QTZ40塔式起重机总体及臂架优化设计【11张CAD图纸与说明书全套资料】
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河北建筑工程学院毕业设计（论文）外文资料翻译 系别： 机械工程系 专业： 机械设计制造及其自动化 班级： 机093 姓名： 侯延峰 学号： 2009307333 外文出处 Ocean Engineering 附 件：1、外文原文；2、外文资料翻译译文。指导教师评语：签字： 年 月 日注：请将该封面与附件装订成册。 海洋工程29（2002）14631477通过油水界面性能数值模拟的威尔斯涡轮机设计参数分析A.布里托-梅洛，L.M.C.加托，A.J.N.A.萨尔门托 机械工程系，西班牙优秀设计员，里斯本技术大学av. Rovisco Pais，1049-001里斯本，葡萄牙 2001年5月22日收到，于2001年8月30日接受 3.结果与讨论 试验研究不同类型的转子叶片，最近对这些叶片进行了管理来提高空气动力学性能（Raghunathan，1995年，加托，1999年a，b）。在这些类型当中，我们考虑两个涡轮叶片配置，这可能会提供一个更大范围的流量，在这个流量范围内涡轮机可以完全以不错的效率运行，与更多标准NACA0015无后掠叶片的涡轮机转子相比：他们是向后掠NACA0015叶片（韦伯斯特和加托，1999年）。图1所示，优化HSIM-15-262123-1576的未波及叶片（加托和恩里克斯，1996年），如图2所示。为了进行比较，我们采取NACA0015无后掠叶片（加托等人，1996年）。图3和图4显示了在IST钻机上单向流动小规模测试的实验结果（韦伯斯特和加托，1999年a，b）。结果列于图3和图4，高坚固性的威尔斯涡轮机的转子（转子外半径为R0.295米，常数翼弦C125毫米，涡轮实度0.64，上文提到的叶片配备，有或无导流叶片。数据显示，在无量纲形式中，实验结果为效率与压力降相关，和转矩,及函数流量系数相关（是空气密度）。无导流叶片的涡轮机在图3中的结果显示，NACA001530无后掠转子在时具有较低的为0.583，NACA001530向后掠转子在时具有较低的为0.583，但没有表现出无后掠转子的转矩急剧下降。此外，在失速条件下，转子扫略的转矩在一个更高流量时变为负，而无后掠转子在时效率变负。无后掠HSIM装有叶片的转子显示出了与向后掠转子类似的hmax，但通过转子叶片产生的逐步失速的流量，在很宽的流速范围内显出更高效率失速的发生。【】图1，转子叶片掠角 【】 图2 NACA0015和HSIM15-262123-1576部分。图4表示配备相应的情节相同的涡轮机转子双列的导流叶片。如图4实验结果，导流叶片的使用为上述任何几何形状增加了hmax，即从0.583到0.706，0.551到0.613和0.553到0.669。此外，我们发现，使用导流叶片缩小了流量范围，在这个范围内涡轮机以正扭矩运转。【1469 图（a）（b）（c）】图3. 无后掠和30向后掠NACA0015和无后掠HSIM转子叶片涡轮机，无导流叶片：效率的测量值（a）中，压力降（b）和扭矩（c）针对流量系数。【1470 图（a）（b）（c）】图4. 无后掠和30向后掠NACA0015和无后掠HSIM转子叶片涡轮机，带导流叶片：效率的测量值（a）中，压力降（b）和扭矩（c）针对流量系数。表1总结了六个涡轮机的性能数据，其中和分别为最小和最大的流量系数，效率名义上是。然而，和给出了这样一个迹象，即当中压力和流率比例近于直线的范围时的操作范围。在上面的性能对比中，稳定的整体坚固性被假定为不同的涡轮机配置。表1中的结果显示出，转子叶片的几何形状对涡轮机的性能具有显著的影响。特别是与在 较窄的流量范围内有更高的峰值效率的涡轮机相比，一些转子的几何形状会给与相当宽的流量范围，在这个范围中，涡轮机可以高效运行。图5-7通过44个有代表性的波浪气候的亚速尔群岛网站的记录给出了平均电力输出，并考虑到各海况的发生频率设定的数值。结果给出了涡轮机特性K为几个值的额定功率。表2表示在哪些设计的不同类型的涡轮机转子及旁通压力释放阀的流量系数值。3.1 NACA0015无后掠转子叶片和导流叶片图5给出了研究NACA0015无后掠叶片的转子使用导流叶片的数值模拟的结果。图5还展示出了有或者无庞统压力溢流阀，导流叶片的使用都会使平均电功率输出有显著的增加。图3和图4中绘制的NACA0015有或者无导流叶片的转子的曲线图，分别显示出带导流叶片的涡轮机具有最大效率.无导流叶片的涡轮机的转矩曲线具有较宽的流量范围，超过这个范围涡轮机将以良好的效率运行。数值模拟的结果揭示了有用的导流叶片。此外，他们显示的是在上述条件下，涡轮机的空气动力学设计标准应该是最大化的涡轮机的峰值效率，即使可能导致相对于流速曲线有一个较窄范围的效率。此外，可以发现，使用导流叶片会导致涡轮机尺寸的一个小的增加，然而，这不应该构成重大的问题，因为涡轮机的成本只有整个工厂成本的一小部分。 【1471 表1】【1472 图（a）（b）（c）】图5（a）的结果还表明，不管导流叶片是否被承认电功率输出作为涡轮机的额定功率的功能的趋势是一样的，。图6显示出电功率输出的显著增加。被看做涡轮机额定功率增大到600KW。最大电功率输出已实现，在模拟中考虑涡轮机额定功率的范围。【1473 图（a）（b）】图6，向后掠的30NACA0015转子叶片涡轮机和无导流叶片（a）和无旁通阀（b）：作为涡轮机的特性函数K，涡轮机的额定功率为几个值。图5（b）给出了使用旁路泄压阀的情况下的数值模拟结果。在这样的条件下，当使用旁路泄流阀时，与相应的情况下相比，总的转换功率相当小。另外，转换后的电功率随涡轮功率增长至900kw，高于这个值时，转换后的功率将减小，因为小流量涡轮机损失和机械损失。转换的最大电力获得的常数k的值，类似于使用旁路溢流阀获得的。【1474 图（a）（b） 】图7中，HSIM转子叶片涡和无导流叶片（a）以及无旁通阀（b）：作为涡的特性函数k，涡的额定功率为几个值。【 1474 表2】3.2NACA0015横扫转子叶片，有或者无导流叶片 图6给出了研究使用导流叶片对NACA0015 30向后掠转子叶片的影响的数值模拟的结果。可以看出，导流叶片的使用不利于没有溢流阀的情况。这是由于与无导流叶片的涡轮机相比，装有导流叶片的涡在停滞流动条件下的性能最差（见图3和图4）。 如图5和图6比较所示，装有横扫叶片转子的涡比无后掠叶片涡的性能差，假定该导流叶片和压力释放阀都被使用。这个结果与图4所示的相同的涡的性能曲线相符。 当考虑使用溢流阀且无导流叶片时，向后掠NACA0015叶片的转子与无后掠NACA0015转子相比，表明向后掠的转子叶片比相应的无后掠的转子生产的能量少。在这些条件下，与向后掠叶片相比，无后掠转子具有比失速之前效率更高的优点。然而向后掠叶片的好处是，目前向后掠叶片没有旁路压力释放阀且不使用导流叶片。 对于一个给定的涡的额定功率，电能的转换作为涡特性k的一个功能，这两种类型的涡是相似的，既当最高转换达到195k119.4Pam3S时，无论是否使用旁路泄压阀都一样。 此外，可以看出，旁通阀的使用与向后掠叶片转子的配合不会导致一个显著增加的平均功率转换，与所发现的相应条件下无后掠叶片涡轮机相反。这是由于在较宽的流速范围内，向后掠叶片转子的涡可以以良好的效益工作。3.3 HSIM转子叶片和无导流叶片图7的结果显示，无后掠HSIM叶片转子在使用旁通阀（图7a）和不使用旁通阀时（图7b），有或无导流叶片相比较，可以看出，当无导流叶片的涡的旁路泄压阀运行时，所产生的电能稍高，可以看出，当无导流叶片的涡的旁路泄压阀运行时，所产生的电能稍高。在涡带导流叶片的条件下也可以发现使用旁路压力泄压阀可以使电功率转换有较大幅度的提高。当HSIM叶片转子与NACA0015（图5）比较时，在无导流叶片和带有旁通阀这两种情况下，我们观察到NACA0015装有叶片的转子的涡可以获得更好的性能。这意味着与HSIM叶片转子的下部较宽的曲线相比应优选高峰窄的NACA0015叶片转子的效率曲线。 图4表明，当使用旁通阀时无后掠NACA0015叶片转子比HSIM叶片转子的转矩曲线更有利。与此相反目前如果没有旁通阀是事实。虽然在流速范围更宽时HSIM转子叶片具有正的转矩，有或者无导流叶片，其结果与NACA转子叶片相比表明，如果使用压力释放阀，至少波气候在亚速尔群岛是不利的。如果不考虑旁通阀，可以通过使用HSIM叶片得到稍微好一点的性能。我们注意到，HSIM叶片会导致更高的额定功率和尺寸更小的涡轮机，然而如果考虑到电动发电机和其电力电子的成本，这可能不会是一个优势。4. 结论通过波线理论来模拟电源转换链的数学模型可以用来匹配多个威尔斯涡轮机油水界面波能量转换器。主要关注的是威尔斯涡轮机的峰值效率与涡轮机可以高效运转的流率范围的宽度（固有有限），尤其是考虑在油水界面工厂中结合使用一个旁路压力释放阀。旁路高压溢流阀被发现为在类似的工作条件下研究的每个涡轮设计，以提供更高级别的电能生产。这一增长对无后掠NACA0015带有导流叶片的转子尤为重要。这个涡轮机为所有研究的涡轮机配置提供最佳的电源转换。无后掠NACA0015无导流叶片，且无旁路减压阀的转子表现最差。数学模型预测这种情况下，最大的平均电能生产只有60%左右，这样可以实现最好的安排，即无后掠NACA0015导流叶片与压力释放阀并行工作的转子。