《节能》或者《制冷学报》或《城市建设理论研究》
暖通专业的论文，最好是发国家级或者核心期刊了，不过审核也相当严的，
（迪西欧论文网）可以为您提供写作和发表服务，保证您文章质量和个人的信息安全，详情可以咨询迪西欧论文网宋老师。希望对你有用，如果有帮助就采纳一下吧 ^_^ ^_^ 兄弟咋又发了一遍？我给你的答复没错的。呵呵。如果档案在重庆，那就肯定有资格评初级，而且不需要考英语和计算机
建筑热能通风空调、制冷与空调、城市建筑等等均可
擦，都到这来找了，一看就522的某个人
这种国内的会议论文都是很水的， 基本不算SCI， 除非被选中优秀论文，会在会议指定的一些期刊进行发布。就看那些期刊的级别了。
（一）暖通空调系统的基本组成一个完整独立的空调系统基本可分为三大部分，分别是：冷热源及空气处理设备、空气和冷热水输配系统、室内末端装置。图8-4是一个典型的空调系统组成示意图，夏季由制冷设备（冷源）提供冷水或液态制冷剂，冬季由锅炉（热源）提供热水或蒸汽。通过冷热水输配系统将冷热水送至空调机组（空气处理设备）将空气处理到送风状态点，通过空气输配系统将处理后的空气送入室内消除热湿负荷，或者将冷热水送至房间末端设备（空气处理设备）换热来满足房间负荷要求。局部处理方式A和集中处理方式B可以分别独立使用，也可以联合使用。 图8-4 空调系统组成示意图 （二）工作原理空调系统的工作原因主要是制冷原理，也就是逆卡诺循环。下面图为“卡诺循环”示意，逆卡诺循环为其相反循环，但原理是一样的。卡诺循环是由四个循环过程组成，两个绝热过程和两个等温过程。它是1824年NLS卡诺（见卡诺父子）在对热机的最大可能效率问题作理论研究时提出的。卡诺假设工作物质只与两个恒温热源交换热量，没有散热、漏气、磨擦等损耗。为使过程是准静态过程，工作物质从高温热源吸热应是无温度差的等温膨胀过程，同样，向低温热源放热应是等温压缩过程。因限制只与两热源交换热量，脱离热源后只能是绝热过程（三）主要的系统类型按使用目的分类舒适性空调——要求温度适宜，环境舒适，对温湿度的调节精度无严格要求、用于住房、办公室、影剧院、商场、体育馆、汽车、船舶、飞机等。工艺性空调——对温湿度有一定的调节精度要求，另外空气的洁净度也要有较高的要求。用于电子器件生产车间、精密仪器生产车间、计算机房、生物实验室等。2．按设备布置情况分类集中式（中央）空调——空气处理设备集中在中央空调室里，处理过的空气通过风管送至各房间的空调系统。适用于面积大、房间集中、各房间热湿负荷比较接近的场所选用，如商场、超市、餐厅、船舶、工厂等。系统维修管理方便，设备的消声隔振比较容易解决，但集中式空调系统的输配系统中风机、水泵的能耗较高。图8-4中，如果没有空气局部处理A，只有集中处理B来进行空气调节，此系统就属于集中式。半集中式空调——既有中央空调又有处理空气的末端装置的空调系统。这种系统比较复杂，可以达到较高的调节精度。适用于宾馆、酒店、办公楼等有独立调节要求的民用建筑，半集中式空调的输配系统能耗通常低于集中式空调系统。常见的半集中式空调系统有风机盘管系统和诱导式空调系统。图8-4中既有空气局部处理A，又有集中空气处理B共同作用，此系统就属于半集中式。局部式空调——每个房间都有各自的设备处理空气的空调。空调器可直接装在房间里或装在邻近房间里，就地处理空气。适用于面积小、房间分散、热湿负荷相差大的场合，如办公室、机房、家庭等。其设备可以是单台独立式空调机组，也可以是由管道集中给冷热水的风机盘管式空调器组成的系统，各房间按需要调节本室的温度。图8-4中如果没有集中空气处理B，只有局部空气处理A，则该系统属于局部式。3．按照承担负荷介质分类全空气系统——仅通过风管向空调区域输送冷热空气，如图8-5 (a)所示。全空气系统的风管类型有：单区风管、多区风管、单管或双管、末端再热风管、定空气流量、变空气流量系统以及混合系统。在典型的全空气系统中，新风和回风混合后通过制冷剂盘管处理后再送人室内，对房间进行采暖或制冷。图8-4中如果只有集中处理B进行空气调节，就属于全空气系统。全水系统——房间负荷由集中供应的冷、热水负担。中央机组制取的冷冻水循环输送到空气处理单元中的盘管（也称为末端设备或风机盘管）对室内进行空气调节，如图8-5(b)所示。采暖是通过热水在盘管中的循环流动来实现。当环境只要求制冷或采暖、或采暖和制冷不同时进行时，可以采用两管制系统。采暖所需的热水是由电加热器或锅炉制取，利用对流换热器、脚踢板热辐射器、翅片管辐射器、标准风机盘管等进行散热。图8-4中如果只有冷媒水进行局部空气处理A，就属于全水系统。空气一水系统——空调房间的负荷由集中处理的空气负担一部分，其他负荷由水作为介质进入空调房间，对空气进行再处理，如图8-5(c)所示。属于空气一水系统的有末端再热系统、新风十风机盘管系统、带盘管的诱导系统。图8-4中，如果既有B处理过的空气承担部分负荷，又有A处理过的冷冻水承担部分负荷，此时为空气一水系统。直接蒸发式机组系统——又称冷剂式空调系统，空调房间的负荷由制冷剂直接负担，制冷系统蒸发器（或冷凝器）直接从空调房间吸收（或放出）热量，如图8-5 (d)所示。其机组组成为：空气处理设备（空气冷却器、空气加热器、加湿器、过滤器等）通风机和制冷设备（制冷压缩机、节流机构等）。图8-4中只有冷媒局部换热A作用，而且冷媒为液态制冷剂时，就属于直接蒸发式系统。
说到暖通空调，相信很多人跟小编一样都不是特别的了解。暖通空调顾名思义就是将采暖、通风和空气调节这三者合为一的空调器。暖通空调也被人们称作HVAC，这是由采暖、通风、空气通风这三个词的英文缩写组合而成。今天小编就来为大家介绍一下暖通空调系统设计原理及特点，希望可以为大家提供一定的帮助，也为有需要的人提供更多的了解。 一、原理 暖通空调是分户的中央空调，中央空调它最大特点,是能够创造一种舒适的室内环境。而家居一般的分体的空调，它只能解决冷暖问题，而解决不了空气处理过程。有了暖通空调就不一样了。其空气处理过程有以下步骤：首先是空气进来以后，除了引进新风以外，可以把空气进行冷却处理，然后就进行过滤处理，过滤处理以后，增加了几大特点：第一就增加电子除尘器，它主要可以捕捉非常小的颗粒的灰尘，一般来讲它可以捕捉一个微米的灰尘，而这个灰尘的范围内大部分都是细菌、病毒、烟尘，或者是异味这样就都可以过滤掉;另外就是会增加一种加湿设备，这个加湿器可以创造我们房间的加湿达到40%左右的相对湿度，这样人会感到很舒适。 二、特点 在现代化暖通空调系统中，变频技术的应用具有较强的必然性。通过变频技术，既可弥补空调系统的工艺问题，也可减少能源消耗，降低运行成本。一般情况下，空调系统仅按照事先设计的额定功率运行，在负荷较低的情况下，如果设备仍以额定功率实行全负荷运行，那么必然产生能源浪费。通过在暖通空调系统中应用变频技术，就可实现空调设备的输出功率随着负荷的变化情况而有所调节，发挥节能减排效果。结合空调的实际负荷状况，适当改变风流量或者水流量，实现节能目标。 一方面，变风量系统，利用空调系统的末端装置实现室内负荷的补偿机制，优化调整送风量，以保持合适的室内温度;与定风量系统相比较，变风量系统可节能约5O%;另一方面，变水量系统，主要通过控制数量来调节温度，比定流量系统更加省电。随着我国工业变频器的推广与使用，通过优化调节风量、水量及主机等，可实现与空调负荷的匹配运行，发挥良好的节能效益。 很多人可能人会问小编，暖通空调系统在生活中常见吗?相信很多人都没有注意，现在很多单位和公共场所都已经开始应用暖通空调系统。暖通空调技术可以选择热源系统的优化，也采用了节能技术。所以自从暖通空调系统面世以来，便受到了广大消费者的喜爱和追捧。小编今天为大家介绍的暖通空调系统设计原理和特点就到这里了，希望可以为大家带来帮助。
网站上去找吧，我也是这个专业的，以前收集了一些，但是都删了
擦，都到这来找了，一看就522的某个人
3万块全包硕士论文
这个空调的质量还是比较好的，而且相对来说的话使用的效果比较好，整个的话制冷也是非常舒服的。
要是有文献资料 我可以帮你翻译 我现在不知道你要什么样的啊？？
