您好，我是科园月刊杂志社责任编辑小王，有文章需要发表的话，可以随时Q我。QQ见用户名。《科园月刊》省级综合性学术期刊（国家新闻出版总署检索，万方龙源全文收录，核心遴选）长期征稿 我们是杂志社直接征稿，出刊及时，随时找我。
我以前有发表过文章，是通过一个叫公务员之家的网站发表的，而且是代写代发，个人觉得挺讲信用，而且速度也不错，可以去了解一下。
我评的是建筑工程师，第一作者两篇论文就行，对论文质量要求不高，大约800－1000字，省级刊物发表，发表也不算难，交钱就行，发表一篇得500元吧，我当时发表了两篇，每篇400元
份建筑环境与设备工程(暖通)相关的英肯定知道的
份建筑环境与设备工程(暖通)相关的英肯定知道的
发过去了 分给我哦～ 发了3篇。。。。pdf格式
lalallalalalallalalalal
暖通空调 就是一个杂志啊或者其他自然学科的杂志
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，
CN就是国内统一刊号，ISSN是国际统一刊号。所谓CN类刊物是指在我国境内注册、国内公开发行的刊物。该类刊物的刊号均标注有CN字母。另外判断期刊是否正规就看CN刊号，即国家新闻出版总署正式批准出版的期刊都应具备CN刊号。
EnergyEnergy and BuildingsRenewable and Sustainable Energy ReviewsApplied EnergyScience & Technology for the Built EnvironmentComputers Environment and Urban Systems暖通空调建筑热能通风空调建筑科学建筑技术开发等等。。
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，
专业的有暖空空调杂志、安装杂志这样的，还有那些论文集的基本不审。
暖通制冷空调新暖通区域供热煤气与热力暖通制冷空调像《热泵技术》类反挺我觉《暖通空调》《新暖通》比较名
暖通空调就很好啊，他的增刊很好，制冷与空调也行，得看自己侧重的行业
这个上面你自己看下能不能用得上哦我不懂这个的
If the entering air condition ischanged to Point D at the same wet-bulb temperature but at a higherdry-bulb temperature, the total heat transfer (Vector DB) remainsthe same, but the sensible and latent components change dramati- DE represents sensible cooling of air, while EB representslatent heating as water gives up heat and mass to the Thus, forthe same water-cooling load, the ratio of latent to sensible heattransfer can vary The ratio of latent to sensible heat is important in analyzing waterusage of a cooling Mass transfer (evaporation) occurs only inthe latent portion of heat transfer and is proportional to the changein specific Because the entering air dry-bulb temperatureor relative humidity affects the latent to sensible heat transfer ratio,it also affects the rate of In Figure 2, the rate of evapo-ration in Case AB (WB − WA) is less than in Case DB (WB − WD)because the latent heat transfer (mass transfer) represents a smallerportion of the The evaporation rate at typical design conditions is approximately1% of the water flow rate for each 7 K of water temperature range;however, the average evaporation rate over the operating season isless than the design rate because the sensible component of total heattransfer increases as entering air temperature In addition to water loss from evaporation, losses also occurbecause of liquid carryover into the discharge airstream and blow-down to maintain acceptable water 如果是进入空调改为D点在同一湿球温度，但在较高干球温度，总传热（矢量数据库）仍然相同，但明智的和潜在的组成部分的剧作家，卡利。明智署署长代表空气冷却，而电子束代表潜热，水热，放弃了对空气质量。因此，同样的水冷却负荷，潜显热比转移有很大的差别。对潜在的比例显热分析是很重要的水使用的冷却塔。传质（蒸发）只发生在潜在的传热部分，是成比例的改变在具体的湿度。由于进入空气干球温度或相对湿度影响潜到显热传递率，它也影响到蒸发率。在图2中，土壤水分蒸发的速度，在个案公司（世行 - WA）的比例是比案例数据库（世行 - 西数较少）因为潜热转移（质）代表一个较小的部分总额。在典型的设计条件下蒸发率约为1％的水每7 K表温度范围内水流量;但是，在工作赛季平均蒸发量低于设计速度，因为总热量合理成分作为进入空气温度降低转移增加。除了水的蒸发损失，损失也可能发生由于结转到液体排放气流吹到保持可接受的水质。
一种很常见的测压设备称为波登管式压力计。这种压力计的基本构造是一条圈状椭圆形管，其一端固定，另一端可自由移动。要测量的压力被传导到管的内侧。当管内外的压力出现差异时，自由端会移动。管的移动传导到压力计的指针。由于管外侧的压力就是大气压力，压力计的读数就是高于或低于大气压力的值。但一般情况下，压力计测量的压力定义为高于大气压力的压力。
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
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