当不考虑溢流阀时，无后掠HSIM带或不带导流叶片的转子提供了最佳的功率转换。旁通压力释放阀的使用已经被发现用来减小涡轮机的尺寸和额定功率。当旁通溢流阀被应用于油水界面的控制时，涡轮机的设计应着眼于涡轮机的峰值效率最大化，即使这会导致一个较窄的效率曲线。相反，当没有考虑旁通阀时，计算结果表明，涡轮机的空气动力学设计会导致涡轮机能够在一个很宽的流速范围内以相当不错的效率运转。致谢根据IDMEC和MAQETEC的合同BD/5148/95，部分财政支持的计划PRAXIXXI，里斯本的工作报告。Ocean Engineering 29 (2002) 14631477/locate/oceanengAnalysis of Wells turbine design parameters bynumerical simulation of the OWC performanceA. Brito-Melo, L.M.C. Gato*, A.J.N.A. SarmentoMechanical Engineering Department, Instituto Superior Te cnico, Technical University of Lisbon, Av.Rovisco Pais, 1049-001 Lisbon, PortugalReceived 22 May 2001; accepted 30 August 2001AbstractThis paper investigates by numerical simulation the influence of the Wells turbine aerody-namic design on the overall plant performance, as affected by the turbine peak efficiency andthe range of flow rates within which the turbine can operate efficiently. The problem of match-ing the turbine to an oscillating water column (OWC) is illustrated by taking the wave climateand the OWC of the Azores power converter. The study was performed using a time-domainmathematical model based on linear water wave theory and on model experiments in a wavetank. Results are presented of numerical simulations considering several aerodynamic designsof the Wells turbine, with and without guide vanes, and with the use of a bypass pressure-relief valve. 2002 Elsevier Science Ltd. All rights reserved.Keywords: Wave energy; Oscillating water column; Equipment; Wells turbine1. IntroductionThe Wells turbine has been the most commonly adopted solution to the air-to-electricity energy conversion problem in oscillating water column (OWC) waveenergy converters. These essentially consist of a capture pneumatic chamber, openat the bottom front to the incident wave, a turbine and an electrical generator. Theincident wave motion excites the oscillation of the internal free surface of theentrained water mass in the pneumatic chamber, which produces a low-pressure reci-* Corresponding author. Tel.: +351-21-841-7411; fax: +351-21-841-7398.E-mail address: lgatohidro1.ist.utl.pt (L.M.C. Gato).0029-8018/02/$ - see front matter 2002 Elsevier Science Ltd. All rights reserved.PII: S0029-8018(01)00099-31464A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477procating flow that drives the turbine. A few full-scale turbine prototypes have beenbuilt and installed in grid-connected power plants in European countries, e.g. the500 kW Wells monoplane turbine with guide vanes installed in the Island of Pico,Azores (Falca o, 2000), and 2250 kW biplane contrarotating turbine of the LIMPETplant, at Islay, Scotland (Heath et al., 2000).The greatest challenges to designers of equipment for wave energy converters arethe intrinsically oscillating nature and the random distribution of the wave energyresource. These features are absent or much less severe in other competing energytechnologies. The air turbine in an OWC converter is subject to flow conditions(randomly reciprocating flow), which, with respect to efficiency, are much moredemanding than in turbines in almost any other application. The Wells turbine, whilereaching only a moderately high peak efficiency as compared with conventional tur-bines, can operate in reciprocating flow without the need of a rectifying valve system.The turbine, on the one hand, is required to extract energy from air whose flow rate,in each of the two directions, oscillates between zero and a maximum value, whichin turn has an extremely large variation from wave to wave and with sea conditions.On the other hand, at fixed rotational speed, turbines in general, and Wells turbinesin particular, are capable of operating with good efficiency only within a limitedrange of flow conditions around the peak efficiency point. The power output of Wellsturbines is known to be low (or even negative) at small flow rates (the flow ratepasses through zero twice in a wave