vol.44, june.2017 issn: 2354-7065 e-issn: 2527-6085isomase.org/jomase/vol.44 jun 2017/vol-44.pdf ·...

ISSN: 2354-7065 & e-ISSN: 2527-6085Vol.44, June.2017

Journal of Ocean, Mechanical and Aerospace -Science and Engineering-

Vol.44: June 30, 2017

ISOMAse

International Society of Ocean, Mechanical and Aerospace -Scientists and Engineers-

Contents

About JOMAse

Scope of JOMAse

Editors

Title and Authors Pages

Characteristics of Dynamic Response of Suspension Hydraulic Motor - Regenerative Shock Absorber (HMRSA)

Kaspul Anuara, Harus Laksana Guntur

1 - 7

Analyze Performance of Double Acting Tanker While Running Astern in Ice Condition

Efi Afrizal, J. Koto

8 - 20


Vol.44: June 30, 2017

ISOMAse


About JOMAse

The Journal of Ocean, Mechanical and Aerospace -science and engineering- (JOMAse,

ISSN: 2354-7065 and e-ISSN: 2527-6085) is an online professional journal which is published

by the International Society of Ocean, Mechanical and Aerospace -scientists and engineers-

(ISOMAse), Insya Allah, twelve volumes in a year. The mission of the JOMAse is to foster free and

extremely rapid scientific communication across the world wide community. The JOMAse is an

original and peer review article that advance the understanding of both science and engineering

and its application to the solution of challenges and complex problems in naval architecture,

offshore and subsea, machines and control system, aeronautics, satellite and aerospace. The

JOMAse is particularly concerned with the demonstration of applied science and innovative

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into the use of computational fluid dynamic, heat transfer, thermodynamics, experimental

and analytical, application of finite element, structural and impact mechanics, stress and strain

localization and globalization, metal forming, behaviour and application of advanced materials in

ocean and aerospace engineering, robotics and control, tribology, materials processing and

corrosion generally from the core of the journal contents are encouraged. Articles preferably should

focus on the following aspects: new methods or theory or philosophy innovative practices, critical

survey or analysis of a subject or topic, new or latest research findings and critical review or

evaluation of new discoveries. The authors are required to confirm that their paper has not been

submitted to any other journal in English or any other language.


Vol.44: June 30, 2017

ISOMAse


Scope of JOMAse

The JOMAse welcomes manuscript submissions from academicians, scholars, and practitioners for

possible publication from all over the world that meets the general criteria of significance and

educational excellence. The scope of the journal is as follows:

• Environment and Safety

• Renewable Energy

• Naval Architecture and Ship Construction

• Computational and Experimental Mechanics

• Hydrodynamic and Aerodynamics

• Noise and Vibration

• Aeronautics and Satellite

• Engineering Materials and Corrosion

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Any queries please send an email to the following addresses: [email protected] or

[email protected]


Vol.44: June 30, 2017

ISOMAse


Editors

Chief-in-Editor

Jaswar Koto (Ocean and Aerospace Research Institute, Indonesia)

(Universiti Teknologi Malaysia, Malaysia)

Managing Editor

Dodi Sofyan Arief (Universitas Riau, Indonesia)

Associate Editors

Ab. Saman bin Abd. Kader (Universiti Teknologi Malaysia, Malaysia)

Adhy Prayitno (Universitas Riau, Indonesia)

Adi Maimun (Universiti Teknologi Malaysia, Malaysia)

Ahmad Fitriadhy (Universiti Malaysia Terengganu, Malaysia)

Ahmad Zubaydi (Institut Teknologi Sepuluh Nopember, Indonesia)

Ali Selamat (Universiti Teknologi Malaysia, Malaysia)

Buana Ma’ruf (Badan Pengkajian dan Penerapan Teknologi, Indonesia)

Carlos Guedes Soares (University of Lisbon, Portugal)

Cho Myung Hyun (Kiswire Ltd, Korea)

Dani Harmanto (University of Derby, UK)

Harifuddin (DNV, Batam, Indonesia)

Hassan Abyn (Persian Gulf University, Iran)

Hassan Ghassemi (Amirkabir University of Technology, Iran)

Iis Sopyan (International Islamic University Malaysia, Malaysia)

Jamasri (Universitas Gadjah Mada, Indonesia)

Mazlan Abdul Wahid (Universiti Teknologi Malaysia, Malaysia)

Mohamed Kotb (Alexandria University, Egypt)

Moh Hafidz Efendy (PT McDermott, Indonesia)

Mohd. Shariff bin Ammoo (Universiti Teknologi Malaysia, Malaysia)

Mohd Yazid bin Yahya (Universiti Teknologi Malaysia, Malaysia)

Mohd Zaidi Jaafar (Universiti Teknologi Malaysia, Malaysia)

Musa Mailah (Universiti Teknologi Malaysia, Malaysia)

Priyono Sutikno (Institut Teknologi Bandung, Indonesia)

Sergey Antonenko (Far Eastern Federal University, Russia)

Sunaryo (Universitas Indonesia, Indonesia)

Sutopo (PT Saipem, Indonesia)

Tay Cho Jui (National University of Singapore, Singapore)

Journal of Ocean, Mechanical and Aerospace -Science and Engineering-, Vol.44

June 30, 2017

1 JOMAse | Received: 8 - April - 2017 | Accepted: 30 - June - 2017 | [(44) 1: 1-7] Published by International Society of Ocean, Mechanical and Aerospace Scientists and Engineers, www.isomase.org., ISSN: 2354-7065 & e-ISSN: 2527-6085

Characteristics of Dynamic Response of Suspension Hydraulic Motor - Regenerative Shock Absorber (HMRSA)

Kaspul Anuar a,* and Harus Laksana Guntur b

a) Mechanical Engineering, Universitas Riau, Indonesia b) Mechanical Engineering, Institut Teknologi Sepuluh November, Indonesia *Corresponding author: [email protected] Paper History Received: 8 - April - 2017 Received in revised form: 28 - April - 2017 Accepted: 30 - June - 2017 ABSTRACT Translational motion that happens in the vehicle's suspension due to unevenness of the road surface can be used as a source of electrical energy. A suspension that can convert translational motion into electrical energy is known as regenerative type suspension. To know the characteristics of the dynamic responds, such as electrical energy potential and driving coziness which resulted by a suspension system, we need to examine the suspension. In this study, a test will be conducted to a suspension system that have been designed by researchers and was named Hydraulic Motor - Regenerative Shock Absorber (HMRSA). The test will be conducted statically and dynamically. The goal of the static testing is to obtain the spring’s constant value and the damping’s constant value of HMRSA. In the dynamic testing, excitation was given in the form of periodic and impulse. Periodic excitations are varies between these several frequencies such as 1.4Hz, 1.75 Hz and 2 Hz. Instead of variant of frequencies, electrical resistivity loads are varies in periodic excitations with each resistive loads such as 6 ohm, 12 ohm and 18 ohm. From dynamic testing, the electricity power values and sprung’s mass acceleration which resulted by HMRSA suspension system on each frequencies and electrical resistivity will be obtained. The sprung’s mass acceleration value will be fundamental on how to analyze driving coziness that produced by HMRSA suspension system. KEY WORDS: Hydraulic Motor - Regenerative Shock Absorber (HMRSA), sprung’s mass, acceleration.

