滑雪场造雪系统中炮式造雪机的整体设计【三维图PPT】【机械毕业设计论文说明书CAD图纸】.zip
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滑雪场造雪系统中炮式造雪机的整体设计【三维图PPT】【机械毕业设计论文说明书CAD图纸】.zip,三维图PPT,机械毕业设计论文说明书CAD图纸,毕业设计论文,设计毕业设计
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课题简介人工造雪是相当昂贵的。然而，当对比成本与收益时，就会发现人工造雪给许多旅游点带来了丰厚的利润。人工造雪能够显著地增加顾客的滑雪体验。机器造雪能够在整个滑雪季中提供足够的雪量。在滑雪季初期，滑雪场几乎完全依赖人工造雪。机器造雪也易于在较长时间内保证雪的质量不变，并且比自然雪更适于抵抗升华及来自光源的热量的影响，滑雪者的装备也能够重复使用。山上的雪一旦失去其晶体结构就会变成球形的，也就不再能被塑造成形。将人造雪覆盖在已经失去晶体结构的雪体的表面可以重新激发其晶体结构，从而使雪恢复生命力。而机器造雪的设计被造雪机的基本原理并没有任何变化。将水注入一个专用喷嘴或喷枪，在那里接触到高压空气，高压空气将水流分割成微小的粒子并喷入寒冷的外部空气中，在落到地面以前这些小水滴凝固成冰晶，也就是人们看到的雪花。造雪机是一种可以迅速把大量液态水转化成为高压雾化冰晶的电气装置，主要用于人工造雪，布置人工滑雪场地、消防等方面。多数配有方便移动的履带式车轮，上载有直径较宽（直径约半米）粗短的喷雪炮筒。数配有方便移动的履带式车轮，上载有直径较宽（直径约半米）粗短的喷雪炮筒。造雪机原理是在-15的蒸发器上结成冰，通过冷却的空气输送到滑雪道方式的不受大气温度影响的崭新造雪系统。人工造雪机不受气候的影响，只要能保持一定的水量就可以造雪。具体任务、内容及要求（包括设计计算、实验分析、绘图质量各类图纸张数、外文翻译、参考文献及撰写外文摘要等要求）设计的主要内容及要求：1、 查阅文献、熟悉课题、撰写开题报告；2、 造雪机的工艺分析；3、 造雪机的运动特性分析；4、 造雪机及其控制电路的设计；5、 俯仰运动的实现与机构设计；6、 确定造雪机设计；7、 运动及动力参数计算；8、 根据课题要求设计整体尺寸及零部件强度计算；9、 用CAD软件绘造雪机的设计总装配图；10、重要零件的2D图。编写一本不少于10000字的设计说明书，撰写1500字以上的文献综述，独立翻译一篇2000字以上、与课题相关的外文参考文献，所有设计内容均由计算机及相应软件形成电子文档并打印，绘图量不少于折成A0号图纸4张，参考文献不少于10篇。时间进度安排实习调研、查阅资料 第 1-2 周上机运算（绘图）第 10-13 周方案确定 第 3-4 周撰写说明书（论文）第 14 周设计计算（实验） 第 5-9 周上交论文（设计）第 15 周教研室主任签章： 毕业论文（设计）领导小组组长签章： 教务处制表外文文献Ejector refrigeration systems have long been an attractive research subject for a lot of researchers due to being heat- driven systems and having simple designs. Two importan drawbacks of these systems are primarily being refrigeration systems with low COP and using mechanically driven pumps. If the two disadvantages could be eliminated, especially the first one, the systems could find a wide area of application in air-conditioning and refrigeration industries. To produce a cooling effect, such heat-driven refrigeration systems could use low-grade energy sources, which are widely available and have low-cost such as solar and geothermal energy and waste heat. If the liquid pumps used in these systems could be ther- mally driven or an ejector refrigeration system without the pump could be developed, then the systems will be indepen- dent of electrical energy. In addition, the problems dealing with the operation of the pump and some additional devices coming with the pump will be eliminated. As expected, recent researches on the ejector refrigeration systems have focused on the improvement of energy conversion performance of these systems. Studies carried out to increase their efficien- cies are generally on a better design of ejector, the selection of a proper refrigerant, the optimization of operating condi- tions and the addition of various (secondary) devices such as a pre-cooler and regenerator to the refrigeration cycle. Scien- tists have densely studied on these research subjects for sev- eral decades and have achieved a significant improvement in the coefficient of performance of the systems using one or more methodsChang and Chen (2000) used a petal nozzle to enhance the performance of a steam-jet refrigeration system. According to their experimental results, when the system is operated at larger area ratios, the performance of system with a petal noz- zle is better than that with a conical nozzle.It is an attractive subject to develop an ejector refrigeration system without a liquid pump or with only thermal-driven. If this is achieved, such systems will not contain any moving part. For this purpose, Huang et al. (2006) have tried develop- ing an ejector cooling system with thermal pumping effect. The authors designed a cooling system with a multi-function generator which serves as both a liquid pump and a generator, thus eliminating a pump from system. These researchers tested the system and concluded that the design of such a sys- tem is feasible and it is possible to produce continuousl a cooling effect by regularly switching the two multi-function generators of the system.Eames (2002) introduced a new method for designing ejec- tors to be used in ejector refrigeration systems. It is assumed in the method that the momentum of flow changes at a con- stant rate within the diffuser passage of a supersonic ejector. The theoretical method produces a diffuser geometry that removes the thermodynamic shock process within the dif- fuser at the design-point