cycle) and it drops sharply for flow rates abovea critical value due to aerodynamic losses produced by rotor blade stalling. Therefore,the turbine is expected to perform poorly in very energetic sea-states or wheneverviolent wave peaks occur. Mounting a bypass pressure-relief valve on the top of theair chamber as proposed in the Azores plant may prevent this problem. The valveis controlled to limit the maximum pressure and suction in the chamber (dependingon the turbine rotational speed) to prevent the instantaneous air flow rate throughthe turbine from exceeding the values above which aerodynamic stalling at the rotorblades would produce a severe fall in power output. Numerical simulations (Brito-Melo et al., 1996; Falca o and Justino, 1999) indicate that a reduction in turbine sizeand a substantial increase in the annual production of electrical energy might beachieved by the use of a bypass pressure-relief valve. Moreover, recent studies(theoretical and model testing) indicate that blade sections especially designed forWells turbine rotors can significantly enlarge the range of flow rates within whichthe turbine operates efficiently and reduce aerodynamic losses under partially stalledflow conditions, in comparison with other blade designs which give a higher peakefficiency within a narrower range of flow rates through the turbine. This raises thequestion of whether, in view of the total annual produced electrical energy and takinginto account the hydrodynamic performance of the OWC device, it is more appropri-ate to select a turbine aerodynamic design which allows an enlarged range of flowrates at which the turbine operates efficiently or whether it is better to adopt a turbinedesign which gives a higher peak efficiency value with a reduced range of flow ratesat which the turbine operates efficiently. Furthermore, it is of interest to know towhat extent this issue might be dependent on the use of a pressure-relief valve.The main objective of the present work is to investigate the influence of the Wells1465A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477turbine aerodynamic design on the overall plant performance, as affected by theturbine peak efficiency and the range of flow rates within which the turbine canoperate efficiently. Realistic characteristics are assumed for the turbine and the useof a bypass pressure-relief valve is also considered. Since the resulting pressurechanges in the chamber are dependent on the turbine characteristics and the pressure-relief valve influences the turbine operation, the hydrodynamic process of energyextraction is also modified. The hydrodynamics of the conversion of wave energyinto pneumatic energy is predicted by using a time-domain mathematical modelbased on linear water wave theory and on model experiments in a wave tank asdescribed in Sarmento and Brito-Melo (1996). The conversion of pneumatic energyinto electrical energy is predicted by a suitable computational model of the powertake-off equipment based on the results extrapolated from aerodynamic tests on ascale-model and on empirical approximations for the generator losses (Brito-Meloet al., 1996). This paper presents the results of numerical simulations consideringseveral aerodynamic designs of the Wells turbine, with and without guide vanes,and the use of the pressure-relief valve. The problem of matching the turbine to anOWC is illustrated by taking the wave climate and the OWC of the Azores wavepower converter.2. Wave-to-wire model2.1. Plant operationThe wave-to-wire model concerns the operation of an OWC equipped with a Wellsturbine, a bypass valve of unlimited capacity and a variable speed turbo-generator,under a set of representative sea-state conditions.The Wells turbine is known to exhibit an approximately linear relationshipbetween the turbine pressure drop p(t) and the flow rate qt(t). Then we may writethe turbine characteristic as K ? p(t)/qt(t) ? ps(?)