NOMENCLATURE

K spring’s constant (N/m) C damping’s constants (Ns/m) m vehicle mass (kg) x displacement (m) �� velocity (m/s) �� acceleration (m/s2) Te electrical torque (N.m) F force (N) N amount of coil turns B Flux density vector 1.0 INTRODUCTION

Realizing the amount of energy wasted on vehicles [1], a device to harvest those wasted energies is needed. By this, the consumption of fuels on vehicles will be saved in such a way. One of a mechanism that be able to relieve energy loss in vehicles is suspension that known as regenerative shock absorber (RSA). Regenerative shock absorber can change the vibrations on vehicles that happened by unevenness of the road surfaces into electrical energy [2].

There are several researches about RSA that have been conducted by some researchers. In 2009, a team from Massachusetts Institute of Technology [3], have developed a regenerative shock absorber system that can produce electricity based on hydraulic principles. Inside this regenerative shock absorber there is a piston that used to push fluid that flows to turbine, after that the spin of the turbine will also spin the generator. This regenerative shock absorber suspension could produce a power of 1 KW on normal roads.

In 2010, Zuo [1] from Stony Brook University design and develop a device to collect energy on a vehicle’s suspension. The wasted vibrations on the suspension are used to produce electricity. This device was developed with 2 different principles of electricity generator, linear electromagnetic absorber and


June 30, 2017


rotational absorber. The power that produced by linear electromagnetic is about 2 - 8 watt, while rotational absorber could produce a power of 80 watt [1].

Arziti [2] develop a regenerative shock absorber that based on piezoelectric. This suspension uses a piezoelectric at the top of the piston as its energy generator. In this regenerative shock absorber, the electrical voltage generated is 0.0176 volts. Next, fang [3] from Wuhan University of technology China, create an absorber from actuator hydraulic component which is connected with the hydraulic generator motor. Based on the test, the energy generated from this regenerative shock absorber reaches 193 watts.

In Indonesia, The development of regenerative type suspension is done by Guntur, L. Harus [4]. This regenerative type suspension converts translational motion into rotational motion to rotate generator. The development of this RSA type rotational absorber is capable of generating power of 15 -18.6 watts. 2.0 FUNDAMENTAL THEORY 2.1 Motion of Base The vibrations in the vehicle suspensions can be modeled as shown in Figure 1. below.

M

K C

Y(t)

X

Base

mX

Figure 1: Base excitation (a) Physical system base excitation (b) Free body diagram for base excitation system [5]

As shown in figure 1(a), The mass-spring-damping system

undergoes harmonic motion. Excitation input y(t) state displacement from base, and x(t) state mass displacement from the static equilibrium position at time t. Then the elongation of the spring is �� and The relative speed between the two ends of the damper is�� . From free body diagram shown in figure 1 (b), we obtain the equation of motion :

�� (1) The force received by the spring mass due to the base exciter can be formulated as follows:

� �� sin�� ∅� �� sin�� ∅� (2) 2.2 Spring Spring is a mechanical connector which widely used in many applicants. Spring is assumed to have no mass and no damping. The type of spring commonly used in a vehicle is a helical coil spring. Any object that has elasticity can be assumed to be a

spring [5], figure 2 shows springs that was charged to compression loads and tensile loads.

Figure 2: Springs that was charged to compression loads and tensile loads [5]

In accordance with the 3rd Newton’s law, if a spring is given a tensile force or compression force then the direction of the spring’s reaction force will be opposite. The spring is called linear if the relation between force and spring deflection satisfies the following equation [3]:

F = k. x (3)

If the relations between the force given and the spring deflection do not satisfy the equation (3), then the spring is called non-linear. Figure 3 shows graph of force vs deflection on non linear springs

Figure 3: Graph of force vs deflection on non linear springs [5]

On non linear springs, the relations between force gain and deflection gain, satisfies the following equation:

∆� �. ∆� (4)

On non linear springs [5], the spring’s Constants can be

linierized by using the following equation:

� ��

�� (5)

2.3 Damping Damping is part of the vibration system that converts vibration energy into heat and sound [5]. Damping is assumed to have no


June 30, 2017


masses and elasticity. In non-linear damping [5], the damping Constanta can be dilinierized by using the following equation:

� ��

�� (6)

2.4 Electric generator A generator is a device that converts mechanical energy into electrical energy. The electric generator induces electrical motion by rotating the coils in a magnetic field. Figure 4 shows Schematic of electric generator

Figure 4: Schematic of electric generator [6]

On the generator, the electric torque generated can be determined using the following equation [6]:

� 2.". #. $. %. & (7)

2.5 The effect of Vehicle Acceleration towards coziness. Information about human body resistance to acceleration is very important as a reference in the design of vehicle body resistance to impact. The comfort criteria based on acceleration according to ISO 2631 standard, is shown in table 1 below:

Tabel 1: Comfort reaction to acceleration – ISO 2631 No. Acceleration (RMS) Details 1. a < 0,315 m/s2 No complaints 2. 0,315 m/s2 to 0,63 m/s2 A bit uncomfortable 3. 0,5 m/s2 to 1 m/s2 Somewhat uncomfortable 4. 0,8 m/s2 to1,6 m/s2 Uncomfortable 5. 1,25 m/s2 to 2,5 m/s2 Very uncomfortable 6. a > 2 m/s2 Very very uncomfortable

3.0 METHODOLOGY 3.1 Static Testing This study begins by measuring the value of the damping’s constants and the spring’s constants of the HMRSA suspension system. The measurement of spring’s constant value is done by giving seven variations of mass load between 217.9 kg up to 277.9 kg above test rig test equipment. From the load we will obtain spring deflection (∆x) HMRSA suspension. Here is shown figure 5 testing the spring defelection value with the rig test equipment.

Figure 5: Testing the spring deflection value

The value of spring’s Constants is obtained by doing calculations according to the equation (4) and (5).

The test to obtain the value of damping’s constants is done by giving 3 variations of mass load on HMRSA suspension. The result will be the speed of vertical motion of the absorber when subjected to mass load. Here is shown figure 6 Scheme about testing the speed of vertical absorber motion when subjected to mass load.

Figure 6: Static testing mechanism of compression damping value

The value of damping constants is obtained by calculating according to the equation (6) 3.2 Dynamic Testing This test is done to obtain dynamic response of HMRSA suspension in the form of spring’s mass acceleration and electric power generated. The testing mechanism being used quarter car’s model. In this test the mass of concrete is assumed as a spring’s mass (mass on the vehicle), while on the bottom plate (excitation sources) used as base exciter.