operating conditions.Eames et al. (1999) examined the effects of ejector geometry on the performance of steam jet-pump refrigerators, using two primary nozzles and three diffusers with mixing chamber. It is seen from their experimental results that the entrainment ra- tio increases almost linearly with the ejector ratio area, if the primary pressure ratio ( pg/pc) and the ratio of the primary noz- zle exit area to throat area (Ane/Ant) are held constant.Chunnanond and Aphornratana (2004) examined the ef- fects of the nozzle geometry and position on the performance of a steam ejector refrigerator with a conical mixing chamber. Based on their tests, they expressed that decreasing the gener- ator pressure, using a nozzle with smaller throat area (hence higher area ratio) and retracing the nozzle out of the mixing chamber can increase the COP and cooling capacity of the re- frigerator, provided that the critical condenser pressure is decreased.No matter what the working refrigerant, the type of ejector and the secondary devices are used in ejector refrigeration systems; it is required to optimize the operating conditions of the systems according to geometrical and flow parameters, to obtain maximum performance from them. Optimum oper- ating conditions of these systems mainly change depending upon the ejector area ratio (Nahdi et al., 1993; Yapc and Ersoy, 2005; Sun and Eames, 1996). In other words, when oper- ating temperatures or pressures were determined based on practical considerations such as heat source, refrigerated me- dium and heat rejected medium temperatures, an ejector should be designed so that its area ratio and nozzle position are optimum for the selected operating conditions. The aim of this study is to determine the optimized operating condi- tions of an ejector refrigeration system using R-123 in a wide In the first part of the study, primary vapor flow rates through these nozzles were measured at various generator temperatures and then the optimum nozzle position for each area ratio was found separately by taking the minimum pressure at the suction chamber of the ejector as the criteria. In the second part of the study, it is described how the opti- mum operating condition is determined for a given area ratio To do this, as an example, the experimental results including the effects of operating temperatures or pressures on the COP of the system were also presented for the smallest area ratio6.5 in the present study. That is, three different experiments were conducted to find the optimum operating point for each area ratio. In consequence, the optimum generator tempera- ture for the area ratio was found at the evaporator temperature of 10 oC and the condenser pressure of 125 kPa.In the last part of the study, the optimum generator tem- peratures for the remaining five area ratios were determined separately by repeating experiments at the same evaporator and condenser conditions. It was seen that there exists only an optimum generator temperature for each area ratio. All ex- perimental curves showing variations in COP against the gen- erator temperature were totally presented on a graph and the optimum performance curve was generated as a function of the temperature. Moreover, the optimum experimental re- sults were compared with the optimized analytical results obtained for the same refrigerant by using the ejector flow model given in the literature (Yapc and Ersoy, 2005).The optimum COP almost linearly increases with the gen- erator temperature at the fixed evaporator and condenser conditions. The experimental curve of the optimum COP coin- cides with that of theoretical COP while the efficiencies of the primary nozzle and diffuser are 0.90.Fig. 1 schematically depicts the experimental setup. The main elements of the test facility are a vapor generator, an ejector with movable nozzle, a condenser, a receiver tank, an expan- sion valve, an evaporator, a sub-cooler and a sliding vane pump; the refrigerant fluid circulates through the devices. The secondary elements of the refrigeration system are a hot water boiler, a circulating pump, pre-heater, through them which water flows, the measuring and control devices. The experimental setup and procedure, and the operation principle of the ejector