testing of an air-cycle refrigeration system for road transportAbstractThe environmental attractions of air-cycle refrigeration are 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 This unique plant operated satisfactorily, delivering sustainable cooling for refrigerated trailers using a completely natural and safe working 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 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 Keywords: 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 IntroductionThe current legislative pressure on conventional refrigerants is well 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 Consequently, as soon as alternative systems, such as non-HFC refrigerants or air-cycle systems are considered, the issue of increased energy consumption arises Concerns 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 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 NormalAir Garrett developed and commercialised an air-cycle air conditioning pack that was fitted to high speed trains in Germany in the As part of an European funded programme, a range of applications for air-cycle refrigeration were investigated and several demonstrator plants were 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 Thermo 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 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 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 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 In practical terms, this meant using the chassis and panels for an existing SL200 unit along with the standard diesel engine and circulation 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 For anything other than the smallest flow rates, the most efficient machines available for the necessary compression and expansion processes are Considerations for the selection of turbomachinery for air-cycle refrigeration systems have been presented and discussed by Spence et [3] a typical configuration of an air-cycle system, which is sometimes called the ‘boot-strap’ 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 Instead, the work recovered by the turbine during expansion is utilised in the secondary The two-stage compression also permits intercooling, which enhances the overall efficiency of the compression An ‘open system’ where the cold air is ejected directly into the cold space, removing the need for a heat exchanger in the cold 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 The compression and expansion processes were modelled using appropriate values of isentropic efficiency, as defined in EThe heat exchange processes were modelled using values of heat exchanger effectiveness as defined in The model also made allowance for heat exchanger pressure 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 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 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 However,illustrates, there is a broad range of pressure ratio and duty over which the system can be operated with only moderate variation of COPThe class of turbomachinery suitable for the demonstrator plant required speeds of around 50 000 rev/ To simplify the mechanical arrangement and avoid the need for a high-speed electric motor, the two-stage compression system shown was 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 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 The two-stage system in incorporated an intercooler between the two compression By dispensing with the intercooler and its associated duct work a larger