High Court · Decided 30 November 2015

Airtrust (Hong Kong) Ltd v PH Hydraulics & Engineering Pte Ltd

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[2015] SGHC 307Case no. Suit No 219 of 2013

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Grounds of decision

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Chan Seng Onn J:
Introduction
[1] The Plaintiff is a company incorporated in the Special Administrative Region of the People’s Republic of China, Hong Kong. The Defendant is a company incorporated in Singapore and is an established supplier, designer and manufacturer of heavy machinery (including tensioners and winches) for offshore use in the marine and oil and gas industry.
[2] In 2007, the Plaintiff purchased a 300 tons reel drive unit (“RDU”) from the Defendant. Attached below is a photograph of the RDU that was fabricated by the Defendant. The photograph shows a 300 tons reel (brown in colour) sandwiched between the two towers of the RDU (blue in colour) mounted on skids (blue in colour) on the deck at the aft of the ship, the “Maersk Responder”.
[3] For a better understanding of the overall structure of the RDU, I have also attached below a colour schematic diagram of the RDU.
[4] The RDU was delivered to the Plaintiff on 10 April 2008 and thereafter mounted on board the “Maersk Responder” for the laying of undersea umbilical for Nexus Energy at the Longtom field in the Bass Straits of Australia. After having laid one complete reel of umbilical and in the course of laying the second reel, a major failure of one of the gearbox assemblies occurred. The hydraulic drive motor and gear assembly on one of the towers of the RDU (ie, Tower A) came off its mounting and fell down. This incident caused the Plaintiff to investigate into whether there were any inherent manufacturing or design defects in the RDU.
[5] In this action, the Plaintiff claims that the RDU already suffered from manufacturing and design defects at the time of delivery. Broadly, these defects pertain to the following:
(a) bolts;
(b) gears;
(c) bearings;
(d) structural strength of the two RDU towers and other structural components;
(e) excessive deflection under load of the sub-frame carrying the drivetrain;
(f) brakes; and
(g) the failure to manufacture the RDU according to the specifications and drawings.
[6] The parties have agreed that the court is to determine this claim under Singapore law notwithstanding the choice of Western Australian law in cl 21 of the Sale and Purchase Agreement (“Sale and Purchase Agreement”) for the RDU.
[8] After a technically challenging trial over several days where 17 factual witnesses testified and seven expert witnesses from different fields of engineering gave their expert opinion as a group in a “hot-tub” instead of sequentially, and after having carefully considered all the relevant evidence and the detailed and comprehensive submissions from both parties, I find on a balance of probabilities, for the reasons stated in this judgment based on the pleadings as amended, the facts that I have found and the totality of the factors that I have evaluated and taken into consideration, that the RDU:
(a) is not of merchantable quality;
(b) is not fit for the purpose for which it is intended to be used;
(c) is not free from defects in design, manufacture or workmanship; and
(d) does not meet the relevant industry standards and certifications, or the specifications and certifications stipulated in the contract between the Plaintiff and the Defendant.
[9] For ease of reference, I have attached at Appendix 1 a “Glossary of Witnesses” with a brief summary of the focus of their evidence at the trial. In the course of this judgment, I shall refer to the evidence of some of these witnesses.
Sale and Purchase Agreement
[10] It is not disputed that the Defendant was aware that the RDU was purchased by the Plaintiff specifically for lease to Trident Offshore Services (“Trident”) for the laying of undersea umbilical in the Bass Straits of Australia for Nexus Energy in the Longtom Project.
[11] The Plaintiff and Defendant entered into the Sale and Purchase Agreement dated 7 September 2007 which provides, inter alia, that:
(a) the Defendant was to sell a 300 tons RDU and deliver it on 14 January 2008 to the Plaintiff;
(b) the total purchase price of the RDU was $895,000 inclusive of “ABS [American Bureau of Shipping] Full Certification” which was itemised at a cost of $20,000 (See Schedule 1 of the Sale and Purchase Agreement at paragraph 2 - Scope of Supply);
(c) the RDU was to be fully certified by ABS;
(d) the RDU would be of merchantable quality, fit for the purpose for which it was intended and free from any latent or apparent defect in material or workmanship (See cl 15 of the Sale and Purchase Agreement); and
(e) the Defendant would perform all work diligently, carefully, in a good and workmanlike manner and in accordance with accepted industry standards (See cl 24 of the Sale and Purchase Agreement).
[12] When it transpired that ABS does not provide certification for machinery such as the RDU, the Defendant suggested in an email dated 12 October 2007 that ABSG Consulting Inc (“ABSG”), an entity related to ABS, carry out the “design review, site survey and final testing, same as what [the Defendant] did for the Acergy 100 Ton reel drive system” instead. The certificate from ABSG for the Acergy 100 Tons spooler tower was attached for the Plaintiff’s consideration. Mr Terry Griffiths, the Principal Pipeline Engineer of Trident (“Mr Terry Griffiths”), replied on behalf of the Plaintiff vide an email dated 17 October 2007 that “the use of ABS Consulting for design review of the reel towers is appropriate given there is no Class applicable, however can you please ensure the design/structural codes selected are appropriate” [emphasis added in bold]. The Defendant’s project manager, Dr Yang Ting, clarified during cross-examination that the certification for the RDU, now agreed to be performed by ABSG instead of ABS, remained a “full certification”.
[13] It is not disputed that the Defendant was solely responsible for designing the entire RDU although this obligation was not expressly stated in the Sale and Purchase Agreement. In fact, the Defendant’s Confidential Bundle of Documents (“Confidential Bundle”), containing inter alia the design drawings and calculations submitted to ABSG for the purpose of obtaining certification from ABSG for the RDU design, was never provided or shown to the Plaintiff at that time. Basically, the Defendant kept the design of the RDU confidential. The Plaintiff was unaware of the engineering details and the design calculations in relation to the design of the RDU.
[14] After the Defendant had designed and fabricated the RDU, a factory acceptance test (“FAT”) was carried out and signed off by Mr Terry Griffiths on behalf of the Plaintiff on 26 February 2008 and counter-signed by Mr Chong, the then ABSG surveyor with ABSG, and Mr Steven Gan, the Assistant Manager-After Sales of the Defendant. ABSG issued certificates for the design reviews it carried out, as well as the manufacturing processes and the FAT.
Inadequate design of the RDU
[15] I will deal with the design inadequacies first and there are several alleged by the Plaintiff. I will address them seriatim.
Failure to take the vessel roll into account
[16] The Plaintiff submits that the Defendant failed to take vessel roll into account in the RDU design. It is not disputed that the inertial forces arising out of the vessel’s accelerations must form part of the design considerations for the RDU. The input design parameters for any computer program (eg, the STAAD.Pro analysis) using a finite element method (ie, Finite Element Analysis or “FEA”) to perform the stress calculations for the various components of the RDU structure mounted on board a vessel must therefore take into account the expected motions of the vessel due to current, wind and wave action.
[17] I accept the explanation of Dr Yang Ting that there are basically two acceptable methods of specifying such input design parameters for the inertial forces due to the vessel’s motions for the purpose of stress calculations:
(a) Specifying the equivalent accelerations that the equipment mounted on a vessel would be subjected to in the X, Y and Z directions corresponding to the ship’s transverse, vertical and longitudinal directions, which are to account for pitch, sway, yaw, heave, roll and surge motions of the ship out at sea; or
(b) Specifying the maximum pitch, yaw and roll angles of the vessel and their respective periods of oscillation, the maximum sway, heave and surge and also the specific location of the equipment installed on the vessel, so that calculations can be performed to translate and resolve them into their corresponding accelerations in the respective X, Y and Z directions.