refrigeration system are described in detail in the literature (Yapc, 2008).The output signals from the measurement devices were transferred to a PC through a data acquisition board and all readings were monitored and also recorded by the computer. All devices used for the measurement were calibrated to- gether with the data acquisition system in their measurement range. The pressure of the vapor generator was measured with the accuracy of 土 10 kPa, whereas other pressures in the system were measured with the accuracies of 土 1 kPa. The water flow rate of the evaporator was measured with an accuracy of 土 0.03 L/min. The temperature of the vapor in the generator was measured with the accuracy of 土 0.5 oC. Other temperatures were measured within 土 0.2 oC accuracy. Based on the inaccuracies in measuring the temperature, vol- ume and time period, the uncertainty in the mass flow rate of the primary vapor is determined to be within 土 4.8 %. The un- certainties in COP at the optimum operating conditions arewithin 土5.7%.The ejector model used in this study is shown in Fig. 2. In this model, inlet to the mixing chamber is with rounded- entry, the mixing chamber is of constant-area and the diffuser is conical. The ejectors were designed based on the constant- area ejector flow model given in the literature (Yapc and Ersoy, 2005; Ersoy, 1999). The configurations and dimensions of the six ejectors used in the experiments are presented in Table 1. These six configurations are obtained by using differ- ent three supersonic nozzles and two mixing chambers with diffusers.The coefficient of performance (COP) is the most important parameter for evaluating the performance of the ejector re- frigeration system. It indicates the cooling capacity relative to the energy input into the system. The energy input to the system in the refrigerant pump is too low compared with the energy input in the vapor generator. Therefore, neglecting the power of the liquid pump, the performance coefficient of the ejector refrigeration system is calculated from the follow- ing expression.COP(Qe/ Qg)/4(1)The cooling capacity was determined from Q_ e m_ cwCpTin Texe(2)where m_ cw is the mass flow rate of chilling water.The heat input to the vapor generator was calculated fromQ_ g m_ phex hing(3)To determine the flow rates of primary vapor at various generator temperatures for each nozzle, the experiments were carried out and the mass flow rates of the primary nozzles were separately found as a function of the vapor generator temperature. Variations in the mass flow rate of primary vapor with the generator temperature are shown in Fig. 3 for three nozzles used in the experiments. Using the method of least squares, the m_ p Tg curves were separately fitted to the experimental data for each nozzle. The equations of the curves were found and used in deter- mining the primary mass flow rates at the various generator temperatures.The relative nozzle position Ln/dm is in the range 0.5 Ln/ dm 2 according to the experimental results for refrigerant R-11 given in Nahdi et al. (1993). In the preliminary experi- ments carried out, similar results were found for ejector with the cylindrical mixing chamber at Pc 125 kPa, Pg 752.4 kPa (98 oC) and Ar 9.97, and using refrigerant R-123. Therefore, the nozzle position was adjusted toLn 5 mm (upstream the entry of cylindrical mixing cham- ber) in each experiment and the optimum generator tempera- tures were determined first separately by keeping evaporator and condenser conditions constant. After that, the variations of the suction chamber pressure with the nozzle position were measured at the optimum generator temperatures. Finally, it was controlled whether the adjusted nozzle distance is in the optimum range. The results for these experiments are shown in Fig. 4. In determination of the optimum nozzle posi- tion, as in the literature (Hamner, 1978), the minimum pres- sure in the suction chamber was taken as criteria.It is clearly seen from the foregoing figure that the optimum range of the adjusted nozzle position for all nozzles is within 10 to 5 mm.When the condenser and evaporator temperatures/ pressures are known, it is possible to find the optimum gener ator temperature for a given ejector area ratio. In this