/qs(?), where ps(?), and qs(?)are maximum values of pressure and flow rate (prior to the onset of aerodynamicstall at the turbine rotor blades), which (for a given turbine) depend on the turbinerotational speed ?. The use of a properly controlled bypass pressure-relief valveprevents the occurrence of stall at the turbine rotor blades. The valve is controlledto ensure that |p(t)|?ps(?). Then |qs(t)|?qs(?). The excess flow rate qv(t) passesthrough the valve to (or from) the atmosphere.The inertia of the rotating parts is assumed large enough so that rotational speed? may be considered approximately constant over the time-intervals simulating agiven sea-state (about 15 minutes). This allows ? to be optimized for each represen-tative record of the sea-state, in order to maximize the electrical energy production.The turbine rotational speed is allowed to vary between the synchronous speed ofthe generator and twice its value. Summing the product of the time-averaged electri-cal power output with the occurrence frequency for all data records gives the overallannual average electrical power output.1466A. Brito-Melo et al. / Ocean Engineering 29 (2002) 146314772.2. Hydrodynamic modelThe hydrodynamic model is based on the pressure model presented in Sarmentoand Falca o (1985). According to the OWC performance description presented inSection 2.1, the mass balance across a control surface enclosing the pneumaticchamber is given byp(t)K? qv(t) ? q(t)?V0gPadp(t)dt(1)where q(t) is the volume flow rate displaced by the free-surface inside the chamber,V0denotes the volume of the air in the chamber under undisturbed conditions, Paisthe atmospheric pressure and g is the ratio of specific heats. As stated in Section2.1, qv(t) ? 0 if |p(t)| ? ps(?) (i.e. when the valve is not operating). According tothe linear water wave theory, the volume flow rate displaced by the free-surfaceinside the chamber may be decomposed as q(t) ? qd(t) ? qr(t), where qd(t) is thediffraction flow rate, due to incident wave action assuming the internal and the exter-nal free-surfaces at constant atmospheric pressure, and qr(t) is the radiation flow ratedue only to the pressure oscillation p(t) in otherwise calm waters. Under the assump-tions of the linearized wave theory, we may apply the convolution theorem to obtainthe solution of a linear problem in terms of an impulse response (Pipes and Harvill,1970), as follows:qr(t) ?thr(t?t)p?(t) dt(2)where p?(t) is the time-derivative of the pressure inside the chamber and t representsa time-lag. The upper limit of the integral in Eq. (2) represents the present instantt because the process is causal (Cummins, 1962). The impulse response functionhr(t) can be obtained from the hydrodynamic coefficients of the OWC computedwith a numerical model, such as the WAMIT (Lee et al., 1996) or the AQUADYN-OWC (Brito-Melo et al., 1999), or by tank testing. Here we use an estimate of theimpulse response function obtained in free-oscillation transient experiments from1:35 scale testing of the Azores OWC wave power plant (see Sarmento and Brito-Melo, 1996, for details).Time series for the diffraction flow, qd(t), have also been obtained in energy extrac-tion experiments with the scaled model subject to a set of 44 sea-states representativeof the Azores power plant site. In these experiments a device producing an equivalentair pressure drop simulated the turbine. The flow rate qt(t) could be obtained as afunction of p(t) from the device calibration curve. The diffraction flow time-seriesfor each of the 44 sea-states was estimated by solving Eq. (1) (with qv(t) ? 