[18] In either case, those equivalent accelerations arising from the vessel’s motions can be used for computing the inertial forces in the X, Y and Z directions for the structural and stress analysis.
[19] But if in either case the maximum roll angle of the vessel for example is expected to be large, then the weight of the equipment itself (which is a gravitational force and not an inertial force) must be resolved into the X and Y components because of the large roll angle and then added to the inertial forces in the respective X and Y directions (which arise separately from the equivalent accelerations due to the ship’s roll) in order to perform the structural and stress analysis. If that is not done, there can be a serious underestimate of the expected total force (due both to inertia and gravity) in the X direction when the vessel is at an inclined position relative to the horizontal at its maximum roll angle.
[20] It is also significant to note that the design parameters in the form of the accelerations in the X, Y and Z directions for use in the design calculations of the RDU were in fact proposed, not by the Plaintiff, but by the Defendant two months after the contract was made and at a point of time when the Defendant was already aware that the RDU was to be utilised in the Bass Straits of Australia. Mr Terry Griffiths requested Dr Yang Ting on 15 November 2007 to advise the Plaintiff of the design accelerations to be used for the RDU reel tower. Dr Yang Ting responded on behalf of the Defendant on the same day that:
[21] I find that these are acceleration values for computing purely inertial forces due to the vessel’s motions. The vertical gravitational force or weight is not included in these acceleration values. If it had been, then the acceleration value for heave for example cannot be as low as 3.7 m/s (or 0.377 g) when the equivalent acceleration value due to gravity is already 9.81 m/s (or 1.0 g). Obviously, the weight of the equipment must be added to the inertial vertical force due to the heave acceleration of 3.7 m/s to derive the total maximum vertical force acting on the structure for the purpose of stress analysis, thereby giving 3.7 m/s + 9.81 m/s = 13.51 m/s (or 0.377 g + 1.0 g = 1.377 g). The same logic should apply in the case of roll. Any X component of gravitational force or weight acting in the X direction (due to the inclined position of the equipment on the deck of the vessel at the maximum roll angle), if significant, must be added to the inertial force in the X direction due to the maximum roll acceleration at that maximum roll angle to give the total maximum force in the X direction acting on the structure for the purpose of stress analysis.
[22] The weight of the equipment must therefore be separately accounted for over and above the inertial forces arising from the stated acceleration values in the structural and stress analysis. In other words, in order to assess the maximum value of the total force in the X, Y and Z directions, the components of the weight (due to the vertical gravitational force) resolved in the X, Y and Z directions (relative to the vessel), if significant, when the vessel is positioned at its expected maximum roll and pitch angles, must be added to the inertial forces due to the stated acceleration values in those corresponding directions as provided in [20] above for the design.
[23] After a careful examination of the calculations, I find that the Defendant had additionally provided for the vertical force due to the RDU’s weight in the vertical Y direction over and above the inertial force due to the heave design acceleration value of 3.7 m/s (or 0.377 g) (as it should), but had simply ignored the possible transverse force due to weight in the X direction when the vessel is at an inclined position due to roll (which it should not do). Effectively, the Defendant simply assumed that the maximum possible roll angle of the vessel out at sea is either zero or always negligible in value, which in my view is an incorrect assumption considering the kind of sea conditions that may be encountered in the expected areas of operation of the RDU. I thus agree with the Plaintiff’s submission that at least in this respect, the Defendant failed to take the vessel’s roll into account in designing the RDU. Bearing in mind that the full reel alone is a massive 300 tons, a small roll angle of say 10° from the vertical is already enough to generate a component of force due to the reel’s weight on the towers in the X direction of some 52.1MT, which is not at all negligible. A roll angle larger than 10 degrees will obviously generate an even higher component force due to the reel’s weight in the X direction on the towers, which must be added to the maximum inertial force in the X direction specified at 300 tons × 2.2 m/s = 66 MT for the structural and stress analysis.
[24] To overcome this criticism of its design, the Defendant argues rather unconvincingly that the 2.2 m/s acceleration in the transverse direction would have included both effects of roll (ie, the roll acceleration to account for the inertial force and the equivalent acceleration due to the resolved X component of the combined weight of the Tower and the full reel of 300 tons at the extreme end of the roll motion of the vessel). If that is the argument, then the Defendant as the designer who had set these design parameters, has not shown to me that the quantitative figure of 2.2 m/s is sufficient in itself to account for both effects during the typical roll motion of the vessel during the RDU umbilical laying operation out at sea and in particular in the Bass Straits, with the typical sea conditions expected in that area. The Defendant simply adopted the accelerations used in its design of a 300 tons RDU that it had constructed for another company, Acergy, without establishing the adequacy of the transverse acceleration design value of 2.2 m/s (or 0.224 g) in the first place for this RDU (to be deployed for use in the Bass Straits) based on its purported premise that both effects (ie, gravitational and inertial effects) had been subsumed under one equivalent acceleration value.
[25] Without the Defendant producing a set of calculations setting out its roll assumptions and the necessary vector calculations to demonstrate the equivalence with the acceleration figure adopted in the RDU design, it is not possible to establish to what extent and with what sufficiency each of the two effects of the roll has been incorporated within the purported composite acceleration figure of 2.2 m/s Without knowing the roll assumptions used by the Defendant, it is also not possible to establish if the acceleration values used for the RDU’s design are acceptable considering the kind of sea and weather conditions that the vessel with the RDU mounted on its deck is expected to be operating in.
[26] In the absence of information from the Defendant of the design assumptions it had adopted in setting the design criteria for the X, Y and Z accelerations, the Plaintiff’s expert, Mr Moore, had no alternative but to compare the acceleration values adopted by the Defendant for the RDU design with the typical acceleration values that classification societies have stipulated for use in the design and construction of ships and offshore structures to ensure compliance with their respective classification standards. At least, this would help to indicate whether there is any serious deficiency in the acceleration values adopted by the Defendant upon which the design of the RDU structure was fundamentally premised.
[27] In this regard, I have no reason to disagree with the following expert opinion and comments of Mr Moore in his affidavit of evidence-in- chief (“AEIC”):
[28] Where there is a significant difference between the acceleration values stipulated in the design codes of the well-known classification society cited by Mr Moore and those accelerations adopted by the Defendant for its RDU design, it behoves the Defendant to justify in more detail how and why those accelerations values can be so much lower. When the Defendant does not explain how its low acceleration values were derived (which would affect the safety and structural integrity of the RDU as a piece of offshore equipment mounted on board a ship for operations out at sea and in particular in the Bass Straits), I will draw an adverse inference that the RDU, designed using the Defendant’s acceleration values that were significantly lower than those recommended by DNV, is under-designed in terms of its structural strength. This is one of the many factors in my overall consideration as to whether the RDU is of merchantable quality and fit for its purpose.