aftercooler could be accommodated with improved effectiveness and reduced pressure Analysis suggested that the improved performance from a larger aftercooler could compensate for the loss of the shows the impact of the recuperator effectiveness on the COP of the plant, which is clearly more significant than that of the other heat As well as boosting cycle efficiency, increased recuperator effectiveness also moves the peak COP to a lower overall system pressure The impact of pressure loss in the recuperator is the same as for the intercooler and aftercooler shown The model did not distinguish between pressure losses in different locations; it was only the sum of the pressure losses that was 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 The specific cooling capacity of the air-cycle increases with system pressure Consequently, if a higher system pressure ratio was used the required cooling duty could be achieved with a smaller flow rate of shows the mass flow rate of air required to deliver 7,5 kW of cooling power for varying system pressure Since 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 and were based on efficiencies of 81 and 85% for compression and expansion, While such efficiencies are attainable with optimised designs, they would not be realised using compromised turbocharger For the design of the demonstrator plant efficiencies of 78 and 80% were assumed to be realistically attainable for compression and Lower turbomachinery efficiencies corresponded to higher cycle pressure ratios and flow rates in order to achieve the target cooling The cycle design point was also compromised to help heat exchanger The pressure losses in duct work and heat exchangers increased in proportion with the square of flow Selecting a higher cycle pressure ratio corresponded to a lower mass flow rate and also increased density at inlet to the aftercooler heat 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 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 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, The system pressure ratio at the design point was 2,14 and the corresponding mass flow rate of air was 0,278 kg/By 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 Also, operating at part-load corresponded to lower flow velocities and anticipated improvements in heat exchanger 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 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 From the thermodynamic model, the pressure ratio for the primary compressor was 1, The centrifugal compressor required a shaft speed of around 55 000 rev/ Other alternatives were evaluated for primary compression with the aim of obtaining a suitable device that operated at a lower Other commercially available devices such as Roots blowers and rotary piston blowers were all excluded on the basis of poor A one-off gearbox was designed and manufactured as part of the project to step-up the engine shaft speed to around 55 000 rev/ 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 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 Consequently, both the secondary compressor and the turbine stage were specially chosen and developed to deliver suitable Most turbochargers use plain oil fed journal bearings, which are low-cost, reliable and provide effective damping of shaft However, plain bearings dissipate a substantial amount of shaft power through viscous losses in the oil 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 