0) usingthe pressure records from the energy extraction experiments, and the experimentalestimate of hr(t) previously obtained in the transient experiments.1467A. Brito-Melo et al. / Ocean Engineering 29 (2002) 146314772.3. Power take-off equipmentThe power take-off sub-model is based on results extrapolated from small-scaleturbine tests (Gato et al., 1996; Webster and Gato, 1999a,b) and on empirical datafor the turbine and generator losses (Brito-Melo et al., 1996). The average power atthe turbine shaft for a period T is given byWs?T?0TL(?,qt(t)?Lm(?) dt(3)where L is the aerodynamically produced turbine-torque and Lmthe torque due tomechanical losses (especially bearing losses). For stall-free conditions, L is approxi-mated by a second-order polynomial. In order to provide the necessary performancedata to study the matching of the power take-off equipment and the pneumaticchamber, the data from small-scale turbine tests are modified using a simple mean-line turbine flow analysis method to take into account the rotor solidity S and thehub-to-tip ratio. Ignoring the postponement of stall when the Reynolds number isincreased, scale effects are taken into account by correcting the torque curve of theturbine model. This is done multiplying (dividing) the positive (negative) values ofL by f ? 0.8/0.706. This corrects the torque curve of the unswept NACA 0015bladed rotor with guide-vanes to match a peak efficiency of hmax? 0.80. For thepreliminary design of the turbine a maximum blade tip speed of 160 ms?1is assumed.The average electrical power output is obtained by subtracting the generator lossesfrom the average power at the turbine shaft. The model for the generator lossesincludes the Joule losses, the iron losses, the ventilation losses and the mechanicallosses (Brito-Melo et al., 1996).3. Results and discussionExperimental research on different types of rotor blades has been conductedrecently to improve the aerodynamic performance of the Wells turbine (Raghunathan,1995; Gato et al., 1996; Curran and Gato, 1997; Webster and Gato, 1999a,b). Amongthese types, we consider two turbine blade configurations, which may give a widerrange of flow rates within which the turbine can operate with fairly good efficiency,in comparison with that of the more standard NACA 0015 unswept bladed turbinerotor: they are the backward-swept NACA 0015 blades (Webster and Gato, 1999a),Fig. 1, and the optimized HSIM-15-262123-1576 unswept blades (Gato and Hen-riques, 1996), Fig. 2. For comparison we take results for the NACA 0015 unsweptblades (Gato et al., 1996).Figs. 3 and 4 show experimental results from unidirectional-flow small-scale test-ing at the IST rig (Webster and Gato, 1999a,b). Results presented in Figs. 3 and 4refer to high-solidity Wells turbine rotors (rotor outer radius R ? 0.295 m, constantchord c ? 125 mm, rotor solidity S ? 0.64, equipped with the blades referred to1468A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477Fig. 1.Rotor blade sweep angle.Fig. 2.The NACA 0015 and HSIM 15-262123-1576 sections.above, with and without guide vanes. The figures show, in dimensionless form,experimentalresultsfortheefficiencyh ? L?/(qtp),pressuredropp?p/(r?2R2), and torque L? L/(r?2R5) as functions of the flow rate coefficient U*(r is the air density). Results in Fig. 3 for the turbines without guide vanes showthat the NACA 0015 unswept rotor has hmax? 0.583 at U? 0.114, and stalls atU? 0.21. The NACA 0015 30 backward-swept rotor has a lower hmax? 0.583,with a lower flow rate for the onset of stall, U? 0.17, but without exhibiting thesharp decrease in the torque that occurs in the unswept rotor. Furthermore, understall conditions, the torque of the swept rotor becomes negative at a much higherflow rate, U? 0.45, whereas for the unswept blades the efficiency becomes nega-tive for U? 0.3. The unswept HSIM bladed rotor shows a hmaxsimilar to that ofthe backward-swept rotor, but produces a soft progressive stall of the flow throughthe rotor blades, with notably higher efficiency for a wide range of flow rates afterthe onset of stall.Fig. 4 shows a corresponding plot for the same turbine rotors when equipped witha double row of guide vanes. The experimental results plotted in Fig. 4 show thatthe use of guide vanes increases hmaxfor