[29] Of course, if the Plaintiff can further demonstrate, as will be seen later, that the RDU is already under-designed when it is based on the Defendant’s transverse acceleration design value of 2.2 m/s (or 0.224 g), then the degree of under-design will necessarily be even greater when the acceleration design value adopted by the Defendant does not in fact include the effect of weight in the X direction due to the vessel’s roll.
[30] The Defendant next argues that it never agreed to design an RDU that could withstand the worst conditions in the Bass Straits, thereby suggesting that they had not in fact done any analysis to determine whether those accelerations adopted for its design were in fact appropriate for vessels operating in the Bass Straits. The Defendant relies on cl 9 of the Sale and Purchase Agreement which states that the purchase order including any attachments constitutes the entire agreement between the supplier and contractor. The Sale and Purchase Agreement makes no reference to the fact that the Defendant was to design the RDU for operations in the Bass Straits, although the Defendant does not dispute that it had been told that the RDU would be used for umbilical laying for the Longtom Project in the Bass Straits.
[31] The Defendant wholly misses the point. When the Plaintiff as the client of the Defendant informs the Defendant of the specific area of operation of the RDU, and the Defendant, as the designer of the RDU, proposes values of accelerations as parameters for the design, which are subsequently stipulated in the Sale and Purchase Agreement, the responsibility for ensuring the appropriateness of those values of accelerations to be adopted as design parameters does not shift to the Plaintiff merely because the Plaintiff agrees or does not object to those design parameters. The Plaintiff is not the designer and does not know whether the design parameters are the correct and appropriate parameters to be used. Neither has the Plaintiff contractually accepted any responsibility to check on the correctness and appropriateness of such parameters. The Plaintiff may not even have the design expertise and domain knowledge. It is the Defendant as the designer under the contract who remains solely liable for the design parameters that it wishes to adopt even though the Plaintiff, as its client, may have accepted the Purchase Order which spells out the design parameters chosen by the Defendant. It must be borne in mind that the Plaintiff is not contractually a co-designer of the RDU. The Plaintiff engaged the Defendant to be the sole designer of the RDU for which the Defendant had been made fully aware that the RDU “[p]rojects can be worldwide but first project is in Australia (Bass Strait).”
[32] I give an example. If after knowing that the RDU is to be used in the Bass Straits, the Defendant proposes extremely low acceleration values as design parameters, which are only appropriate for machinery installed on a vessel operating in relatively calm inland waters and which are therefore totally inappropriate or inadequate for machinery mounted on an ocean going vessel, can the Defendant, who has in the first place contractually undertaken the sole responsibility for the RDU design, turn around later, rely on its entire agreement clause and then raise a defence that the Plaintiff has already accepted those very low acceleration values proposed by the Defendant, and therefore the RDU, as designed, is deemed fit for its purpose simply because those design parameters have been contractually accepted by the Plaintiff? I do not think that the Defendant can do so. When the Defendant has been informed of an important purpose and the area of operation of the RDU (ie, that it is to be used out at open sea, and its first area of operation is in the Bass Straits), it is incumbent on the Defendant, as the designer, to select and use appropriate acceleration values for the structural strength design that ensures that the RDU is at least operable out at open sea, if not in the Bass Straits, and which at the minimum is consistent with the typical acceleration values stipulated by classification societies (eg, DNV). I cannot accept that the values of accelerations selected by the Defendant for the RDU design would make the RDU fit for a purpose already made known to the Defendant, when those values are way below those recommended by classification societies for use as design parameters for machinery mounted on ocean going vessels. The comparison table of acceleration values set out at [27] above shows that inappropriate acceleration values were selected, proposed and used by the Defendant as the designer. I therefore reiterate that I have to take into consideration the likely under-design in the structural strength of the RDU, based on the relatively low acceleration values selected and used by the Defendant (ie, in comparison with those set out by the classification society DNV) having regard to the fact that the Defendant had been made fully aware of the known purpose and area of operation of the RDU, when I consider whether the RDU is of a merchantable quality and fit for its purpose.
[33] I also accept the Plaintiff’s submissions that the Defendant is under a statutory obligation to ensure that the RDU is fit for the particular purposes made known to it: see National Foods Ltd v Pars Ram Brothers (Pte) Ltd [2007] 2 SLR(R) 1048 at [54]. It has never been the law that in order for this obligation to bite, the particular purpose must itself be a term of the agreement: see Benjamin's Sale of Goods (M G Bridge, gen ed) (Sweet & Maxwell, 9th Ed, 2014) at para 11-055. In short, it must be part of the design obligations of the Defendant to stipulate appropriate design parameters in the contract specifications in the light of the known purpose and areas of operation of the RDU when the Defendant has undertaken full responsibility for its design.
[34] Another troubling aspect is that the Defendant wrongly treated the reel to be fixed with a pin joint at both ends of the tower in its STAAD.Pro analysis. Clearly, it is erroneous to assume that the reel is not free to slide along its shaft in the transverse direction because the tower reel inserts are simply inserted into the reel hub and the reel is therefore free to slide. Accordingly, the transverse force calculations for the FEA analysis are affected.
[35] After reviewing the Defendant’s own design calculations, Dr Yang Ting eventually admitted that the Confidential Bundle shows that the Defendant had wrongly calculated the transverse force acting on the RDU tower due to the transverse acceleration by attributing only half of the transverse force to each of the RDU towers. Based on the Defendant’s design parameters, the mass of the full reel of 300 tons subjected to a transverse acceleration of 2.2 m/s in fact produces a total transverse force on the RDU of 66 MT (ie, 300 tons x 2.2 m/s). Since the reel can slide on the shaft (because it is not secured on both sides), it is wrong to assume that the 66 MT transverse force is shared equally by both RDU towers. If the frictional force is disregarded, the full 66 MT transverse force (and not 33 MT as was adopted in the calculations in the Confidential Bundle) must be taken to bear down entirely on the single tower that is restraining the full effect of the roll motion of the 300 tons reel.
[36] I therefore find that the Defendant had wrongly computed the transverse force acting on the tower based on the transverse acceleration of 2.2 m/s that it had itself proposed and adopted as one of the input criteria for the design of the 300 tons RDU. Accordingly, the Defendant significantly underestimated the transverse inertial force acting on each RDU tower during its structural design calculations premised on its own input criteria.
[37] The ingenious response from the Defendant’s expert, Mr Huang, is that the frictional force restraining the slide in the transverse direction, which was previously omitted and not considered in the calculations, must now be taken into account for the structural analysis. The Defendant had to find a way to remedy the error in the structural design calculations.
[38] Ms Renuka Devi, the Lead Structural Engineer from ABSG, very helpfully re-computed her structural analysis pursuant to directions from the court, using the same transverse acceleration of 2.2 m/s, but with the addition of the friction force. She summarised her results at Exhibit D3 (reproduced in the table below (“the Table”)) showing the “Beam Unity Checks” under various scenarios and with different assumed friction coefficients.
