看大学教材吧 《建筑环境与设备工程专业英语》
在网上找本暖通英汉词典就行了吧
20分帮你手动翻译那么多东西。。。。。。这个是不可能的，没事做也不要这样自寻烦恼啊
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
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《暖通空调》是本杂志，本专业的核心期刊，自己去随便下载几篇都是暖通行业相关的。 翻译借助百度翻译和自己那点水平，这小事儿还用求助..要原创的我可以提供
英文参考文献格式举例
参考文献是毕业论文的重要组成部分，对其进行统计分析，不仅有利于本科生的.教育和管理，而且能为图书馆文献保障和读者服务等工作提供一定的参考依据。下面是我整理的英文参考文献格式举例，希望大家重视。
一、参考文献的类型
参考文献(即引文出处)的类型以单字母方式标识，具体如下：M——专著C——论文集N——报纸文章J——期刊文章D——学位论文R——报告
对于不属于上述的文献类型，采用字母“Z”标识。对于英文参考文献，还应注意以下两点：
①作者姓名采用“姓在前名在后”原则，具体格式是：姓，名字的首字母.如：MalcolmRichardCowley应为：Cowley，.，如果有两位作者，第一位作者方式不变，&之后第二位作者名字的首字母放在前面，姓放在后面，如：FrankNorris与IrvingGordon应为：Norris，F.&.;
②书名、报刊名使用斜体字，如：MasteringEnglishLiterature，EnglishWeekly。
二、参考文献的格式及举例
1.期刊类
【格式】[序号]作者.篇名[J].刊名，出版年份，卷号(期号)：起止页码.【举例】
[1]王海粟.浅议会计信息披露模式[J].财政研究，2004，21(1)：56-58.[2]夏鲁惠.高等学校毕业论文教学情况调研报告[J].高等理科教育，2004(1):46-52.
[3]Heider，[J].ForeignLanguageTeachingandResearch，1999，(3):62–67.
2.专著类
【格式】[序号]作者.书名[M].出版地：出版社，出版年份：起止页码.【举例】[4]葛家澍，林志军.现代西方财务会计理论[M].厦门：厦门大学出版社，2001：42.
[5]Gill，[M].London:Macmillan，1985:42-45.
3.报纸类
【格式】[序号]作者.篇名[N].报纸名，出版日期(版次).【举例】
[6]李大伦.经济全球化的重要性[N].光明日报，1998-12-27(3).
[7]French，[N].AtlanticWeekly，198715(33).
4.论文集
【格式】[序号]作者.篇名[C].出版地：出版者，出版年份：起始页码.【举例】
[8]伍蠡甫.西方文论选[C].上海：上海译文出版社，1979：12-17.
[9]Spivak，G.“CantheSubalternSpeak?”[A].(eds.).VictoryinLimbo:Imigism[C].Urbana:UniversityofIllinoisPress，1988，.
暖通空调杂志就可以的
一般杂志投稿将Word电子文档发给杂志社就可以了，对格式没有太多要求，杂志社自己会排版的。
1. 去暖通空调的官网，. 点击左侧杂志订阅3. 按需选择购买
要给你个电话么 我们单位定着呢
《暖通空调》创刊于 1971 年，是中国建筑科学类核心期刊， 国家期刊奖最高奖项获奖期刊， 中国暖通空调行业惟一的中央级科技期刊，由建设部主管， 亚太建设科技信息研究院、 中国建筑设计研究院、 中国建筑学会（暖通空调专业委员会）联合主办。 本刊以实用技术为主，兼具学术性和信息性，在行业中最具影响力，被誉为权威刊物，深受广大读者喜爱，发行量在国内同行业刊物中遥遥领先。 《暖通空调》始终以 “ 新颖、实用、准确、精练 ” 为办刊方针，以提高全行业素质、推动全行业技术交流与发展为宗旨，及时报道国家有关建筑节能和环境保护的重大技术政策，建筑环境与设备工程中供暖、通风、空调、制冷及洁净技术方面的研究成果、学术论文、先进技术、工程总结、设计经验、设备开发与运行管理以及行业学术活动与设备市场信息。 《暖通空调》是世界最著名的建筑专业数据库 —— 国际建筑文献数据库 ICONDA 收录期刊，中国科技论文与引文数据库统计分析数据源刊，中国科学引文数据库来源期刊，中国学术期刊综合评价数据库统计源期刊，中国核心期刊（遴选）数据库收录期刊，中国期刊全文数据库收录期刊。 《暖通空调》栏目设置：专题研讨、科技综述、标准规范、专业论坛、专题讲座、设备开发、设计参考、工程实例、技术交流、运行管理。 《暖通空调》发行对象：从事建筑环境与设备工程中供暖、通风、空调、制冷、洁净等相关领域的工程设计、科研教学、施工安装、设备制造、运行管理的专业技术人员、管理人员、院校师生、房地产开发商和业主，以及对暖通空调制冷技术感兴趣的各界朋友。 编辑单位：《暖通空调资讯》编辑部总编：王曙明执行总编：潘晓福执行主编：刘昊编辑部地址：常州市新北区黄山路99-5号4楼
