暖通翻译论文题目

提供一些英语专业翻译方向的论文题目，供参考。 商务合同英汉互译技巧英文品牌汉译知识对于翻译的重要性中英文化差异及其对英汉互译的消极影响英语广告中修辞手法的应用及其翻译 <<红楼梦>>金陵判词两种译文的比较及评析从红楼梦诗词翻译看翻译中的文化补偿关于李后主“虞美人”的3种英译本的鉴赏跨文化交际与商标翻译中式菜肴的命名与翻译浅谈英语电影片名的翻译英文电影片名的翻译策略英文化妆品广告之美学翻译数字在中西文化中的内涵差异及数字习语翻译初探浅析原语文本在目标语文本中文体的适应性英语习用语翻译中的等效性研究论语境在英汉翻译中的作用浅析英语动画片翻译的基本原则中英服饰广告的翻译论英汉翻译中语篇连贯的重要性论译者的风格与译风经济英语中的隐喻及其翻译从翻译的美学角度浅析旅游资料的中英译

学术堂提供了十五个好写的英语翻译论文题目，供大家参考：1、谈英语谚语的翻译2、谈英语幽默的翻译3、英语汉译技巧初探4、地方名胜古迹汉译英5、翻译中常见错误分析6、中英思维方式的差异对翻译的影响7、会话含义的推导与翻译8、词汇的文化内涵与翻译9、语境在翻译中的作用10、翻译技巧探索11、商标词翻译12、广告语言的翻译13、论英汉互译中的语义等值问题14、英汉文化差异对翻译的影响15、英汉谚语的理解和翻译

翻译方向的.英语习语方面的.觉得这方面的很有趣,并且材料也很好搜集,写作角度也很广.在写作过程中,无论是从中西方文化方面还是从语言方面确实都学到不少东西. 每个人的兴趣不一样,不知道你会不会对这方面的感兴趣.“论电影片名的翻译”，这个题目比较好写，答辩的时候老师听的也比较有兴趣，资料也不难找。楼主可以试试。合同翻译中的经济、法律术语翻译如果以后要做翻译工作，也非常实用

testing of an air-cycle refrigeration system for road transportAbstractThe environmental attractions of air-cycle refrigeration are considerable. Following a thermodynamic design analysis, an air-cycle demonstrator plant was constructed within the restricted physical envelope of an existing Thermo King SL200 trailer refrigeration unit. This unique plant operated satisfactorily, delivering sustainable cooling for refrigerated trailers using a completely natural and safe working fluid. The full load capacity of the air-cycle unit at −20 °C was 7,8 kW, 8% greater than the equivalent vapour-cycle unit, but the fuel consumption of the air-cycle plant was excessively high. However, at part load operation the disparity in fuel consumption dropped from approximately 200% to around 80%. The components used in the air-cycle demonstrator were not optimised and considerable potential exists for efficiency improvements, possibly to the point where the air-cycle system could rival the efficiency of the standard vapour-cycle system at part-load operation, which represents the biggest proportion of operating time for most : Air conditioner; Refrigerated transport; Thermodynamic cycle; Air; Centrifuge compressor; Turbine expander COP, NomenclaturePRCompressor or turbine pressure ratioTAHeat exchanger side A temperature (K)TBHeat exchanger side B temperature (K)TinletInlet temperature (K)ToutletOutlet temperature (K)ηcompCompressor isentropic efficiencyηturbTurbine isentropic efficiencyηheat exchangerHeat exchanger effectiveness1. IntroductionThe current legislative pressure on conventional refrigerants is well known. The reason why vapour-cycle refrigeration is preferred over air-cycle refrigeration is simply that in the great majority of cases vapour-cycle is the most energy efficient option. Consequently, as soon as alternative systems, such as non-HFC refrigerants or air-cycle systems are considered, the issue of increased energy consumption arises over legislation affecting HFC refrigerants and the desire to improve long-term system reliability led to the examination of the feasibility of an air-cycle system for refrigerated transport. With the support of Enterprise Ireland and Thermo King (Ireland), the authors undertook the design and construction of an air-cycle refrigeration demonstrator plant at LYIT and QUB. This was not the first time in recent years that air-cycle systems had been employed in transport. NormalAir Garrett developed and commercialised an air-cycle air conditioning pack that was fitted to high speed trains in Germany in the 90s. As part of an European funded programme, a range of applications for air-cycle refrigeration were investigated and several demonstrator plants were constructed. However, the authors are unaware of any other case where a self-contained air-cycle unit has been developed for the challenging application of trailer King decided that the demonstrator should be a trailer refrigeration unit, since those were the units with the largest refrigeration capacity but presented the greatest challenges with regard to physical packaging. Consequently, the main objective was to demonstrate that an air-cycle system could fit within the existing physical envelop and develop an equivalent level of cooling power to the existing vapour-cycle unit, but using only air as the working fluid. The salient performance specifications for the existing Thermo King SL200 vapour-cycle trailer refrigeration unit are listed .It was not the objective of the exercise to complete the design and development of a new refrigeration product that would be ready for manufacture. To limit the level of resources necessary, existing hardware was to be used where possible with the recognition that the efficiencies achieved would not be optimal. In practical terms, this meant using the chassis and panels for an existing SL200 unit along with the standard diesel engine and circulation fans. The turbomachinery used for compression and expansion was adapted from commercial . Thermodynamic modelling and design of the demonstrator plantThe thermodynamics of the air-cycle (or the reverse ‘Joule cycle’) are adequately presented in most thermodynamic textbooks and will not be repeated here. For anything other than the smallest flow rates, the most efficient machines available for the necessary compression and expansion processes are turbomachines. Considerations for the selection of turbomachinery for air-cycle