[39] After considering Ms Renuka Devi’s detailed explanation in court, I accept that the figures in scenario 4 ie, columns 4a, 4b and 4c in the Table above are inaccurate because they do not correctly simulate the actual forces on the structure in the as-built condition. Furthermore, all loading conditions (referred to as “LC” in the Table above) both relevant and irrelevant have been included in scenario 4. The right approach is to take into account only the relevant LC and model the actual forces acting on the actual as-built structure as follows under scenario 3:
(a) with “Struts bolted” (ie, correctly modelling fixed joints and not ball joints at the strut connections);
(b) with “reel sliding” (ie, correctly modelling the as-built condition where the 300 tons reel is free to slide along its shaft in the transverse direction),
where the coefficients of friction are respectively assumed to be “0” under column 3b (for a conservative design where friction forces are ignored), “0.16” under column 3c (for wet or lubricated steel to steel contact) and “0.3” under column 3d (for dry steel to steel contact).
[40] In the opinion of Ms Renuka Devi, the coefficient of friction of “0.3” is inappropriate because of weather and wet conditions out at sea. Basically, if friction forces are to be taken into account for the structural analysis, then in Ms Renuka Devi’s view, the appropriate value for the coefficient of friction is “0.16” and accordingly, column 3c of the Table is the correct column to look at and not the other columns.
[41] In the Table produced by Ms Renuka Devi, a Beam Unity Check value that exceeds “1” indicates that the beam fails the unity check in terms of its structural strength. The beam designer therefore has to redesign that beam to strengthen it. The higher the value is above “1”, the greater is the inadequacy of that particular beam as a structural strength member. As can be seen, that critically depends on the correct choice of the appropriate coefficient of friction for the sliding surfaces.
[42] Based on a friction coefficient value of “0.16”, it appears that only one structural beam member marginally fails the Beam Unity Checks (ie, Beam member number “107” which has a unity check value of 1.022) as shown in column 3c of the Table. If one adopts a more conservative design where the effects of friction are ignored entirely as was done in the initial calculations, then many more structural beam members (ie, 10 beams) fail the Beam Unity Checks as can be seen under column 3b in the Table above.
[43] I note that Ms Renuka Devi had opined that she would not normally take friction into account in her structural analysis. She said “I mean, if the thing has been left to me, then I will not consider any friction, as a conservative design”. I agree with Ms Renuka Devi in this regard. It is a correct conservative engineering approach to take for the RDU design. I am of the view that having to rely on frictional forces as a restraining force to help to reduce the stresses on a ship deck mounted structure exposed not only to all the weather and sea elements, but also to the roll motion of a vessel out at open sea where rough sea state conditions can be expected is really scrapping the bottom of the barrel in terms of good engineering design for deck mounted equipment on board a ship such as the RDU. I would have expected instead to find adequate safety factors built into the RDU design (ie, with unity check values well below “1” for the structural beams) in order to provide for some over-design to cater for many potentially unknown and unexpected factors in the course of the actual RDU operations out at open sea. A practical engineering design for ship board equipment in my view should not normally be pushed to its engineering limits since weight is not as significant a consideration for design purposes on board a large ship unlike the design of a fighter jet where light weight is such a critical design consideration due to the need to achieve high speed, have a large fuel capacity for long range and yet be able carry a heavy weapons payload. Designing for stronger and heavier structural elements with sufficient safety factors should be the correct approach rather than one that results in an RDU design that teeters at its engineering limits, even when based on a 2.2 m/s transverse acceleration, which by DNV standards is already far too low and unacceptable as a design criterion for portable offshore equipment.
[44] For the reasons stated above, I find that the structure of the RDU is inadequately designed as the assistance of friction should be ignored and far too low acceleration values were used as design criteria by the Defendant.
Failure to design for fatigue
[45] The Defendant admits that fatigue was not considered in the design of the RDU structure. The Defendant takes the view that it is not necessary to design for fatigue. As the RDU is subject to cyclical loading, the Plaintiff submits that fatigue ought to have been considered.
[46] I agree with Mr Drago that it “simply doesn’t make engineering sense to design rotating machinery without considering fatigue.” The bearings and shafts in the RDU must be designed for fatigue as there will be several hundred thousands of hours of operation expected within the lifetime of the RDU. However, is the same applicable to the non-rotating structural parts of the RDU?
[47] The Defendant says that the towers of the RDU were designed and manufactured in accordance with the AISC 360-05. As highlighted by Mr Moore, AISC 360-05 is a specification for structural steel buildings and Section 3.1 of Appendix 3 stipulates that an evaluation of fatigue must be made if the application of live loads is more than 20,000 cycles. I agree with Mr Moore that the RDU will see more than 20,000 cycles of loading in its lifetime due to wave action on the vessel. The Defendant cannot pretend for the purpose of engineering analysis that the RDU is installed on land and not on board a ship. Accordingly, the RDU should have been assessed in accordance with the fatigue section of that code. Ms Renuka Devi also accepted that a check for fatigue should be carried out on the welding and connections of the RDU but the Defendant had not asked ABSG to do it. Dr Eccles and Mr Natarajan, testifying as experts on behalf of the Plaintiff, asserted that the connections are in many cases the most relevant areas to check for fatigue failure because they often fail first. I agree because welding connections typically have much lower fatigue strength than the underlying metal substrate of the structural members due to stresses from welding and heat distortion at the heat affected zone around the weld.
[48] It appears that the Defendant took no account of the cyclical loading that the RDU structure would be subjected to. Dr Yang Ting had to concede that there was some “oversight” not to have considered fatigue in the structural design calculations due to cyclical loading. As such, structural failure due to fatigue may well occur within the operating lifespan of the RDU.
[49] The fact that both the rotating and non-rotating parts of the RDU were not designed to withstand fatigue is yet another important factor that I have to take into consideration in my evaluation of the merchantable quality and fitness for purpose of the RDU.
[50] Dr Yang Ting’s explanation for the omission of fatigue analysis was that site-specific information on the load spectrum and the areas of operation of the RDU were needed. He said “[b]ecause the customer cannot give you an input [on] how the load spectrum looks like. We cannot do the calculation.” That to me is a lame excuse for a designer. The Plaintiff had informed the Defendant of one specific area of deployment of the RDU ie, Bass Straits. It is for the designer to research and find out for himself the load spectrum for the specified area of operation (or the equivalent load spectrum for similar areas of operation). It is not for the designer to complain that the customer did not provide the load spectrum information (which the designer in any event had not even asked his customer), or to simply ignore the engineering design requirement to perform the fatigue analysis merely because the customer had not volunteered any load spectrum information for that area of operation to enable the designer to perform fatigue calculations. If any design data is needed, I am of the view that the designer himself is responsible for finding out the required design data to enable him to undertake the fatigue calculations required for the RDU design. That is what the Defendant had been contractually engaged to do.
[51] The Plaintiff urges me to infer that the Defendant deliberately omitted to ask ABSG to perform any fatigue calculations because the Defendant knew that such calculations if done would have shown up the structural design to be inadequate and ABSG would not have certified the RDU. At this stage, I am not prepared to make this inference although I am prepared to find a high level of technical or engineering incompetence on the part of the Defendant’s designers not to have taken fatigue into account in the RDU design.