建筑热能通风空调、制冷与空调、城市建筑等等均可
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登陆他们的网站，点击个性化订阅。以下是联系方式《暖通空调》编辑部 地址： 北京市西城区德胜门外大街36号A楼4层 邮编： 100120电话： 010－88361727（主编） 010－88383814（杨爱丽） 010－68316357（刘学民） 010－57368821（于松波）010－68335394（李丽萍） 010－68330305（胡竹萍） 010－57368823（龚雪） 010－88362746（查询） 010－68363186（广告） 010－68362755（邮购）
《暖通空调》创刊于 1971 年，是中国建筑科学类核心期刊， 国家期刊奖最高奖项获奖期刊， 中国暖通空调行业惟一的中央级科技期刊，由建设部主管， 亚太建设科技信息研究院、 中国建筑设计研究院、 中国建筑学会（暖通空调专业委员会）联合主办。 本刊以实用技术为主，兼具学术性和信息性，在行业中最具影响力，被誉为权威刊物，深受广大读者喜爱，发行量在国内同行业刊物中遥遥领先。 《暖通空调》始终以 “ 新颖、实用、准确、精练 ” 为办刊方针，以提高全行业素质、推动全行业技术交流与发展为宗旨，及时报道国家有关建筑节能和环境保护的重大技术政策，建筑环境与设备工程中供暖、通风、空调、制冷及洁净技术方面的研究成果、学术论文、先进技术、工程总结、设计经验、设备开发与运行管理以及行业学术活动与设备市场信息。 《暖通空调》是世界最著名的建筑专业数据库 —— 国际建筑文献数据库 ICONDA 收录期刊，中国科技论文与引文数据库统计分析数据源刊，中国科学引文数据库来源期刊，中国学术期刊综合评价数据库统计源期刊，中国核心期刊（遴选）数据库收录期刊，中国期刊全文数据库收录期刊。 《暖通空调》栏目设置：专题研讨、科技综述、标准规范、专业论坛、专题讲座、设备开发、设计参考、工程实例、技术交流、运行管理。 《暖通空调》发行对象：从事建筑环境与设备工程中供暖、通风、空调、制冷、洁净等相关领域的工程设计、科研教学、施工安装、设备制造、运行管理的专业技术人员、管理人员、院校师生、房地产开发商和业主，以及对暖通空调制冷技术感兴趣的各界朋友。 编辑单位：《暖通空调资讯》编辑部总编：王曙明执行总编：潘晓福执行主编：刘昊编辑部地址：常州市新北区黄山路99-5号4楼
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《暖通空调》创刊于 1971 年，是中国建筑科学类核心期刊， 国家期刊奖最高奖项获奖期刊， 中国暖通空调行业惟一的中央级科技期刊，由建设部主管， 亚太建设科技信息研究院、 中国建筑设计研究院、 中国建筑学会（暖通空调专业委员会）联合主办。 本刊以实用技术为主，兼具学术性和信息性，在行业中最具影响力，被誉为权威刊物，深受广大读者喜爱，发行量在国内同行业刊物中遥遥领先。 《暖通空调》始终以 “ 新颖、实用、准确、精练 ” 为办刊方针，以提高全行业素质、推动全行业技术交流与发展为宗旨，及时报道国家有关建筑节能和环境保护的重大技术政策，建筑环境与设备工程中供暖、通风、空调、制冷及洁净技术方面的研究成果、学术论文、先进技术、工程总结、设计经验、设备开发与运行管理以及行业学术活动与设备市场信息。 《暖通空调》是世界最著名的建筑专业数据库 —— 国际建筑文献数据库 ICONDA 收录期刊，中国科技论文与引文数据库统计分析数据源刊，中国科学引文数据库来源期刊，中国学术期刊综合评价数据库统计源期刊，中国核心期刊（遴选）数据库收录期刊，中国期刊全文数据库收录期刊。 《暖通空调》栏目设置：专题研讨、科技综述、标准规范、专业论坛、专题讲座、设备开发、设计参考、工程实例、技术交流、运行管理。 《暖通空调》发行对象：从事建筑环境与设备工程中供暖、通风、空调、制冷、洁净等相关领域的工程设计、科研教学、施工安装、设备制造、运行管理的专业技术人员、管理人员、院校师生、房地产开发商和业主，以及对暖通空调制冷技术感兴趣的各界朋友。 编辑单位：《暖通空调资讯》编辑部总编：王曙明执行总编：潘晓福执行主编：刘昊编辑部地址：常州市新北区黄山路99-5号4楼
暖通空调 就是一个杂志啊或者其他自然学科的杂志