The input on the dynamic testing is excitation impulse and harmonic with amplitude of 1.5 cm and varies of excitation frequency of 1.4 Hz, 1.75 Hz and 2 Hz. The testing load being used is 250 kg (quarter of urban vehicles weight). Here is shown figure 7, Dynamic testing of HMRSA suspension.


June 30, 2017


Figure 7: Dynamic testing of HMRSA suspension

Details of figure 7: 1. Installation of HMRSA on Suspension test rig; 2. Test load (250 kg) 3. Hydraulic motor, gear transmission and HMRSA generator 4. Lamp installation as a load 5. Accelero sensor on spring mass 6. Accelero sensor on base exciter 7. Rectifier (AC to DC rectifier circuit) 8. Limit switch (excitation amplitude regulator) 9. Control Panel

4.0 RESULTS AND DISCUSSION 4.1 Static test results The spring deflection data retrieval process is done using seven variations. Here is shown figure 8 Data of spring deflection test results.

Figure 8: Data of spring’s test results

Results of spring deflection test on figure 8 then ploth on graphical of force vs deflection when there is compression and rebound from spring. Figure 9 shows graphical plot of force vs spring deflection as test results.

Figure 9: graphical plot of force vs spring deflection

From the graph above, Graph’s gradient of force vs deflection experienced a slight increase on 2628.1 N to 2726.2 N loads. This indicates the beginning of nonlinearity from spring’s constants on that load. According to the equation (5), the value of spring is constants when compression and rebound are 44564 N/m and 44151 N/m respectively.

The tests to find the damping value of HMRSA were done by giving forces for compression. Three kinds of loads were given on the test and repeated as much as three times. Below shown figure 10 data of HMRSA damping test results

Figure 10: Data of HMRSA damping test results

From the test result data, then ploth a graph force vs velocity so that the total damping compression value will be obtained from HMRSA suspension. Here is shown figure 11. Graph force vs velocity test results.


June 30, 2017


Figure 11: graphical plot force vs velocity

On figure 11 above gradient of the graph damping force (Fd) vs velocity (v) is directly proportional, the bigger the force given, the higher the velocity. This is in accordance with the equation (6) where c is the gradient or the damping constant of HMRSA suspension. From the graph above given the gradient or HMRSA suspension damping value equal to 10492 N.s/m. 4.2 Dynamic test results Excitation impulses are given only on amplitude 1.5 cm with varies of electrical resistance load on the generator of 6 ohm, 12 ohm and 18 ohm. The results of tests with impulse excitations are shown in Figure 12 below.

Figure 12: Graph of results of the mass spring acceleration responses test due to impulse excitation with varies of electrical loads of 6 ohms, 12 ohms and 18 ohms

From figure 12 above, we can see the difference of response

on sprung’s mass on each loading. The smallest peak amplitude of sprung’s mass occurs in the 6 ohm electrical resistance with value 0.95 m/s2. On the 12 Ohm load, the peak amplitude is bigger than the 18 ohm loading of 1.55 m/s2. And the highest peak amplitude value occurs at 18 ohm power loading which equal to 1.605 m/s2. This shows that the smallest damping value occurs on HMRSA that used 18 ohm load which create big peak amplitude. From the figure above also seen, the time it takes to reach stable condition on every load is almost the same which about 0.3 seconds.

Varies of electrical resistance values will affect the strong current generator generated. On the test with bigger electrical resistance will cause the current generated become smaller. If the generated current generator is smaller it will cause electric torque values (electrical damps) also small. This corresponds to the equation (7). With the shrinking value of electrical damps

(electrical torque) will decrease HMRSA damping coefficient so that the peak amplitude value of the spring mass acceleration will get bigger.

Here is shown figure 13 graph of spring mass acceleration response due to periodic excitation on 1.4 Hz frequency and amplitude of 1.5 cm.

Figure 13: Test results graph of sprung’s mass of periodic excitation response with load of 6 ohm, 12 ohm and 18 ohm on frequency of 1.4 Hz

In the test using excitation frequency as much as 1.4 Hz, the highest root-mean square (RMS) of sprung’s acceleration occurred at 18 ohm loading with value 0.402m/s2. On 12 ohm load, the RMS value of sprung’s mass acceleration as much as 0.387 m/s2. The smallest RMS of sprung’s mass acceleration occurs at 6 ohm load which is 0.248 m/s2. This shows that the smallest damping value occurs in HMRSA using 18 ohms resistance so that the RMS of sprung’s mass acceleration is the greatest.

According to ISO 2361, the three sprung’s mass acceleration responses have an RMS value below 0.8m/s2, This indicates that HMRSA is able to maintain driving comfort in periodic excitation with frequency 1.4 Hz and amplitude 1.5 cm.

At periodic excitation frequency of 1.75 Hz, the load of electrical resistance is also varied by 6 ohm, 12 ohm and 18 ohm. Here is shown figure 14 graph of sprung’s mass acceleration response due to periodic excitation at 1.75 Hz and amplitude 1.5 cm.

Figure 14: Test results graph of Sprung’s mass response of periodic excitation at 6 ohm, 12 ohm and 18 ohm loading at 1.75 Hz frequency

C = 10492x

Fo

rce

(N

)

Velocity (m/s)

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7-1.5

-1

-0.5

0

0.5

1

1.5

2

Time (s)

Acc

eler

atio

n (

m/s

2)

6 ohm12 ohm18 ohm

0 0.5 1 1.5 2 2.5-1

0

1

2

3

Time (s)

Acc

ele

ratio

n (

m/s

2)

6 ohm12 ohm18 ohm

0 0.5 1 1.5 2 2.5-2

-1

0

1

2

3

Time (s)

Acc

eler

atio

n (m

/s2)

6 ohm12 ohm18 ohm


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From figure 14 it can be seen that the difference of sprung’s mass acceleration response at each loading is not so much because the variation of load of electrical resistance being used also not so much difference. In the test using the excitation frequency of 1.75 Hz, the greatest RMS of sprung’s acceleration occurred at 18 ohm loading with value 0.729 m/s2, followed by 12 ohm loading of 0.5979 m/s2 and the smallest RMS occurs at 6 ohm loads with value 0.5534 m/s2. This also shows that the smallest damping value occurs in HMRSA using 18 ohm resistance so that the rms mass spring acceleration will be the greatest.

According to ISO 2361, the three spring mass acceleration responses have an RMS value below 0.8m/s2. This indicates that HMRSA is able to maintain driving comfort in periodic excitation with frequency 1.75 Hz.

Periodic excitation at a frequency of 2 Hz varies the load of electrical resistance by 6 ohms, 12 ohms and 18 ohms. Here is shown figure 15 graph of sprung’s mass acceleration response due to periodic excitation at 2 Hz frequency and amplitude 1.5 cm.