any of the above geometries, i.e. from0.583 to 0.706, 0.551 to 0.613 and 0.553 to 0.669, for the NACA 0015 unswept and1469A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477Fig. 3.Unswept and 30 backward-swept NACA 0015 and unswept HSIM bladed rotor turbines, withoutguide vanes: measured values of efficiency (a), pressure drop (b) and torque (c) against flow rate coef-ficient.1470A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477Fig. 4.Unswept and 30 backward-swept NACA 0015 and unswept HSIM bladed rotor turbines, withguide vanes: measured values of efficiency (a), pressure drop (b) and torque (c) against flow rate coef-ficient.1471A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477backward-swept rotors and the HSIM unswept rotor, respectively. Furthermore, wefind that the use of guide vanes narrows the range of flow rates within which theturbine works with positive torque.Table 1 summarizes the performance data for the six turbines, where UaandUbare the minimum and maximum flow rate coefficients respectively, at which theefficiency is nominally h ? 0.5hmax. Therefore, ? ? Ua/Uband ? ? Ua?Ubgivean indication of the operational range while (?p0/U)h ? hmaxis the pressureflowratio in the approximately rectilinear region. In the above performance comparison,constant overall solidity was assumed for the different turbine configurations. Resultsin Table 1 show that the rotor blade geometry has a remarkable influence on theturbine performance. In particular, some rotor geometries give a considerable widerrange of flow rates within which the turbine operates efficiently, in comparison withothers that have higher peak efficiency within a narrower range of flow rates.Figs. 57 plot the average electrical power output as given by the numerical simul-ation for the set of the 44 representative records of the wave climate for the AzoresPlant site, taking into account the frequency of occurrence of each sea-state. Theresults give the turbine characteristic K for several values of the rated powerW0? psqs. Table 2 indicates the values of the flow coefficient Usat which thedifferent types of turbine rotor were designed and the bypass pressure-relief valveis actuated.3.1. NACA 0015 unswept bladed rotor with and without guide vanesFig. 5 presents the results of the numerical simulation to study the effect of theuse of guide vanes with the NACA 0015 unswept bladed rotor. Fig. 5 shows thatthe use of guide vanes provides a significant increase in the average electrical poweroutput, both with and without the presence of the bypass pressure-relief valve. Thecurves plotted in Figs. 3 and 4 for the unswept NACA 0015 rotor, with and withoutguide vanes, respectively, show that the turbine with guide vanes has hmax?0.72Table 1Peak efficiency, useful flow rate range and damping ratio for several turbine models (overall solidityS=0.64)Turbine rotorWith guide vanesWithout guide vanesNACA 0015 NACA 0015 HSIMNACA 0015 NACA 0015 HSIMunsweptswept-backunsweptunsweptswept-backunswepthmax0.7060.6130.6690.5830.5510.553(U)h ? hmax0.1240.1370.1540.1140.1290.131Ua0.0500.0620.0570.0510.0580.059Ub0.1970.2090.2750.2510.2320.360?0.2540.2970.2070.2030.2500.164?0.1470.1470.2180.2000.1740.301(?p0/U)h ? hmax2.191.872.382.542.042.791472A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477Fig. 5.Unswept NACA 0015 bladed rotor turbine with and without guide vanes working (a) with and(b) without the bypass valve: average electrical power conversion as a function of the turbine characteristicK, for several values of the turbine-rated power.whereas for the turbine without guide vanes hmax? 0.60. Nevertheless, the torquecurve for the turbine without guide vanes exhibits a wider range of flow rates overwhich the turbine performs with good efficiency. The results of the numerical simula-tions reveal the usefulness of the guide vanes. In addition, they show that, under theabove conditions, the aerodynamic design