Forrest and G C S wensson, 'Application of contrarotating
propulsion system to a US flag merchant vessel: technical
Trans IMarE, Vol 102, pp 141-151 report, phase 1, model testing and preliminary installation
design', MARAD report No HA-RD-91-80072 (1980). G A Lane, 'Study of contra-rotating propeller arrangements
for merchant ships', Final Report by the Swedish Working
Group in 1965 for Studies and Development of Contra-Rotat-ing Propeller Arrangements (1970).
J D van Manen and M W C Oosterveld, 'Model test on contra-rotating propellers', International Shipbuilding Progress, Vol 15, No 172 (1968).
J B Hadler, W B Morgan and K A Meyers, 'Advanced propeller
propuLsion for high-powered single-screw ships', TransSNAME,
Vol 72 (1964).
Y Ukon, Y Kurobe and C Arai, 'Cavitation tests of contraro-Eating propellers', Proc of the 44th Technical Presentation of Ship Research Institute (1984).
Y Ukon et al, 'On the design of contrarotating propellers',
Trans West-Japan Society of Naval Architects, no 75 (1988). S Nishiyama, 'Juno 37 000 dwt class contra-rotating propel-ler vessel', lshikawajima-1 larirna Gino, Vol 29, No 4(1989).
8. .9. hall-13. 14. 16. J7.
Unconventional propulsion systems discussion
Discussion
A Stevens (MoD) My question is addressed to the authors of both papers on multi-fuel diesel engines. Are there secondary effects from the use of gas in diesel engines, for example
lubricating oil contamination and consumption, which might reflect on the economics of this type of operation?
0 Grime (MAN B&W Diesel A/S) This question gives me the opportunity to highlight one of the special additional advan-tages of the high pressure gas injection diesel engines. As the
gas is injected into the engine only after the compression stroke, and after the pilot oil is already injected, gas is not at all in contact with the lub oil film in the cylinders, as is the case in a low pressure gas engine. Therefore, gas and oil cannot mix, and no special contamination takes place.
Unlike low pressure gas engines, high pressure gas engines do not, therefore, require special lubricants. On top of that gas is clean and low in sulphur so the additives content of the lub oil can be low.
In any case please note also that in low speed crosshead diesels the crankcase and the cylinder section are separated in way of the piston rod stuffing box, which is why the crankcase oil contamination is negligible.
B Engesser (Sulzer Diesel Ltd) There are perhaps four points that Mr Stevens might consider as secondary effects.
The boil-off gas compressor imposes a small power
con-sumption penalty that can be 3-6% of the engine power,
depending upon the compressor suction conditions (see Fig 14). In addition the thermal efficiency of the engine itself is very slightly less in dual-fuel operation than in liquid-fuel
operation (see Fig 24). Thus both effects would result in
slightly higher operating costs for the dual-fuel engine com-pared with a liquid-fuelled diesel engine. But the dual-fuel engine still offers economical advantages over steam turbine plant, while a liquid-fuelled diesel engine would need a boil-off gas reliquefaction plant.
Cylinder lubricating oil consumption will be somewhat higher in dual-fuel operation than in liquid-fuelled operation because of the missing sulphur from heavy fuel oil which has a certain lubricating effect. On the other hand, when operating on minimum heavy fuel oil, low-alkalinity cylinder lubricating oil can be used and that is cheaper than high-alkalinity oil. The net result is probably no effect on operating costs compared with a liquid-fuelled diesel engine.
It is noticeable that all combustion space components, the
exhaust system and the piston underside spaces are much cleaner for dual-fuel operation as natural gas is a much cleaner fuel than heavy fuel oil. In general, cold corrosion problems in the combustion space and exhaust system will also become less significant in dual-fuel operation. The result of both of these aspects is lower maintenance costs compared with a heavy-fuel burning diesel engine.
Dr A Fowler (The University of Newcastle upon Tyne) With regard to the boil-off and compression of gas could the authors of both papers on multi-fuel diesel engines add some details as to the type of compressors used, the temperatures at various locations, whether or not cooling is required and if so by what medium.
0 Grone (MAN B&W Diesel A/S) We, as engine designers, only specify the amount of gas and outlet pressure and
tempera-152
ture. How this is met by the compressor maker is not our
concern, except for the power consumption for compressor work, which influences total economy.
The compressors used are also used in the chemical industry and several manufacturers can comply with the requirements.
B Engesser (Sulzer Diesel Ltd) In reply to Dr Fowler, the boil-off gas compressor envisaged for a 17 MW Sulzer RTA dual-fuel engine would be of the balanced-opposed reciprocating type from Sulzer Burckhardt. It would be a five-stage inter-cooled compressor with the first three stages having unlubri-cated labyrinth pistons while the last two stages would be lubricated. The maximum temperature at the outlet of each stage would be 160°C while the final temperature after the end cooler would be 47°C. The delivery pressure of 250 bar would be regulated by thyristor control of driving motor speed.
A Stevens (MoD) Dr Fowler, in his presentation, described the method by which the 'gamma' of the working fluid in the Argo-diesel system was measured. Could I ask whether, for the measurement of 'gamma', it would be possible to use pressure measurements from 'continuous' adiabatic compression of the working fluid in a nozzle instead of from the 'discrete' com-pression of a sample of the working fluid in an engine cylinder?
Dr A Fowler (The University of Newcastle upon Tyne) In answer to Mr Stevens, the issue of gamma measurement using a `venturi' nozzle was in fact given serious consideration
during the earlier phases of the project. In principle the value of gamma may be obtained from the relationship between the upstream and throat pressures when a gas expands through a convergent/divergent nozzle arrangement.
The advantage of the system is its simplicity; a small compressor draws a sample of gas from the system, compresses and discharges through the nozzle, and discharges back into the main system. Highly accurate pressure transducers are re-quired since the variations in pressure corresponding to the range of gamma experienced is very small.
Such a system was developed and tested on the Argo-diesel engine during the Newcastle trials, but consistent problems were encountered with respect to accuracy, repeatability and calibration.
This was attributed to the effect of relatively trivial fouling of the venturi, thereby causing disproportionate variations in
readings, despite repeated attempts at filtration of the gas sample. The basic problem is that the resolution of the process is proportional to the pressure ratio which, in the case of a choke nozzle, is limited to the range of about 2:1.
By resorting to the higher pressure ratios associated with the reciprocating piston, the resolution of the device is improved by an order of magnitude and, in practical terms, this system was found to give far superior results.
Rr Adml M A Vallis (Deputy President, The Institute of Marine Engineers)
In presenting Mr Person's paper Cdr Forbes had spoken of underwater endurance of the Neicken being improved by a factor of five to seven. Could he say whether this referred to time spent underwater or to distance travelled? It appeared that the power of the Stirling engines now
being produced in Sweden was 75 kW max, 65 kW