Figure 15: Test results graph of sprung’s mass of periodic excitation response with load of 6 ohm, 12 ohm and 18 ohm on frequency of 2 Hz

In the test using the excitation frequency of 2 Hz, The largest RMS of sprung’s mass acceleration occurs at 18 ohm loading with value of 1.02 m/s2. On 12 ohm load, the RMS value of sprung’s mass acceleration is 1.009 m/s2. RMS acceleration of the smallest sprung’s mass occurs at 6 ohm electrical load which equal to 0.6383 m/s2.

According to ISO 2361 the sprung’s mass acceleration response at 12 ohm and 18 ohm electrical resistance has an RMS value of more than 0.8 m/s2. This indicates that HMRSA is not able to maintain the comfort of driving when working at a frequency of 2 Hz. 4.3 The Power Generated by HMRSA From the above periodic excitation testing, we obtained the amount of energy generated at each excitation frequency. Here is shown table 2 Electric power generated HMRSA suspension.

Table 2: HMRSA Generated Electric Power

Loads Power generated (watt)

1,4 Hz 1,7 Hz 2Hz

6 ohm 0.198 0.063 0.024

12 ohm 0.6771 0.623 0.3465

18 ohm 1.1078 0.5157 0.4932

From table 2, then displayed in the form of a bar chart as shown in Figure 16. below.

Figure 16: Electric power generated by HMRSA

From figure 16 above, it is known that the electric power

generated HMRSA suspension is very small that is ranging from 0.024 watts to 1.1078 watts. The small electric power generated by HMRSA suspension is due to the small excitation amplitude being used is 1.5 cm. Furthermore, the electric generator rotation of the HMRSA suspension is relatively slow. This is inversely proportional to the specification of HMRSA’s generator which is high-speed type power generator.

Viewed from the increasing frequency side, generally the electric power generated by HMRSA suspension keeps increasing as frequency increases. The highest power produced by HMRSA occurs at 2 Hz excitation frequency and 6 ohm electrical resistance which is 1.1078 watts. 5.0 CONCLUSION In this research, HMRSA suspension testing is done statically and dynamically. Static test aims to determine the value of spring and damping constants. While the dynamic testing aims are to determine the electric power and acceleration of spring mass produced HMRSA suspension. The following describes the conclusions of this study. 1. HMRSA uspension is able to maintain the driving comfort in

periodic excitation of frequency 1.4 and Hz 1.7 Hz on all variations of electrical resistance. Whils at the frequency of 2 Hz, HMRSA suspension only able to maintain driving comfort at 6 ohm electrical resistance.

2. There is an increasing trend of RMS value of sprung’s mass acceleration along with increasing of electrical resistance value.

3. The power generated by HMRSA suspension is very small which ranges from 0.024 watt to 1.078 watts. This is because the excitation amplitude being used is only 1.5 cm.

0 0.5 1 1.5 2 2.5-2

-1

0

1

2

3

4

Time (s)

Acc

ele

ratio

n (m

/s2

)

6 ohm12 ohm18 ohm


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REFERENCE 1. Zuo, Lei, et.al, 2010. Design and Characterization Of An

Electromagnetic Energy Harvester For Vehicle Suspension. New York State University, USA.

2. Arziti, Marcos. 2010. Harvesting Energy From Vehicle Suspension. Tempere University of Technology: Spanyol.

3. Fang, et.al, 2013.Experimental Study of Damping and Energy Regeneration Characteristics of a Hydraulic Electromagnetic Shock Absorber.Wuhan University of Technology.China.

4. Guntur, L. Harus. 2013. Development and Analysis of a Regenerative Shock Absorber for Vehicle Suspension. JSME Journal of System Design and Dynamics.

5. Rao, S Singiresu. 2011. Mechanical Vibration. Prentice Hall PTR: Singapore.

6. Frederick, Close. 1995. Modeling and Analysis of Dynamic System. John Wiley &Sons : USA

7. Kadaryono. 2013. Pengembangan dan Studi Karakteristik Prototipe Regenerative Shock Absorber Sistem Hidrolik. Institut Teknologi Sepuluh Nopember, Surabaya.


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8 JOMAse | Received: 12-April-2017 | Accepted: 30-June-2017 | [(44) 1: 8-20] Published by International Society of Ocean, Mechanical and Aerospace Scientists and Engineers, www.isomase.org., ISSN: 2354-7065 & e-ISSN: 2527-6085

Analyze Performance of Double Acting Tanker While Running Astern in Ice Condition

Efi Afrizal,a and Jaswar Koto,a,b,*

a)Faculty of Mechanical Engineering, Universiti Teknologi Malaysia, Skudai, Johor, Malaysia b)Ocean and Aerospace Research Institute, Indonesia *Corresponding author: [email protected] & [email protected] Paper History Received: 12 - April - 2017 Received in revised form: 25 -June - 2017 Accepted: 30 - June - 2017

ABSTRACT The increasing of shipping activities through the Northern Sea Route (NSR) and growth of oil and gas activities in Arctic and Sub-Artic regions require suitable design of ice-going ships and planning operations in ice. In 2002, Sumitomo Heavy Industries has built advanced ice-ship called “Double Acting Tanker”. This paper discussed application of new method to determine ice resistance of Double Acting Tanker running ahead in ice condition. The simulation was carried out at 1 m ice thickness in unfrozen and frozen channels and 0.5 m ice thickness in level ice condition. The simulation results were compared with experimental results. KEY WORDS: Running Ahead; Ice Thickness; Double Acting Tanke. NOMENCLATURE AAT Aker Artic Technology DAT Double Acting Tanker DWT Deadweight MW Mega Watt NSR Northern Sea Route

1.0 INTRODUCTION Ice–going ships have been developed called as Double Acting Tanker (DAT) which can be travel more efficient in astern than ahead at ice conditions as shown in Figure 1 (Juurmaa et al. 2002). A lot of researches have been developed to find the optimum hull design of double acting tanker while operating in sea ice as astern mode. Recent development is by optimization diesel-electric power plan concept combine with an azipod on the propulsion system of DAT. Sasaki et al. (2004) reported experimental result at the full-scale Double Acting Tanker "Mastera" and "Tempera" with 106000 DWT of weight and 16 MW of powering. The experiment had been done at Sagami Bay, Japan for Mastera and at route between Porvoo to Primorsk, Rusia for Tempera. Improvement on performance was obvious when ship could be sailing at the astern mode in the frozen seas where it does not need escort anymore by icebreaker ship.

Figure 1: Double Acting Tanker in ice condition (Juurmaa et al, 2002)

Based on previous findings, the special design was required

for ships to be operated in open water and ice conditions. The


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phenomenon of interaction between ice and ship has been carried out by researchers through empirical mathematical simulation such as Chen and Lee (2003), Lee (2006), Islam, Veitch and Liu (2007) and Tan et al. (2013, 2014). In the phenomenon, there are two forces acting at the same time that compressed by the hull and sucked by the propeller. Jaswar (2005) has developed an empirical mathematical model to predict resistance of Double Acting Tanker (DAT) without taking into account the impact of suction force caused by the propeller as the ship walked toward the rear. This paper discusses mathematical model to predict the strength of the suction force caused by the propeller of DAT during sailing astern.