criterion for the turbine should be to max-imize the turbine peak efficiency even if that may result in a narrower curve ofefficiency versus flow rate. Furthermore, it may be found that the use of guide vanesleads to a small increase in the turbine size, which, however, should not constitutea significant penalty since the turbine cost is only a small fraction of the overallplant cost.Results in Fig. 5a also show that the trend for the electrical power output as afunction of the turbine-rated power is the same regardless of whether guide vanesare considered or not. A significant increase in the electrical power output is seen1473A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477Fig. 6.30 backward-swept NACA 0015 bladed rotor turbine with and without guide vanes working(a) with and (b) without the bypass valve: average electrical power conversion as a function of the turbinecharacteristic K, for several values of the turbine-rated power.to occur as the turbine-rated power increases up to 600 kW. Maximum electricalpower output is achieved for a turbine characteristic 95?K?119.4 Pa m?3s, withinthe range of turbine-rated powers considered in the simulations.Fig. 5b presents the results of the numerical simulation in the absence of a bypassrelief valve. Under such conditions, the total converted power is considerablysmaller, in comparison with the corresponding cases when a bypass relief valve ispresent. Furthermore, the converted electrical power increases with turbine-ratedpower up to 900 kW. Above this value, the converted power decreases due to theturbine losses at small flow rates and the increase in the electrical and mechanicallosses. Maximum converted power is obtained for a value of the constant K similarto that obtained using the bypass relief valve.1474A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477Fig. 7.HSIM bladed rotor turbine with and without guide vanes working (a) with and (b) without thebypass valve: average electrical power conversion as a function of the turbine characteristic K, for severalvalues of the turbine-rated power.Table 2Design value of the flow coefficient for which the bypass valve is actuatedTurbine rotorUsWith guide vanesWithout guide vanesNACA 0015 unswept0.1760.213NACA 0015 swept0.1700.178HSIM unswept0.2750.381475A. Brito-Melo et al. / Ocean Engineering 29 (2002) 146314773.2. NACA 0015 swept bladed rotor, with and without guide vanesFig. 6 presents the results of the numerical simulation to study the effect of theuse of guide vanes with the NACA 0015 30 backward-swept bladed rotor. It canbe seen that the use of guide vanes is not beneficial in the absence of the reliefvalve. This is due to the poorest performance of the guide vane equipped turbineunder stalled flow conditions, as compared with the turbine without guide vanes (seeFigs. 3 and 4).The comparison between Figs. 5 and 6 shows that the performance of the sweptbladed rotor is poorer than that of the unswept bladed rotor, assuming that guidevanes and a pressure-relief valve are used. This result agrees with the performancecurves shown in Fig. 4 for the same turbines.The performance comparison of the backward-swept NACA 0015 bladed rotorwith that of the unswept NACA 0015 rotor, both without guide vanes and consideringthe use of a relief valve, shows that the backward-swept rotor blades produce lessenergy in comparison with that from the corresponding unswept bladed rotor. Inthese conditions, the advantage of the unswept rotor arises from its higher efficiencyprior to stall, in comparison with that of the backward-swept blades. However, asmall benefit is obtained with the backward-swept blades if no bypass pressure-reliefvalve is present and no guide vanes are use.For a given turbine-rated power, the electrical energy conversion as a function ofthe turbine characteristic K is similar for both turbine types, i.e. maximum conversionis obtained when 95?K?119.4 Pa m?3s, regardless