Failure to design for the effects of wind
[52] Mr Loh, the General Manager (Engineering and Operations) of the Defendant, admitted that the Defendant should have checked for wind loading on the RDU structure but did not do so at the design stage. Ms Renuka Devi was of the same view that the effects of wind should be considered but she did not take that into account in her design review of the RDU as the Defendant told her that it was not necessary to do so.
[53] The ABS Guide for Certification of Lifting Appliances July 2007 sets out at Chapter 2 the requirements for the certification of, inter alia, heavy lift and gantry types of cranes installed on board vessels classed by ABS and requires data on the environmental loads on the crane structure including the effects of wind to be submitted together with other plans and supporting data for review and approval. Clearly, the dynamic wind loading is a design consideration.
[54] As the RDU has to hoist a reel of some 300 tons upwards from the deck to the operating height prior to laying the umbilical, I consider the RDU to be akin to a heavy lift crane and the engineering considerations stipulated in the ABS Guide would be a useful benchmark for the engineering standards to which the RDU should have been designed. As the reel diameter is some 9.2 metres, the reel has a very big surface area subject to wind load. As such, I disagree that dynamic wind loading can be ignored in the design of the RDU.
[55] I find that the failure of the Defendant to take account of wind load has contributed to the overall structural under-design of the RDU.
Bolting arrangements were under-designed
[56] I accept the opinion of the Plaintiff’s experts that the bolting arrangements for the following were inadequately designed:
(a) the bolts securing the struts to the main tower frame;
(b) the bolt connections between the pinion bearing housing and the sub-frame;
(c) the bolts securing the Sauer Danfoss gearbox to the motor mounting;
(d) the bolt connections between the pinion and gearbox flange; and
(e) the bolt connections for the end plates of the drive shaft.
[57] The Plaintiff’s Multimedia Bundle contains photographs and videos showing that loosening of the bolted joints had been observed. I also refer to an observation in a report on the tests conducted on 26 and 27 September 2012 titled “ATHK 300 T Reel Handler System – Load, Deflection, Alignment and Dimension” (“Nobletech Report”):
[58] The available video and photographic evidence and observations in the test report strongly support the theoretical analysis by the Plaintiff’s experts that the bolting arrangements were in fact under-designed. This is a serious defect because the structural integrity of the RDU depends on the bolts holding the various critical components of the RDU together.
Bolts securing the struts to the main tower frame
[59] Mr Moore explained that bolted struts are not generally used for RDU construction as was done in the case of the Defendant’s 300 tons RDU. According to Mr Moore, construction for this range of RDUs usually comprises an all welded or welded/bolted construction and uses much heavier section members. Although some manufacturers do employ the welded/bolted method of construction, however they are for lighter duty RDU applications, typically in the region of 80 tons to 150 tons. Mr Moore cites RDUs supplied by Aquatic Engineering & Construction (UK) and Sparrows Baricon (UK). Their RDUs in the 300 tons range differ insofar as they are of a more robust, all welded construction giving much greater structural rigidity than the bolted/welded construction of the Defendant’s RDU.
[60] I agree with the Plaintiff’s submissions that although the bolting arrangements for the main struts are of critical significance to the structural integrity of the Defendant’s RDU, yet the Defendant has not produced any evidence to show that they had carried out any calculations, at the design stage, to show the adequacy of these bolting arrangements. I draw the inference that no such calculations were in fact done at the design stage.
[61] Dr Eccles carried out detailed calculations to check the adequacy of the bolted joints at the struts. He found that they were inadequate for both the tensile as well as shear forces. See Dr Eccles’s AEIC from pp 109 to 125. Mr Huang on the other hand produced a set of calculations based on a pre-tension of 50% yield on four bolts of M30, each with a diameter of 30 mm, and a friction coefficient of 0.5. Mr Huang attempted to demonstrate that the maximum static friction generated was greater than the maximum slippage force of 353,009 N. However, Dr Eccles pointed out the erroneous approach in Mr Huang’s calculations where the stress area of the cross section of the cylinder of the bolt was used when it should have been the actual thread stress area which is the relevant effective stress area. On re-calculating based on the same parameters assumed by Mr Huang, Dr Eccles explained at Exhibit P15 that slippage would occur:
[62] Mr Hooper, the Defendant’s expert witness, then pointed out that the yield strength of the bolts used was actually 640 MPa instead of 550 MPa. However, even with the higher yield strength figure, Mr Hooper basically accepted that this would still be “low” and in my view, inadequate from a design perspective if the necessary safety factors are to be included:
[63] The second deficiency in Mr Huang’s calculation is his adoption of 0.5 as the value for the coefficient of friction on the basis that the bolted joint is not painted. However, the bolt joint areas were in fact painted and the appropriate coefficient of friction for a painted joint is much lower. According to Dr Eccles, it is between 0.2 to 0.3 according to a research paper by Heistermann, C., Behaviour of pre-tension bolts in friction connections, Lulea University of Technology, 2011. Mr Huang finally had to concede during the hot-tubbing that with a painted joint, the coefficient of friction of 0.5 that he had assumed could not be achieved. In short, the bolted joints would slip and come loose.
[64] Dr Eccles also pointed out a third error in Mr Huang’s calculations. Mr Huang wrongly assumed that the bolt was fully threaded into the joint, but that does not appear to be the case in the design of the bolted joints. It can be seen from the engineering drawings that there would not have been full engagement of the entirety of the thread length (see Confidential Bundle at A127).
[65] The effect of the failure of the bolts was explained by Dr Eccles and Mr Natarajan as follows:
[66] On the totality of the expert evidence, I find on a balance of probabilities that the bolting arrangements for the RDU struts were inadequately designed. I accept the explanation provided by Dr Eccles and Mr Natarajan on the severe implications arising from a potential failure of the bolting arrangements on the strength and integrity of the RDU structure as a whole, bearing in mind that the RDU has to carry a massive 300 tons reel and may be subjected to high wind load and considerable inertial forces due to wave action depending on the kind of sea condition encountered during the umbilical laying operation.
[67] The Defendant’s only response to the Plaintiff’s submission that the bolting arrangements for the RDU struts were under-designed is simply that Dr Eccles’s calculations are unreliable because Dr Eccles used data from Mr Natarajan’s FEA analysis, which the Defendant submits is unsound. Since I have found to the contrary at [145] and [151] that Mr Huang’s FEA analysis is unsound and have accepted the results of Mr Natarajan’s FEA analysis instead, this puts to rest the Defendant’s argument that Dr Eccles’s calculations are unreliable. In any event, Dr Eccles had also analysed the adequacy of the bolts for the tensile stress independently of Mr Natarajan’s report. Dr Eccles re-calculated using the tensile forces that the Defendant’s own experts had assumed to be acting on the bolts and demonstrated that the bolting arrangements and the strut bolts were nevertheless still under-designed.
[68] The Defendant in its closing submissions simply attached a further report from Mr Huang dated 1 October 2015 at Appendix E in a final desperate attempt to show that Dr Eccles’s analysis is erroneous. At the conclusion of the trial, I did not ask for any further report on bolt analysis by any expert from either side. The Plaintiff strongly objects to its introduction as further evidence. As the trial is over, the Plaintiff no longer has any opportunity to bring further expert evidence to rebut the further report from Mr Huang or to cross-examine him on it. As the Plaintiff will be prejudiced, I am not admitting the further report of Mr Huang as evidence.