The uniqueness concept of double acting ship is that could be operated ahead mode if ship was sailing in the open water or to be operated astern mode when the ocean was covered by ice. The performance could be achieved because the DAT ship is using a podded propulsion system which has ability to rotate 360° freely on its axis. The podded propulsion system so that can always be in a state of pulling even when the ship is sailing at ahead or astern modes. “Mastera and Tempera” are examples of ships which were developed with this concept. They had been operating since 2002 and 2003 (Sasaki et al. 2004). Below of this, it would be discussed fundamental concept of Double Acting Tanker and application of proposed method called “Efi-Koto Method” on Double Acting Tanker. 2.0 ICE RESISTANCE WORKING ON SHIP SAILING IN ICE CONDITION As reported by Jones (2004) in the book His review, below some of those involved in this study will be rewritten again to make clear the double acting ship concept. Significant contribution begin by Jansson (1956[a] and 1956[b]). He discussed in detail the history of icebreaking ship from what he considered the earliest true icebreaker, Eisbrecher 1. The ice–breaker was operated between Hamburg and Cuxhafen, it was built in 1871 and in 1956 it was began to use bow propeller while penetrated on ice. Jansson also discussed the science of icebreaking. He quoted, values for the physical properties of freshwater ice, at -

3o

C, as shown in Table 2.1:

Table 2.1 Properties of Freshwater Ice (Jones, 2004) Elastic Modulus 70,000 kg/cm2

(6,900 MPa) Tensile and bending strength 15 kg/cm2

(1.5 MPa) Compressive strength 30 kg/cm2

(2.9 MPa) Shear strength 7 kg/cm2

(0.7 MPa)

There was not mentioned of details experiments including that value, some addition information were only for coefficient of friction between ice and metal as 0.10 to 0.15 for fresh or Baltic ice and 0.20 for salt water or polar ice. He gave a simple formula for the total ice resistance as described in Equation (2.1): �� = �� . ℎ + ��. ℎ. �� . � (2.1)

Where; C� and C� are experimental constants, h is ice thickness, v

is vessel speed and B is breadth of vessel at waterline. After that, Jones (2004) in his report said credited to

Kashteljan et al. (1968) whom the first detailed attempt to analyse level ice resistance by breaking it down into components. Where on the paper, it was appeared like an Equation (2.2) to determine the total of ice resistance, ��: �� = �� !�"ℎ + �� !�#ℎ� + �$ 1&� �'(�') (2.2)

Where; " is ice strength, � is ship beam, ℎ is ice thickness, � is ship speed, and # is the density of ice. ! and η� are related to Shimansky’s ice cutting parameters, and �� , �� , �$ , �* , �+ are coefficients experimentally determined (0.004, 3.6, 0.25, 1.65, and 1.0 respectively).

In the Equation (2.2), that compose of several parts like, first component R

1 = !�"ℎ is resistance due to breaking the ice,

second component represented of R2

= �� !�#ℎ�, is resistance

due to forces connected with weight (such as submersion of broken ice, turning of broken ice, change of position of icebreaker, and dry friction resistance) and third is component of $ = �$ �,- �'(�') for determined of resistance due to passage

through broken ice Lewis and Edwards (1970) gave a good review of previous

work and derived the Equation (2.3); . = �!"ℎ� + ��#/�ℎ� + ��#�ℎ�� (2.3)

Where; . = mean resistance excluding water / = acceleration due to gravity �0, ��, �� =

non-dimensional coefficients to be determined experimentally

The first term represents ice breaking and friction, the second

accounts for all resistance forces attributable to ice buoyancy, and the third accounts for all resistance forces attributable to momentum interchange between the ship and the broken ice. They conducted non-dimensional analysis by dividing by "ℎ� to obtain the Equation (2.4): 2 = �0 + ��234 + ��235 (2.4)

Where;

′ = . "⁄ , non-dimensional mean ice resistance

�′= � ℎ⁄ , non-dimensional beam

3Δ = #/ℎ⁄ , volume metric number

35 = #" �⁄ , inertial number

Crago et al. (1971) describe a set of model test in “wax-type”

ice on 11 icebreakers. By considering a simple bow geometry and the vertical force acting on the ice sheet, they derived Equation (2.5) for the ice thickness, ℎ ;


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ℎ√89:; = 1.539tan �A + B� (2.5)

Where; 8 = ice tensile strength :; = Thrust A = stem angle B = tanC� D ; D is the coefficient of friction

Enkvist (1972) made a major addition to the literature of ship performance in level ice. On the article narrated of experimental on model tests for three ships; Moskva-class, Finncarrier, and Jelppari, and was able to compare his results with limited full-scale data from all three. From a combination of analytical work, dimensional analysis, and a few assumptions, they derived a semi–empirical Equation 2.6 where defined ice resistance based on three terms: �� = ��ℎ" + ��ℎ:#∆/ + �$�ℎ#�� (2.6)

Where; : = draft of ship #F = density of water # = Density of ice #∆ = #F G #

Milano (1973) made a significant advance in the purely theoretical prediction of ship performance on ice. He considered the energy needed for a ship to move through level ice, which varied somewhat with ice thickness. For example, for very thick ice the ship moves through the ice-filled channel �H�� , impacts the various bow and cusp wedges causing local crushing �H��, climbs onto the ice �H$� until sufficient force is generated to cause fracture, at which time the ship falls �H*� , and moves forward, forcing the ice downward �H+�. The total energy loss due to ship motion can be calculated using Equation (2.7); H� = H� + H� + H$ + H* + H+ (2.7)

Vance (1975) obtained an “optimum regression equation”

from five sets of model and full-scale data, of the Mackinaw same data as used by Edwards et al. (1972), Moskza, Finncarrier, Staten Island, and Ermak. Equation (2.8) was resulted by regression to define ice resistance: �� = �I#∆/�ℎ� + �J"�ℎ + �K#L�Mℎ0.N+�0.$+ (2.8)

Where; �� is the resistance due to ice, M is length of vessel, and �I , �J , �K

are empirically determined values. The first term

is a submergence term, the second a breaking term, and the third term is a velocity dependent resistance.

An example of a fit to his equation is shown in Figure 4.1 in which the Mackinaw full-scale data (label FS) are shown fitted to his equation above (label FSR) and a model-scale regression to his equation (MSR) is also shown. Good agreement is found between the model and full-scale results.