refrigeration systems have been presented and discussed by Spence et al. [3]. a typical configuration of an air-cycle system, which is sometimes called the ‘boot-strap’ configuration. For mechanical convenience the compression process is divided into two stages, meaning that the turbine is not constrained to operate at the same speed as the primary compressor. Instead, the work recovered by the turbine during expansion is utilised in the secondary compressor. The two-stage compression also permits intercooling, which enhances the overall efficiency of the compression process. An ‘open system’ where the cold air is ejected directly into the cold space, removing the need for a heat exchanger in the cold space. In the interests of efficiency, the return air from the cold space is used to pre-cool the compressed air entering the turbine by means of a heat exchanger known as the ‘regenerator’ or the ‘recuperato ’. To support the design of the air-cycle demonstrator plant, and the selection of suitable components, a simple thermodynamic model of the air-cycle configuration shown in was developed. The compression and expansion processes were modelled using appropriate values of isentropic efficiency, as defined in heat exchange processes were modelled using values of heat exchanger effectiveness as defined in The model also made allowance for heat exchanger pressure drop. The system COP was determined from the ratio of the cooling power delivered to the power input to the primary compressor, as defined in illustrate air-cycle performance characteristics as determined from the thermodynamic model:illustrates the variation in air-cycle COP and expander outlet temperature over a range of cycle pressure ratios for a plant operating between −20 °C and +30 °C. The cycle pressure ratio is defined as the ratio of the maximum cycle pressure at secondary compressor outlet to the pressure at turbine outlet. For the ideal air-cycle, with no losses, the cycle COP increases with decreasing cycle pressure ratio and tends to infinity as the pressure ratio approaches unity. However, the introduction of real component efficiencies means that there is a definite peak value of COP that occurs at a certain pressure ratio for a particular cycle. However,illustrates, there is a broad range of pressure ratio and duty over which the system can be operated with only moderate variation of class of turbomachinery suitable for the demonstrator plant required speeds of around 50 000 rev/min. To simplify the mechanical arrangement and avoid the need for a high-speed electric motor, the two-stage compression system shown was adopted. The existing Thermo King SL200 chassis incorporated a substantial system of belts and pulleys to power circulation fans, which severely restricted the useful space available for mounting heat exchangers. A simple thermodynamic model was used to assess the influence of heat exchanger performance on the efficiency of the plant so that the best compromise could be developed show the impact of intercooler and aftercooler effectiveness and pressure loss on the COP of the proposed two-stage system in incorporated an intercooler between the two compression stages. By dispensing with the intercooler and its associated duct work a larger aftercooler could be accommodated with improved effectiveness and reduced pressure loss. Analysis suggested that the improved performance from a larger aftercooler could compensate for the loss of the the impact of the recuperator effectiveness on the COP of the plant, which is clearly more significant than that of the other heat exchangers. As well as boosting cycle efficiency, increased recuperator effectiveness also moves the peak COP to a lower overall system pressure ratio. The impact of pressure loss in the recuperator is the same as for the intercooler and aftercooler shown in. The model did not distinguish between pressure losses in different locations; it was only the sum of the pressure losses that was significant. Any pressure loss in connecting duct work and headers was also lumped together with the heat exchanger pressure loss and analysed as a block pressure specific cooling capacity of the air-cycle increases with system pressure ratio. Consequently, if a higher system pressure ratio was used the required cooling duty could be achieved with a smaller flow rate of air. shows the mass flow rate of air required to deliver 7,5 kW of cooling power for varying system pressure the demonstrator system was to be based on commercially available turbomachinery, it became important to choose a pressure ratio and flow rate that could be accommodated efficiently by some existing compressor and turbine rotors. and were based on efficiencies of 81 and 85% for compression and expansion, respectively. While such efficiencies are attainable with optimised designs, they would not be realised using compromised turbocharger components. For the design of the demonstrator plant efficiencies of 78 and 80% were assumed to be realistically attainable for compression and turbomachinery efficiencies corresponded to higher cycle pressure ratios and flow rates in order to achieve the target cooling duty. The