Bolting arrangement on the pinion bearing joint
[69] The pinion gear is designed to drive the main gear. It is connected to the sub-frame housing by a bearing. At the other end, it is bolted to the Sauer Danfoss gearbox. The gearbox is itself mounted by a bolting arrangement. The pinion bearing is attached to the structure through an arrangement of six M-16 bolts, each 30 mm long. As the grade of the bolts was not specified in the drawings, the Defendant’s practice according to Dr Yang Ting would be to use a bolt grade of 8.8.
[70] Dr Eccles explained in detail in his AEIC why he considered the bolting arrangement on the pinion bearing joint to be inadequate. Mr Hooper confirmed that the bolts would slip on the design load:
[71] Mr Hooper further accepted that the under-design of the bolting arrangements at the pinion bearing housing would result in misalignment of the gears and an increase in the loads and stresses on the gears, which I am inclined to believe may also cause premature gear failure due to the increased gear loading:
[72] The Defendant’s answer to this in its submission is that if there was indeed any misalignment of the gears due to slip, then uneven wear patterns across the gear teeth would result. But there was uniform burring found across the teeth of the gear wheel and pinion gear of Tower A of the RDU during the investigations following the failure thereby indicating that there had been no misalignment and hence no slippage at the pinion bearing housing. The Defendant further highlights that prior to the catastrophic failure of the RDU on 20 May 2009 (see [201]), a successful full torque test during the FAT was carried out on 28 February 2008 and the RDU had also finished laying umbilical No. 1 and completed a partial laying of umbilical No. 2 during the Longtom project. The Defendant is thus suggesting in its submission that the successful torque test and the completion of laying of more than one reel of umbilical demonstrate that the RDU is of merchantable quality and fit for its purpose.
[73] The Defendant wrongly assumes that the RDU was operated to its design load during the course of the Longtom project for a sufficiently long period of time prior to the catastrophic failure, which therefore proves that at the design loads, the RDU is still of merchantable quality and fit for its purpose. However, there is no evidence that the RDU had in fact been operated to its full design load during the relatively short umbilical laying operation prior to the catastrophic failure. Dr Eccles’s theoretical analysis of the slippage of the bolts that hold the pinion bearing housing in place was based on the full design load condition. If the RDU had not been operated to its full design load condition for a sufficient period of time out at sea, I would not at all be surprised if serious slippage and misalignment (as predicted by the theoretical analysis) had not occurred yet. The fact per se that there was no uneven wear found on the gears (even if true) is not sufficient for me to reject the conclusions of both the Plaintiff’s and Defendant’s experts, ie, Dr Eccles and Mr Hooper respectively, that they were “in agreement that the joint would slip on the design load, as designed.”
[74] As I have no good reason to disagree with the conclusions of the experts from both sides that the bolts would slip on the design load, I accordingly conclude that the bolting arrangement on the pinion bearing joint was under-designed.
Bolting arrangement on the mounting of the gearbox and the pinion/gearbox flange interface
[75] The pinion is connected to the output flange of the Sauer Danfoss gearbox by a series of bolts. The gearbox is itself located on a motor mounting as follows:
[76] The Defendant submits that the Sauer Danfoss gearbox mounting was not under-designed because the Defendant had followed the manufacturer’s specification by using three M20 bolts of a higher grade 12.9 on each side of the mounting and by tightening them to the specified torque of 540 Nm. However, no records were produced to support the Defendant’s assertion that these bolts were in fact tightened to the torque of 540 Nm at the time of manufacture or testing prior to the delivery of the RDU to the Plaintiff.
[77] In any event, the main question that has not been answered by the Defendant is what are the design parameters and load assumptions used by the manufacturer for the Sauer Danfoss drive unit and gearbox for its specification of those bolt sizes, grades and tightening torques, and whether the load assumptions of the manufacturer exceed those loads computed for this RDU design where the Sauer Danfoss drive units and gearboxes are being used for a totally different purpose (ie, to drive a 300 tons reel and not a much smaller and lighter cement mixer drum). Furthermore, the manner in which the Sauer Dandfoss drive unit and gearbox is mounted on a road-going cement mixer truck to turn the cement mixer drum is very different from that for the RDU. There is a very substantial eccentricity, non-existent for the cement mixer truck, but present in the RDU, which creates very significant shear loads for the mounting bolts as can be seen in the illustrations below. I accept the Plaintiff’s submissions that the Sauer Danfoss drive units and gearboxes were never designed to take these additional large loads present in the RDU at its design loads.
[78] Without knowing the manufacturer’s design parameters and load assumptions for comparison with the expected loads arising from the RDU operation, I do not accept the Defendant’s bald assertion the Sauer Danfoss drive unit and gearbox mounting is correctly designed merely because the Defendant had adopted the manufacturer’s specifications for the mounting bolts and tightening torque requirements.
[79] Dr Eccles explained in his AEIC why the bolts at the Sauer Danfoss gearbox mountings were under-designed. Considering the maximum working design tension of 15 tonnes for the umbilical, the shear force generated is enormous and far exceeds the capacity of those bolts to restrain and the joint would slip. Slippage, even by fractions of a millimetre, would be sufficient to loosen the bolts, induce fatigue loading and reduce its operating life.
[80] Mr Hooper did not challenge Dr Eccles on this point but took the position that in the case of the design of the foundation at the base of the Sauer Danfoss gearboxes, restraining keys or stoppers were in fact put there to take the shear load that Dr Eccles was referring to but “unfortunately, there are, from what we can see, gaps between those stoppers and the gearbox”.
[81] It appears to me that even if the design intent was to have the restraining keys or stoppers (instead of the bolts) absorb the very high shear loading, the fact that there were actual gaps present, as acknowledged by Mr Hooper, means that they were not effective as stoppers and the whole design intent is not achieved rendering the potential of slippage and fatigue of the bolts at the Sauer Danfoss gearbox mounting, a real potential problem as highlighted by Dr Eccles.
[82] I turn next to consider the evidence of Mr Drago that the design of the interface between the Sauer Danfoss gearbox and the pinion was defective. As can be seen in the diagram below “Figure 4 – Drive Unit to Final Pinion Assembly and Pinion Mounting” reproduced from the AEIC of Mr Drago, the final drive pinion flange “Arrow 3” on the pinion gear “C” is mounted to the output flange “Arrow 2” of the Sauer Danfoss gearbox “B” by a series of bolts.
[83] Mr Drago explained at p 65 of his AEIC that:
[84] Mr Hooper attempted to demonstrate that the presence of a spigot on the flange of the pinion was sufficient to cure the problem. However, the following discourse at the hot-tubbing eventually led to Mr Hooper conceding that the spigot itself had a clearance of a maximum of 0.2 mm and with such a clearance, there was definitely allowance for some movement:
[85] I note that Mr Hooper smiled and hence, tacitly acknowledged that the problem of alignment to achieve concentricity alluded to earlier by Mr Drago remained because of the clearance at the spigot. Accordingly, the design does not appear to be adequate.
[86] I agree with the conclusion of Dr Eccles and Mr Drago that as a result of the potential movement arising from the 0.2 mm clearance at the spigot, the potential misalignment in concentricity at the interface between the Sauer Danfoss gearbox and the pinion flanges is unacceptably large for the bolts and the gears. After considering the totality of the evidence of the experts, I find that the bolting arrangements on the mounting of the Sauer Danfoss gearbox and the pinion/gearbox flange interface were both inadequately designed.