Figure 2.1 Result using optimum Regression (Vance, 1975)

Edwards et al. (1976) presented full-scale data for the Louis

S. St. Laurent collected by analysing ramming type tests using Equation (2.9). The equation is in non-dimensional form. #F/�ℎ� = 4.24 + 0.05 "#F/ℎ + 8.9 L9/ℎ (4.9)

Kotras et al. (1983) in his paper proposed an equation based

on Nagle’s thesis (Nagle, unpublished) which describe yet by another semi-empirical approach. In his proposed equation the total ice resistance is given by �� = J + JT + � + �T + I + IT (2.10)

Where ; �� = total ship ice resistance

J, JT = normal and frictional resistance due to breaking of level ice

� , �T = normal and frictional resistance due to broken ice floes

I, IT = normal and frictional resistance due to submerging broken ice

Since 1985, development of new icebreaking forms has been

having significant value, more scientific approach had been used such as modelling of ships in ice with extensive model testing and, most recently, numerical methods. Canadian Arctic oil exploration and development led to new designs such as the Kigoriak, and Terry Fox, while other activities led to the Oden, double acting tankers (DAT) with Azipods, FPSO’s in ice, and research ships such as the Nathaniel B. Palmer, USCGC Healy, and the converted CCGS Franklin now called CCGS Amundsen.

An interesting development in the mid-80’s, Zhan et al. (1987) was made a full-scale resistance trial of the Mobile Bay in uniform level ice. Denny (1951) had been reported the same term, in the principle the experiment method is parallels with the open water trials of the Greyhound (Froude, 1874) and Lucy Ashton.


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While such tests are clearly difficult to perform, in theory they provide a direct measurement of full-scale resistance. They also conducted full scale propulsion tests. They found the best fit to their towed resistance results was with the Equation (2.11): ��#F/�ℎ� = �0 + �� /� . M$ℎ (2.11)

Where; �0 = 4.25

��= 3.96V10C+

Lindqvist (1989) had included submersion component in the

Equation (4.12) to determine ice resistance working. Based on his observation from the full scale experimental, he conclude that ice would be fractured in the one continue cycle including rotating and sliding of broken ice floes.

= W#. /. ℎ X: � + :� + 2:+ Y0.7 M G :tan ∅ G �4 tan \+ : cos ∅. cos ` a 1sin� ∅ + 1tan� \cd . e1+ 9.4 �9/. Mf

(2.12) Where W# is the density difference between the water and the ice, / is the acceleration of gravity, ℎ is ice thickness, M , �, and : are the length, breadth and draft of the ship, is the frictional coefficient, ∅ is the stem angle, \ is the waterline entrance angle, � is the ship speed in ice and ` is the angle between the normal of the hull surface and the vertical vector and can be define by ;

=̀ ghijgk tan ∅sin \

Tan et al. (2013) has rearranged formula of Lindqvist and

present coefficients that were applied to represent each of step on the ice breaking including crushing, braking and submersible. That is showed in Equation (2.13): lm�� = inℎ� + eon + 1.4�n9/ �mf ℎ�.+

+ epn + e1.4on9/ + 9.4pn9/MFqf �mf ℎ (2.13)

3.0 HULL RESISTANCE IN OPEN WATER

The performance of a ship related to resistance loading at the ship. Total resistance, � at the open water is the force required to make the ship traveling in the certain speed. Resistance working at the ship consists of several components. Commonly, general measured are pressure resistance and friction resistance r (Frisk and Tegelhall, 2015). Other components often used in analysis performance of ship are viscous resistance, K and wave resistance, F . If vessel traveling through open water, some resistance can be in form of shear force is known as viscous resistance. In addition, the resistance also could be found in a form of wave resistance due to water wave. Besides, other resistance could be working on the up side of vessel caused by air specified as minor resistance. The total resistance of ship in open water can be expressed as Equation (3.1). � = r + K + s (3.1)

The total resistance coefficient, �� is a dimensionless quantity

which can be defined as Equation (3.2). This coefficient used to characterize total resistance in the different hulls. �� = �� # t� uF (3.2)

Where; � is the total resistance, # is density of water, t is hull speed and uF is the wetted surface area. 4.0 ICE SHIP SAILING IN ICE CONDITION

The first model of ice basin was built in the Soviet Union by AARI, 1955. That was needed to observe either of hull form or propulsion could be effect to ice–breaking ship performance. In ice model test, Froude scaling law is using to associate ice model test and full scale situation. Wilkman (2015) revealed that total resistance is the summation of ice resistance and open water resistance. Ice resistance is an amount of resistance to breaking ice, resistance of some component sink ice under hull and resistance velocity due to dynamic working. Experiment in ice model scale test can be contributed to reducing huge investment before the real ship was manufactured.

Performance ship on ice was measurable in capability of ship to break ice and to manoeuvre in ice condition. That could be confirmed through achieved speed by ship when sailing in uniform or certain ice thickness, ice ridges or in the level ice condition (Wilcox, 1994). The velocity of ship on ice condition can be determined through thrust of propeller available to overcome the ice resistance. Performance of propulsion system can be improved thru modification on hull shape and some change into propulsion design, both of that could minimize an effect of resisting forces and maximize the propulsive forces. The ice resistance is assuming linear to ship speed which composed of three components, like described in Equation (4.1): �� = v + w + T (4.1)


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Each component v , w and T , successively are breaking, submersion and friction components. The breaking component is related to break the ice such as crushing, bending and turning of ice. The submersion component is concerned to push the broken ice down along the ship hull. The friction component is connected to slide the broken ice along the ship hull. In general velocity of the ship depends on working ice resistance associated to friction component. The total resistance working on ice, ;!;xq is the sum of ice resistance, �� and open water resistance, !F , as

expressed in Equation (4.2): ;!;xq = �� + !F (4.2)

Figure 4.1 describe interaction happening between hull and

ice including crushing, bending, submersion and friction of ice at bow hull.

Figure 4.1 Hull and ice interaction (Wilcox, 1994) 4.1 Ice-Ship Sailing Ahead in Ice Condition

4.1.1. Ice-Ship Sailing Ahead in Ice Condition – Unfrozen Channel In this condition, the ship resistance is defined as the average value of all longitudinal forces caused by ice acting on the structures. stated that total resistance to ship’s continuous motion through ice-infested waters is expressed as the summation of resistance developed in failing the ice, the momentum interchange between the ship and the ice, ice buoyancy forces and the open water resistance as in Equation (4.3) (Jaswar 2005):

lwnyCz{| = lFx;�}CT}�; + lw~v + lT}�;+ l.!.��; (4.3)

Where; the above equation is summation that consists of all longitudinal forces as stated below. a. Water Friction Resistance (��C��) To determine the frictional resistance coefficient, William Froude’s approach yielded that the frictional resistance coefficient was related to the resistance coefficient of plate with the same length and wetted surface area of the ship or model hull, which can express in Equation (4.4) and Equation (4.5):

�T = T�� ρ L� � (4.4)

Or:

lFx;�}CT}�; = lFTz�~r� = 12 �FTz�~r� ρ L� � (4.5)

Where:

lFTz�~r� is head unfrozen water friction resistance

�FTz�~r� is head unfrozen water friction resistance coefficient

T is frictional resistance (N) ρ is density of water (kg/m3) L is ship or model speed (m/s)

� is wetted surface of ship or model hull (m2)