cycle design point was also compromised to help heat exchanger performance. The pressure losses in duct work and heat exchangers increased in proportion with the square of flow velocity. Selecting a higher cycle pressure ratio corresponded to a lower mass flow rate and also increased density at inlet to the aftercooler heat exchanger. The combined effect was a decrease in the mean velocity in the heat exchanger, a decrease in the expected pressure losses in the heat exchanger and duct work, and an increase in the effectiveness of the heat exchanger. Consequently, a system pressure ratio higher than the value corresponding to peak COP was chosen in order to achieve acceptable heat exchanger performance within the available physical space. The below optimum performance of turbomachinery and heat exchanger components, coupled with excessive bearing losses, meant that the predicted COP of the overall system dropped to around 0,41. The system pressure ratio at the design point was 2,14 and the corresponding mass flow rate of air was 0,278 kg/ moving the design point beyond the pressure ratio for peak COP, it was anticipated that the demonstrator plant would yield good part-load performance since the COP would not fall as the pressure ratio was reduced. Also, operating at part-load corresponded to lower flow velocities and anticipated improvements in heat exchanger performance. Part-load operation was achieved by reducing the speed of the primary compressor, resulting in a decrease in both pressure and mass flow rate throughout the . Prime mover and primary compressorThe existing diesel engine was judged adequate to power the demonstrator plant. The standard engine was a four cylinder, water cooled diesel engine fitted with a centrifugal clutch and all necessary ancillaries and was controlled by a microprocessor the thermodynamic model, the pressure ratio for the primary compressor was 1,70. The centrifugal compressor required a shaft speed of around 55 000 rev/min. Other alternatives were evaluated for primary compression with the aim of obtaining a suitable device that operated at a lower speed. Other commercially available devices such as Roots blowers and rotary piston blowers were all excluded on the basis of poor one-off gearbox was designed and manufactured as part of the project to step-up the engine shaft speed to around 55 000 rev/min. The gearbox was a two stage, three shaft unit which mounted directly on the end of the diesel engine and was driven through the existing centrifugal . Cold air unitThe secondary compressor and the expansion turbine were mounted on the same shaft in a free rotating unit. The combination of the secondary compressor and the turbine was designated as the ‘Cold Air Unit’ (CAU). While the CAU was mechanically equivalent to a turbocharger, a standard turbocharger would not satisfy the aerodynamic requirements efficiently since the pressure ratios and inlet densities for both the compressor and the turbine were significantly different from any turbocharger installation. Consequently, both the secondary compressor and the turbine stage were specially chosen and developed to deliver suitable turbochargers use plain oil fed journal bearings, which are low-cost, reliable and provide effective damping of shaft vibrations. However, plain bearings dissipate a substantial amount of shaft power through viscous losses in the oil films. A plain bearing arrangement for the CAU was expected to absorb 2–3 kW of mechanical power, which represented around 25% of the anticipated turbine power. Also, the clearances in plain bearings require larger blade tip clearances for both the compressor and the turbine with a consequential efficiency penalty. Given the pressurised inlet to the secondary compressor, the limited thrust capacity of the plain bearing arrangement was also a concern. A CAU utilising high-speed ball bearings, or air bearings, was identified as a preferable arrangement to plain bearings. Benefits would include greatly reduced bearing power losses, reduced turbomachinery tip clearance losses and increased thrust load capacity. However, adequate resources were not available to design a special one-off high speed ball bearing system. Consequently, a standard turbocharger plain bearing system was secondary compressor stage was a standard turbocharger compressor selected for a pressure ratio of 1,264. Secondary compressor and turbine selection were linked because of the requirement to balance power and match the speed. Since most commercial turbines are sized for high temperature (and consequently low density) air at inlet, a special turbine stage was developed for the application. Cost considerations precluded the manufacture of a custom turbine rotor, so a commercially available rotor was used. The standard turbine rotor blade profile was substantially modified and vaned nozzles for turbine inlet were designed to match the modified rotor, in line with previous turbine investigations at QUB (Spence and Artt,). An exhaust diffuser was also incorporated into the turbine stage in order to improve turbine efficiency and to moderate the exhaust noise levels through reduced air velocity. The