Bolting arrangement on the shaft end plates
[87] The shaft end plates provide attachment of both the main gear and the hub drive insert to the main driveshaft between the gear and the reel hub. Ten M20 x 55 mm, grade 8.8 bolts attach the plate to the hub/gear, with a further ten bolts of the same specification attaching the same plate to the shaft.
[88] Dr Eccles explained in his AEIC from pp 9 to 33 why the bolting arrangements on the shaft end plates were inadequate. At the hot-tubbing, Dr Eccles calculated that if the brakes were applied to slow down the 300 tons reel spinning at 0.29 rpm, corresponding to a lay speed of 500 m per hour, it would cause the bolts at the drive shaft end plates to be overloaded. With the clearances present, measuring about 1.3mm to a maximum of 2.5mm, between the hexagonal shaft and the gear on the one end and the reel insert on the other, the bolts would come loose. Mr Hooper disagreed with Dr Eccles and asserted that the hexagon in the hexagonal shaft took the shear load/torque transmission and not the bolts which merely held the gear hub in place. However, the hexagonal shaft would be able to take the shear load only if the clearances were absent and the hexagon fitted tightly within the hub. Mr Hooper then speculated that “it may have been that on the end of the hexagon, there was a slight taper put on in order to get the hub to fit on...”, which then took up all the clearance on the hexagonal section.
[89] Mr Moore then pointed out that the design drawings made no mention whatsoever of any taper on the hexagonal portions of the drive shaft or any taper on the gear or the hub insert.
[90] Mr Hooper then offered another explanation that even if there were clearances and the bolts loosened and slip occurred, the hexagon would still take over to bear the torque loading and “the hexagon would wind up on the hexagon on the shaft and become very, very tight.”
[91] However, I note that the RDU is designed not only to rotate the 300 tons reel to pay out the umbilical, but it can also rotate the reel in the opposite direction to wind back or pull in the umbilical. There can also be a sudden emergency stopping of the reel, whilst it is rotating in the clockwise or anti-clockwise direction. Given the manner in which the reel can be operated, it is therefore wrong to assume that the hexagon will always “wind upon the hexagon on the shaft” and therefore, the bolts will not loosen thereafter because they no longer bear the torque loading:
[92] I accept the explanation of Mr Hooper that the hexagonal drive would be able to transmit the very large torques generated especially during the braking of the 300 tons reel without the need for any assistance from the bolts. Since the Defendant had put bolts in both the gear wheel as well as the shaft, these bolts were made to bear these large torques unnecessarily and they would slip and come loose because they would not be able to take the shear loads. The shaft end plate might well come loose and fall off. As this is a rotating part, the bolting arrangement for the shaft end plates does not appear to me to be a safe design from this point of view.
[93] The Defendant refers to the Plaintiff’s Multimedia Bundle/Photo set 001.pdf and points out that after the repairs were completed on the RDU, the photograph of the new shaft end plate installed shows only a single circle of bolts ie, those engaged into the shaft end. The Defendant then surprisingly asserts in its submission that the design intent from the onset was for the torque in the main shaft to be transmitted through the hexagon fit and not through the bolts. I do not believe that this was the design intent from the beginning. It is only with hindsight that the Defendant realised the problem in its original design and then redesigned the bolting arrangement to allow the hexagon fit to take the full torque load instead.
Gears were under-designed
[94] Mr Drago carried out a detailed analysis based both (a) on the full rated torque of the motor/gearbox of 61,200 Nm for the non-constant tension mode; and (b) on what he termed the “system’s rated constant-tension load…” of “…15 tons of pull at 9.2 meter diameter for the entire system” in order to assess the “Service Factor” for each of the seven fatigue life conditions for the pinion and gear of the RDU. Mr Drago summarised his results in the following table for the RDU Pinion and Gear in both the Constant Tension Mode and the Non-Constant Tension Mode at Figure 8 at p 52 of his AEIC:
[95] Mr Drago explained the significance of evaluating the Service Factor “defined as the non-dimensional ratio of the inherent load capacity of the gear set divided by the actual applied load” at p 51 of his AEIC:
[96] Mr Drago concluded at p 60 of his AEIC that:
[97] In any event, Mr Hooper basically accepted that the design of the gears was flawed:
[98] It is clear from the totality of the expert evidence that the gears for the RDU were under-designed. I also note that the gears are essential operating components of the RDU. Both of these are important factors in my consideration whether the RDU as a whole is of merchantable quality and fit for its purpose.
Braking system was under-designed
[99] The video presented at the opening of the Plaintiff’s case illustrates the dire consequences of a reel spinning out of control. Any reasonable designer will consider this to be a grave risk requiring careful, cogent and complete mitigation. The adequacy of the design of the braking system for the RDU is thus critical.
[100] However, the Defendant did not carry out any calculations for its design of the braking system:
[101] The design of the mechanical brakes appears to be flawed. The mechanical brakes do not allow for a gradual increase in braking force. They simply spring into place the moment the hydraulic pressure is dropped below a particular pressure.
[102] It is not known specifically what that pressure is. Mr Hooper’s evidence was that the brakes would be applied when the pressure dropped to 500 psi or below:
[103] Mr Hooper then explained that the mechanical brake was simply designed as a parking brake and was not intended to stop a rotating 300 tons reel:
[104] I note however that there is no direct evidence to suggest that this was in fact the design intent of the Defendant. More importantly, Mr Hooper overlooked the fact that there is already an emergency stop function provided on the control panel of the RDU, which drops the hydraulic pressure to zero, and the mechanical brakes are immediately engaged to stop a rotating reel. Mr Moore said:
[105] The Plaintiff submits that absent a designed safe stopping process, pressing the emergency stop button would cause the RDU to self-destruct. This is a highly unsafe design. If the Defendant had tested the emergency stop function with an actual 300 tons rotating reel, as required by the ABS Guide for Certification of Lifting Appliances (see Chapter 2, Section 5 Testing of Cranes), this deficiency in the braking system would have become immediately apparent during testing. Such a test was however not performed.
[106] No parameters whatsoever were provided in the operations manual as to the appropriate angular decelerations in order to stop the 300 tons rotating reel. All the manual says is this:
[107] The Plaintiff submits that this is woefully insufficient. A reasonable designer would have provided specific operational parameters, and provided ways of ascertaining the angular velocity of the reel. Indeed, the design of the operations console makes it next to impossible for an operator to know the angular velocity of the reel.
[108] From what Mr Hooper had described, I find that the design flaws in braking system create obvious and serious risks:
[109] Mr Hooper responded that the Defendant had “trained the operators on how to use the equipment”. He quoted an extract from the manual (see [106] above) and said that the manual has clear instructions on how to operate the RDU.
[110] I agree with the Plaintiff’s submissions that:
(a) Mitigating a risk of this nature by training alone is insufficient. These deficiencies ought to have been designed out, and not left subject to human error.
(b) In any event, this does not answer the point that the mechanical brakes are also supposed to be able to stop the rotating 300 tons reel in an emergency. It makes no engineering sense to include as an emergency braking system a brake which cannot be safely operated when the reel mounted on the RDU is still rotating. Put another way, an emergency (stopping) brake which can only be operated when the reel is already stopped is quite pointless.