of the use or not of the bypassrelief valve.Furthermore, it can be seen that the use of the bypass valve in conjunction withthe backward-swept bladed rotor does not lead to a significant increase in the meanpower conversion, in contrast with what was found for the corresponding conditionsin the case of the unswept bladed turbine. This is due to the wider range of flowrates within which the backward-swept bladed rotor turbine can operate with fairlygood efficiency.3.3. HSIM bladed rotor with and without guide vanesResults presented in Fig. 7 compare the performance of the unswept HSIM bladedrotor with and without guide vanes in the presence (Fig. 7a) or absence (Fig. 7b) ofa bypass valve. It can be seen that the electrical energy produced is slightly higherwhen the bypass pressure-relief valve operates with the turbine without guide vanes.In the case of a turbine with guide vanes, it may be found that a more substantialincrease in the electrical power conversion is obtained by using a bypass pressure-relief valve.When comparing the HSIM bladed rotor with the NACA 0015 (Fig. 5), in bothcases without guide vanes and with a bypass valve, we observe that better perform-ance is obtained from the turbine with the NACA 0015 bladed rotor. This meansthat the higher-peaked narrower efficiency curve of the NACA 0015 bladed rotorshould be preferred to the lower wider curve for the HSIM bladed rotor.1476A. Brito-Melo et al. / Ocean Engineering 29 (2002) 14631477Fig. 4 shows that the torque curve of the unswept NACA 0015 bladed rotor ismore beneficial than that of the HSIM bladed rotor when a bypass valve is present.The opposite is true if no bypass valve is present. Although the HSIM bladed rotorhas positive torque over a wider range of flow rates than the NACA bladed rotor,with and without guide vanes, the results show that this is not a benefit if the press-ure-relief valve is used, at least for the wave climate at the Azores plant.If the bypass valve is not considered then a slightly better performance can beachieved using the HSIM blades. We note that the HSIM blades lead to higher ratedpower and smaller turbines, which, however, may not be an advantage if the costsof the electrical generator and its power electronics are taken into account.4. ConclusionA computational model that simulates the power conversion chain from the wave-to-wire was used to match several Wells turbine designs to an OWC wave energyconverter. A main concern was the Wells turbines peak efficiency versus the(inherent limited) width of the range of flow rates within which the turbine canoperate efficiently, especially if this is considered in conjunction with the use of abypass pressure-relief valve in an OWC plant.The bypass pressure-relief valve was found to provide higher levels of electricalenergy production for each of turbine designs studied under similar working con-ditions. This increase is more significant for the unswept NACA 0015 bladed rotorwith guide vanes. This turbine gives the best power conversion among all the turbineconfigurations studied.The poorest performance is provided by the unswept NACA 0015 bladed rotorwithout guide vanes and without the bypass pressure-relief valve. The mathematicalmodel predicts for this case a maximum average electrical power production that isonly about 60% of what can be achieved with the best arrangement, i.e. the unsweptNACA 0015 bladed rotor with guide vanes working in parallel with a pressure-relief valve.When the relief valve is not considered, the unswept HSIM bladed rotor gives thebest power conversion, with or without guide vanes.The use of the bypass pressure-relief valve was found to provide a reduction inthe turbine size and rated power.When the bypass pressure-relief valve is used in the control of an OWC, then theturbine design should aim at maximizing the turbine peak efficiency even if thatre
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