exhaust diffuser exited into a specially designed exhaust performance of the turbine stage was measured before the unit was incorporated into the complete demonstrator plant. The peak efficiency of the turbine was established at 81%.5. Heat exchangersDue to packaging constraints, the heat exchangers had to be specially designed with careful consideration being given to heat exchanger position and header geometry in an attempt to achieve the best performance from the heat exchangers. Tube and fin aluminium heat exchangers, similar to those used in automotive intercooler applications, were chosen primarily because they could be produced on a ‘one-off’ basis at a reasonable cost. There were other heat exchanger technologies available that would have yielded better performance from the available volume, but high one-off production costs precluded their use in the demonstrator different tube and fin heat exchangers were tested and used to validate a computational model. Once validated, the model was used to assess a wide range of possible heat exchanger configurations that could fit within the Thermo King SL200 chassis. Fitting the proposed heat exchangers within the existing chassis and around the mechanical drive system for the circulation fans, but while still achieving the necessary heat exchanger performance was very challenging. It was clear that potential heat exchanger performance was being sacrificed through the choice of tube and fin construction and by the constraints of the layout of the existing SL200 chassis. The final selection comprised two separate aftercooler units, while the single recuperator was a large, triple pass unit. Based on laboratory tests and the heat exchanger model, the anticipated effectiveness of both the recuperator and aftercooler units was 80%.6. InstrumentationA range of conventional pressure and temperature instrumentation was installed on the air-cycle demonstrator plant. Air temperature and pressure was logged at inlet and outlet from each heat exchanger, compressor and the turbine. The speed of the primary compressor was determined from the speed measurement on the diesel engine control unit, while the cold air unit was equipped with a magnetic speed counter. No air flow measurement was included on the demonstrator plant. Instead, the air flow rate was deduced from the previously obtained turbine performance map using the measurements of turbine pressure ratio and rotational . System testingDuring some preliminary tests a heat load was applied and the functionality of the demonstrator plant was established. Having assessed that it was capable of delivering approximately the required performance, the plant was transported to a Thermo King calorimeter test facility specifically for measuring the performance of transport refrigeration units. The calorimeter was ideally suited for accurately measuring the refrigeration capacity of the air-cycle demonstrator plant. The calorimeter was operated according to standard ARI 1100-2001; the absolute accuracy was better than 200W and all auxiliary instrumentation was calibrated against appropriate performance capacity of transport refrigeration units is generally rated at two operating conditions; 0 and −20 °C, and both at an ambient temperature of +30 °C. Along with the specified operating conditions of 0 and −20 °C, a further part-load condition at −20 °C was assessed. Considering that the air-cycle plant was only intended to demonstrate a concept and that there were concerns about the reliability of the gearbox and the cold air unit thrust bearing, it was decided to operate the plant only as long as was necessary to obtain stabilised measurements at each operating point. The demonstrator plant operated satisfactorily, allowing sufficient measurements to be obtained at each of the three operating conditions. The recorded performance is summarised .In total, the unit operated for approximately 3 h during the course of the various tests. While the demonstrator plant operated adequately to allow measurements, some smoke from the oil system breather suggested that the thrust bearing of the CAU was heavily overloaded and would fail, as had been anticipated at the design stage. Testing was concluded in case the bearing failed completely causing the destruction of the entire CAU. There was no evidence of any gearbox deterioration during . Discussion of measured performanceFrom the calorimeter performance measurements, the primary objective of the project had been achieved. A unique air-cycle refrigeration system had been developed within the same physical envelope as the existing Thermo King SL200 refrigeration unit, w

标题1：如何利用辞典学习英语目的：找到利用辞典学习英语的方法。内容：1、调查；2、查资料；3、写作；4、修改。参考资料：1、国内期刊：《现代外语》、《外语教学》、《外语研究》；2、书籍；3、互联网。标题2：英语口语练习中的语音训练 目的：探索语音训练的方法以提高英语口语。内容：1、调查；2、查资料；3、写作；4、修改。参考资料：1、国内主要期刊：《中小学英语教学》、《中国教育》、《外语界》、《外语教学》等；2、书籍；3、互联网。标题3：英汉翻译中的“信”的原则目的：了解英汉翻译练习的“信”的本质并遵循此原则。内容：1、调查；2、查资料；3、写作；4、修改。参考资料：1、国内期刊：《现代外语》、《外语教学》、《外语研究》；2、书籍；3、互联网。标题4：幼儿英语教学中的智力开发目的：探索幼儿学习英语与智力开发的关系。内容：1、调查；2、查资料；3、写作；4、修改。参考资料：1、国内主要期刊：《中小学英语教学》、《中国教育》；2、书籍；3、互联网。标题5：新时代报刊阅读的特点目的：了解新时代报刊的新特点并探讨如何更好地阅读报刊。内容：1、调查；2、查资料；3、写作；4、修改。参考资料：1、国内期刊：《现代外语》、《外语教学》、《外语研究》；2、书籍；3、互联网。