(c) Finally, the Defendant failed to provide any documents setting out the content of the training given to the Plaintiff’s operators which provides the evidential basis for Mr Hooper’s assertion that the operators were trained in respect of the proper operation of the RDU, for instance, training on the intricacies of how to safely apply the emergency brake, how to avoid excessive angular deceleration of the reel (with maximum deceleration values being stipulated) and how to control the lever in the constant tension mode to avoid any accidental knocking of that lever to the neutral or inadvertent movement of the lever back to neutral, which will abruptly apply the brakes on the rotating reel with disastrous consequences.
[111] Another aspect of concern is the capacity of the brakes themselves. When Dr Eccles was informed by Mr Hooper during the hot-tubbing that the maximum torque capacity of each mechanical brake was approximately 1.5 times the motor torque, he opined that the bolts and gears would be overloaded by the braking torque. Dr Eccles pointed out that there would be a tremendous amount of energy in the system of some 2.9 MJ at the reel’s maximum operating speed of 1.9 rpm and if the brakes were to absorb that amount of energy, they could well burn out and break:
[112] I note that Mr Hooper did not dispute what Dr Eccles had stated above. Although during the hot-tubbing, the manufacturer’s specifications of the mechanical braking capacity of the system became of much interest and Mr Hooper said that the Defendant was arranging to obtain it, the Defendant subsequently did not furnish the manufacturer’s specifications for the brakes to prove that the brakes have the capacity to absorb at least 2.9 MJ of energy present in a 300 tons reel rotating at 1.9 rpm. The Plaintiff accordingly submits that the proper inference is that the brakes are themselves unfit for purpose.
[113] I therefore conclude on a balance of probabilities that the braking system as a whole was improperly designed. Indeed, as Mr Moore had pointed out, the Sauer Danfoss drive system and gearbox is designed to be utilised on a road-going cement mixer truck, an entirely different application from a 300 tons offshore RDU which the Defendant had adapted it for. In the circumstances, it is clear to me that the brakes designed for the RDU were deficient:
[114] The Plaintiff rightly highlights that the Defendant had not tested the mechanical brakes in its intended application ie, in safely stopping a rotating 300 tons reel. All that was effectively tested during the FATs was the brakes’ capacity to act as a parking brake when a 15 tonnes line tension was applied at 9.2 m diameter. Thus the Defendant cannot rely on the successful FAT test of the brake as a mere parking brake to show that the braking system has been designed to safely stop a rotating 300 tons reel.
Sub-frame Bearing Housing was under-designed
[115] Mr Moore examined the fabricated sub-frame structure of the RDU that housed the bearings for the pinion gears, the main gear and the drive shaft and opined that they were inappropriately designed. Accordingly to Mr Moore, the relatively thin braces and the I–section beams within the structure provided little in the way of stiffening and torsional stability. The CAD model below shows the internal construction:
[116] Mr Moore noted that the RDU’s bearing housings were being supported and located by the bolts. In his view, heavy-duty fabricated structures are normally used to support such a drivetrain. The bearing housing is usually designed to be an integral part of the fabrication. This negates the need to rely on the bolts to attach the bearing housing to the main fabrication. In such cases, the item would be designed and manufactured as a casting or heavy-walled fabrication and would also incorporate strengthening members as part of its design. This ensures sufficient local rigidity and support of the loading experienced by the pinion bearing and provides a means of load transfer into the sub-frame fabrication which in turn transfers its load into the main support towers. Further, the bearing is normally located directly into the heavy walled structure via a very tight clearance fit. This provides a rigid base to seat and support the bearings whilst under load and provides an accurate machined datum, consequently giving better alignment of the bearing to the outboard mounting of the pinion. A similar arrangement is also appropriate for the bearing housing for the main driveshaft. This is critical to ensure correct alignment of the pinion gear to the spur gear. A similar arrangement appropriate for the bearing housing for the main driveshaft is also critical to ensure correct alignment of the pinion gear to the spur gear. In his report, he showed the following examples of correctly designed bearing and gearbox housings.
[117] In the light of Mr Moore’s criticisms of the design from a conceptual perspective, Mr Hooper conceded that the whole bearing housing could be better designed and it was “definitely not an optimum housing arrangement.” In fact, the Defendant had shown no technical justification from departing from what one would normally expect for such a design. There are no calculations or engineering analyses on record in the Defendant’s possession that would support the sufficiency or structural robustness of the Defendant’s design. I therefore find that the Defendant had not given due consideration to whether the bearing housing was appropriately designed at the time it was designed.
[118] Thus the experts, Mr Huang and Mr Natarajan, had to separately perform a FEA Analysis to verify whether or not there was in fact an under-design of the sub-frame bearing housing. This will be discussed below under “Finite Element Analysis”.
[119] It suffices for me to state at this juncture that I reject Mr Huang’s FEA analysis that the sub-frame bearing housing is structurally adequate and fit for its purpose because he had not used a sufficiently fine mesh for his model and had modelled incorrectly the pinion shaft as fully rigid when it was not the case. I accept Mr Natarajan’s FEA as being more accurate as he had used a fine mesh analysis and modelled the pinion shaft as non-rigid. Mr Natarajan’s FEA analysis shows that the stresses within certain parts of the sub-frame exceed the yield strength of the steel used and more importantly, the sub-frame deflections (in the range of millimetres instead of microns) are simply too large to maintain gear to pinion alignment within acceptable tolerances, which are in the order of microns.
[120] Apart from the theoretical analysis, the practical tests and measurements performed on the sub-frame under actual loads also show deflections in the order of several millimetres, thus broadly validating the results of Mr Natarajan’s FEA analysis: see 6 AB 5067 to 5068; deflection videos in Plaintiff’s Multimedia Bundle at SANY0016; SANY0020; SANY0021.
[121] Accordingly, I find that the sub-frame bearing housing, which I regard as a key component in the RDU, was under-designed and did not have sufficient rigidity. I accept Mr Drago’s evidence that the drive gear system is thus made to operate under severe misalignment conditions. This is another cumulative factor to be taken into consideration in my assessment whether the RDU as a whole is of merchantable quality and fit for its purpose.
Bearing arrangement for the main shaft was under-designed
[122] Mr Moore took issue with the fact that Mr Hooper had adopted the wrong bearing separation distance for his calculations of the loads on the bearings for the main shaft. Mr Hooper accepted this was an error on his part. Mr Moore then produced a set of calculations at Exhibit P18 based on the correct bearing separation distances for various roll conditions and different safety factors having regard to the shock loads expected during the operation of the RDU. There was some debate as to whether the roll conditions were already incorporated within the three acceleration values that were adopted for the design.

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Airtrust (Hong Kong) Ltd v PH Hydraulics & Engineering Pte Ltd

Grounds of decision

Introduction

Sale and Purchase Agreement

Inadequate design of the RDU

Failure to take the vessel roll into account

Failure to design for fatigue

Failure to design for the effects of wind

Bolting arrangements were under-designed

Bolts securing the struts to the main tower frame

Bolting arrangement on the pinion bearing joint

Bolting arrangement on the mounting of the gearbox and the pinion/gearbox flange interface

Bolting arrangement on the shaft end plates

Gears were under-designed

Braking system was under-designed

Sub-frame Bearing Housing was under-designed

Bearing arrangement for the main shaft was under-designed