Several friction lines based only on the Reynolds number

were developed later, both theoretically using boundary layer theory and experimentally. For laminar flows, the resistance coefficient was formulated from boundary layer theory by Blasius, as shown in Equation (4.6): �T = 1.328. √k (4.6)

So-called plate lines were developed for turbulent boundary layer flows from the leading edge. These plate lines were extended to include full scale Reynolds numbers. The formulations, such as the Schoenherr Mean Line or the ITTC-1957 Line (Molland et al. 2011), which are determined, as indicated in Equation (4.7) and Equation (4.8):

�iℎ��kℎ�hh ∶ 0.2429�T = ��/�0�k. �T�

(4.7)


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�::� G 1957 ∶ �T = 0.075�M�/�0�k� G 2��

(4.8)

The latter one is accepted as a standard by the International

Towing Tank Conference (ITTC). As a matter of fact, it is not too important that a flat plate with a certain length and wetted surface has a resistance coefficient exactly according to one of the mentioned lines. The Froude hypothesis is crude and correlation factors are required afterward to arrive at correct extrapolations to full-scale values. These correlation factors will depend on the plate line which is used. b. Submersion Resistance (��)

The submersion resistance is assumed arise from work required to tip and submerge the broken an ice cusps. The submersion resistance depends on buoyancy of force of the ice cusp due to different density between the ice cusp and seawater, as expressed in Equation (4.9):

lw~v = lw~vz�~r� = �wz�~r� . �#Fx;�}G #��. /. �. �. ℎ (4.9)

Where:

lw~vz�~r� is head (unfrozen) submersible resistance

�wz�~r� is head unfrozen submersion coefficient

#Fx;�} is water density 1.025 ton/m3 #�� is ice density 0.918 ton/m3

/ is acceleration of gravity 9.81 m/s2

� is depth of ice cusp � is width of the ice cusp c. Friction Resistance (��) The frictional resistance is found when buoyancy force of the broken ice is against the hull and underside of the broken ice field as well as the effect of hull form such as friction between ice and hull and broken ice piece and under surface of the broken ice cover. The frictional resistance can be expressed as Equation (4.10):

lT}�; = lTz�~r�= �Tz�~r�. #�� . /. h. ℎ. �. L 9M. /� . D�g, o, �F� (4.10)

Where:

lTz�~r� is head unfrozen friction resistance

�Tz�~r� is head unfrozen friction coefficient

M is ship length

�F is water plane area coefficient of entrance part

L is speed of ship c. Momentum Resistance (��)

Loss momentum resistance is developed when resistive force attributable to extract momentum from the ship and imparting it to broken ice pieces. The time rate of change momentum of the ship is equal to resultant force on ship, which can be expressed as Equation (4.11):

l.!.��; = l.z�~r�= �.z�~r�. #�� . �. ℎ. L�. D�g, o� (4.11)

Where:

l.z�~r� is head unfrozen momentum resistance

�.z�~r� is head unfrozen momentum coefficient

#�� is density of ice � is breadth of ship ℎ is thickness of ice L is velocity of the ship

g, o are component angle fore or aft parts

Figure 4.2 showed ice resistance working on Double Acting Tanker in graph and related with velocity of the ship, while sailing ahead in unfrozen condition at 1 m ice thickness.

It can be seen from that graph, through full scale experiment testing at 1m ice thickness, Double Acting Tankers had been sailing using two variations value velocity of the ship 1.2m/s and 2.5m/s while ice resistance working measured 790 kN and 980kN respectively. These results showed an increasing value of ice resistance through the length of increasing velocity of ship.

Ice resistance of simulation results generated adjacent with experimental full scale data and approximate not slightly different. Velocity of ship 1.2m/s and 2.5 m/s related at ice resistance 810 kN and 977 kN. On the other side, total ice resistance from simulation indicated not much different. Both of lines curve simulation result (ice resistance and total resistance) almost coincided. It is indicated there are no resistance wave arising due to motion influence of water. It might be said that resistance total at unfrozen condition almost entirely is ice resistance.

Figure 4.2: Ice resistance of DAT at unfrozen channel Ice [1 m Ice Thickness], Ahead condition


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Figure 4.3: Breakdown ice resistance of DAT at unfrozen channel Ice [1 m Ice Thickness], Ahead condition

4.1.2. Ice-Ship Sailing Ahead in Ice Condition – Frozen Channel In the frozen channel of ice condition, the ice resistance is important parameter as the additional component to determine the total ship resistance that operated in this region. The breaking force is related to the breaking of ice. The total ship resistance can be stated below as the Equation (4.12) (Jaswar 2005):

lwnyCzr| = lFx;�}CT}�; + lw~v + lT}�;+ l.!.��; + l�� (4.12)

Where :

lwnyCzr| is the head frozen channel resistance

lFx;�}CT}�; is water friction resistance lw~v is submersion resistance lT}�; is friction resistance

l.!.��; is the momentum resistance

l�� is the ice breaking resistance

a. Ice Breaking Resistance (��)

The ice resistance is acting on the ship which can be defined below as the Equation (4.13): l�� = lJz�r� = �Jz�r�. ". ℎ�. . D�g, o� (4.13) Where : lJz�r� is head frozen icebreaking resistance

�Jz�r� is head frozen icebreaking coefficient " is ice flexural strength ℎ is ice thickness

is coefficient of kinetic friction of ice and hull

g, o are component angle fore or aft parts b. Water Friction Resistance (��C��) Water friction resistance which is working in the head frozen condition could be calculated as the same way at the head unfrozen channel as had been shown in the Equation (4.20) channel but for the head frozen condition there are being changed in the coefficient of water friction resistance. It was explicitly observed in the Equation (4.14).

lFx;�}CT}�; = lFTz�r� = 12 �FTz�r� . ρ . L�. � (4.14)

Where :

lFTz�r� is head frozen water friction resistance

�FTz�r� is head frozen water friction resistance coefficient

c. Submersion Resistance (��)

To determine submersion resistance in the frozen channel, it would be calculated using the same equation where previously used for submersion resistance in the unfrozen channel but foremost, it must be changed into the coefficient submersion resistance. It could be referred to Equation (4.15).

lw~v = lwz�r� = �wz�r�. �#Fx;�}G #��. /. �. �. ℎ (4.15)

Where:

lwz�r� is submersion head frozen resistance

�wz�r� is head frozen submersion coefficient

d. Friction Resistance (��) When the ship running ahead at the frozen condition, that would be making friction resistance by structure interaction of ship to ice. That can be calculated like friction resistance at the unfrozen situation but in the different of friction resistance coefficient, as to be writing in the Equation (4.16).

lT}�; = lTz�r�= �Tz�r�. #�� . /. h. ℎ. �. L 9M. /� . D�g, o, �F� (4.16)

Where:

lTz�r� is head frozen friction resistance

�Tz�r� is head frozen friction coefficient

e. Momentum Resistance (��) Ship while sailing on ice could become loss of some momentum when collision with ice. At the frozen condition, momentum loss


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can be determined using Equation (4.17). It was difference of coefficient momentum resistance with the Equation (4.11). l.!.��; = l.z�r� = �.z�r�. #. �. ℎ. L� . D�g, o� (4.17) Where:

l.z�r� is head frozen momentum resistance

�.z�r� is head frozen momentum coefficient

Figure 4.4 showed ice resistance working on double acting

tanker in graph and related with velocity of the ship, while running ahead in frozen condition at 1 m ice thickness.