Also, the clearances in plain bearings require larger blade tip clearances for both the compressor and the turbine with a consequential efficiency Given the pressurised inlet to the secondary compressor, the limited thrust capacity of the plain bearing arrangement was also a A CAU utilising high-speed ball bearings, or air bearings, was identified as a preferable arrangement to plain Benefits would include greatly reduced bearing power losses, reduced turbomachinery tip clearance losses and increased thrust load However, adequate resources were not available to design a special one-off high speed ball bearing Consequently, a standard turbocharger plain bearing system was The secondary compressor stage was a standard turbocharger compressor selected for a pressure ratio of 1, Secondary compressor and turbine selection were linked because of the requirement to balance power and match the Since most commercial turbines are sized for high temperature (and consequently low density) air at inlet, a special turbine stage was developed for the Cost considerations precluded the manufacture of a custom turbine rotor, so a commercially available rotor was 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 The exhaust diffuser exited into a specially designed exhaust The performance of the turbine stage was measured before the unit was incorporated into the complete demonstrator The peak efficiency of the turbine was established at 81% 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 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 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 Several different tube and fin heat exchangers were tested and used to validate a computational Once validated, the model was used to assess a wide range of possible heat exchanger configurations that could fit within the Thermo King SL200 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 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 The final selection comprised two separate aftercooler units, while the single recuperator was a large, triple pass Based on laboratory tests and the heat exchanger model, the anticipated effectiveness of both the recuperator and aftercooler units was 80% InstrumentationA range of conventional pressure and temperature instrumentation was installed on the air-cycle demonstrator Air temperature and pressure was logged at inlet and outlet from each heat exchanger, compressor and the 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 No air flow measurement was included on the demonstrator 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 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 The calorimeter was ideally suited for accurately measuring the refrigeration capacity of the air-cycle demonstrator 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 The 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 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 The demonstrator plant operated satisfactorily, allowing sufficient measurements to be obtained at each of the three operating The recorded performance is summarised In total, the unit operated for approximately 3 h during the course of the various 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 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 A unique air-cycle refrigeration system had been developed within the same physical envelope as the existing Thermo King SL200 refrigeration unit, w
你是学建筑环境也设备工程的不