It can be read through Figure 6.3 that ice resistance was occurred 920 kN when ship running in 0.8 m/s but if velocity increased to 1.67 m/s, it can produce 1050 kN working ice resistance. If it is looked by way of simulation program approach, both of lines are ice resistance and total resistance still tends to coincide or not seen existence of wave resistance due to influence motion of water. The differences are quite acquired prominent when ship speed is 0.8 m/s its getting 990 kN ice resistance using simulation whereas experimental results obtain 920 kN. But it is not happen at the second point when ship's speed is 1.67 m/s simulation results having ice resistance 1040 kN and that is close to experimental value.

If observed on Figure 4.2 concerning an unfrozen condition ice resistance was quite increasing into frozen as shown on Figure 4.4. It can be evidenced while full-scale experimental result having the same value. It confirmed in the frozen condition had occurred consolidated piece of broken ice so that was needed more thrust power of ship to break ice then had proven with increasing value of ice resistance

Figure 4.4: Ice resistance of DAT at frozen channel Ice [1 m Ice Thickness], Ahead condition.

Figure 4.5: Breakdown ice resistance of DAT at frozen channel Ice [1 m Ice Thickness], Ahead condition. 4.1.3. Ice-Ship Sailing Ahead in Ice Condition – Level Ice Ice resistance which was occurring when ship was sailing on the level ice is the whole sum of some resistance including water friction resistance, submersion resistance, friction resistance, momentum resistance and force to breaking ice resistance, as can be seen in Equation (4.18) (Jaswar 2005):

lwnyCz�5 = lFx;�}CT}�; + lw~v + lT}�;+ l.!.��; + l�� (4.18)

It could be confirmed, that is a similar method, was using to

determine resistance at the frozen situation as referred to Equation (4.27) previously. Anyway in the next subchapter below of this will be explained as clearly every piece of its. a. Ice Breaking Resistance (��)

It can be seen from Equation (4.19) ice resistance was re-emerged like Equation (4.13) where using to calculate at the frozen

condition but it was difference in a, �Jz��, that is coefficient only for breaking ice in the level ice, when ship was sailing ahead. l�� = lJz�� = �Jz��. ". ℎ�. . D�g, o� (4.19) Where : lJz�� is an icebreaking coefficient

�Jz�� is head level ice breaking coefficient

b. Water Friction Resistance (��C��) Water friction resistance in the level ice can be determined using

Equation (4.20). There is applicable �FTz�� as a water friction coefficient when ship was running in the head level ice condition.

lFx;�}CT}�; = lFTz�� = 12 . �FTz�� .ρ . L� . � (4.20)

Where :


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lFTz�� is head level ice water friction resistance

�FTz�� is head level ice water friction coefficient

c. Submersion Resistance (��)

Normally in the level ice condition, there are increasing in the submersion resistance which would be coming from fragments of ice. After ship structure interacted with ice, some fragments could be still floating and shear a shape hull of the ship and other fragments were rubbing the bottom of the hull make. Equation (4.21) could be used to determine the submersion resistance with �wz�� to be useable as submersion coefficient in the head level ice condition.

lw~v = lwz�� = �wz��. �#Fx;�}G #��. /. �. �. ℎ (4.21)

Where:

lwz�� is head level ice submersion resistance

�wz�� is head level ice submersion coefficient

d. Friction Resistance (��) Friction resistance which to be calculated in head level ice condition considers some of parameters consist of waterline angle at the fore, stem angle at the bow, dimension of the ship and density of ice, as can be found in Equation (4.22).

lT}�; = lTz��= �Tz��. #�� . /. h. ℎ. �. L 9M. /� . D�g, o, �F� (4.22)

Where:

lTz�� is head level ice friction resistance

�Tz�� is head level ice friction coefficient

e. Momentum Resistance (��) The last contribution in total resistance which occurs when the

ship was sailing in ahead mode is momentum resistance, l.z��. Equation (4.23) can be used to define its. It was similar with the momentum resistance appears in the frozen channel but because

of situation in the level ice now, so �.z�� is the head level ice momentum coefficient used to covering it. l.!.��; = l.z�� = �.z��. #. �. ℎ. L�. D�g, o� (4.23) Where:

l.z�� is head level ice momentum resistance

�.z�� is head level ice momentum coefficient

Figure 4.6 showed ice resistance acting on Double Acting

Tanker in the function of velocity of the ship. The ice resistance predicted is for the condition where the ship running ahead in level ice condition at 0.5 m ice thickness.

The Figure 4.6 shows that Range value of ice resistance between each speed of DAT results on the experimental full-scale did not show a significant difference, 2980 kN and 3100 kN, similar with range speed which can reach by ship, at the two state are almost same 0.5 m/s. If looked at the value of ice resistance and a total resistance of simulation results still coincide and that value almost the same as to experimental full-scale 3125 kN. By compare the Figure 4.6 to Figure 4.2 and Figure 4.4, this can clearly found that the increasing in value of ice resistance in level ice condition (Figure 4.6) is almost three times the value of ice resistance in the unfrozen condition (Figure 4.2) and frozen condition (Figure 6.3). Any subject difference is on value speed of vessel which can be reached. Due to higher ice load at the level ice condition, the maximum speed of DAT Tempera achieved with a power of 16 MW is 0.5 m/s, while at the unfrozen condition and frozen condition are 2.5 m/s and 1.7 m/s respectively.

Another thing needs to be examined, if simulation proceeds continuous in other high variable of ship speed, it did not turn out increasing value of ice resistance, and in the predicted results of this simulation was nearly appropriated to average value of ice resistance of experiment full-scale. So in this level ice conditions, distribution strength of ice has been uniform that means there had never been ship running on the route at least since two years ago.

Figure 11: Ice resistance of DAT at level ice [0.5 m Ice Thickness], Ahead condition


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Figure 11: Ice resistance of DAT at level ice [0.5 m Ice Thickness], Ahead condition 5.0 CONCLUSION In conclusion, this paper has discussed new method to determine performance of Double Acting Tanker operated in ice conditions such as: unfrozen and frozen channels and level ice conditions. The obtained results using the proposed method were compared with the experiment data. ACKNOWLEDGEMENTS The authors would like to convey a great appreciation to Ocean and Aerospace Engineering Research Institute, Indonesia and Universiti Taknologi Malaysia for supporting this research. REFERENCES 1. Aker Artic Technology Inc. (2010) Artic Shuttle Tanker

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