study, he evaporator temperature is selected as 10 oC, which is ppropriate for air-conditioning. To ensure the operation of he system at the choked operating conditions even at the smallest ejector area ratio, the condenser pressure is taken as 125 kPa (saturation temperature nearly 34 oC). As an exam- ple, the result of experiment carried out to determine the optimum generator temperature is shown in Fig. 5 for the ejector area ratio Ar 6.56. The maximum COP for this area ratio was obtained at Tg 83 oC and hence the temperature is optimum generator temperature.In order to determine the critical condenser pressure and hence to control whether the ejector operates at the condi- tions with choking in the mixing chamber, the variation of COP with condenser pressure is investigated at Tg 83 oC and Te 10 oC. According to the curve shown in Fig. 6, the crit- ical condenser pressure is about 126 kPa and thus operating point of the ejector is very near the critical operating condi- tion, that is, the system operates in maximum cooling capac- ity at the given area ratio.Fig. 7 shows how COP changed with the evaporator tem- perature at the same area ratio. Actually, this last experiment for the area ratio was done to verify the COP value at the opti- mum operating condition specified above. Thus the experi- mentation at an operating point becomes repeated three times. Referring to Fig. 7, we see that COP value at Te 10 oC is about 29% again. On the other hand, COP increasing with the evaporator temperature is an expected result under such operating conditions.The experimental procedure described above was repeated for the remaining five ejectors and the COPTg curves were de- termined experimentally for the six area ratios and are all shown in Fig. 8 by different marks. From these curves, it is clearly seen that an optimum generator temperature for every area ratio exists at the fixed evaporator and condenser tem- peratures. As shown in the same figure, the optimum COP curve is the line which is a tangent to the peaks of COPTg curves. Optimum COP increases nearly linearly with the opti- mum generator temperature. The value of COPopt reaches to around 41% at the area ratio 11.45 from 29% at the area ratio6.56. Based on the results shown in Fig. 7, it can be expressed that a higher COPopt can be obtained by increasing the evapo- rator temperature. Other important result which should be expressed here is that the slope of COPTg for a given area ratio is too high at lower temperatures than the optimum generator temperature. Therefore, if the difference between the optimum and operating temperatures of the generator at an area ratio is higher than a given value, which is around 10 oC here, the system does not achieve its function.Comparison of the optimum experimental results with optimum theoretical results is presented in Fig. 9 for the same operating conditions. The optimum theoretical results were determined for three different efficiencies of ejector elements, using the methods provided in literature (Yapc and Ersoy, 2005). To obtain these theoretical results, efficien- cies of the supersonic nozzle and diffuser were kept constant at the values shown in Fig. 9. When the efficiencies were 90%, the experimental data agree very well with the theoretical data.Fig. 10 shows variations of the optimum experimental and theoretical area ratios with the generator temperature. According to both experimental and theoretical results, the optimum area ratio increases almost linearly with the temper- ature. In other words, a lower generator temperature means a smaller ejector area ratio. To obtain a higher COP at a fixed generator temperature, the ejector area ratio should be in- creased together with efficiencies of ejector elements. More- over, the theoretical results for efficiencies of 90% agree with the experimental results. The theoretical area ratios are just slightly higher than the experimental ratios.中文文献喷射器制冷系统作为一种热驱动系统，由于其设计简单，长期以来一直是许多研究者关注的研究课题。这些系统有两个主要缺点制冷效率低和使用机械驱动泵。如果能消除这两个缺点，特别是第一个缺点，该系统将在空调和制冷行业得到广泛的应用。生产这种热驱动制冷系统具有制冷效果，可以使用低品位的能源，如太阳能、地热能和余热,这些能源广泛存在，而且成本低廉。如果这些系统中使用的液体泵能够独立驱动，或者能够开发出没有泵的喷射式制冷系统，那么这些系统可以不使用电能。此外，泵和一些与泵相关的额外设备将被消除。近年来对喷射器制冷系统的研究主要集中在提高系统的能量转换性能上。为提高它们的效率而进行的研究一般包括更好地设计喷射器、选择适当的制冷剂、优化操作设备以及在制冷循环中增加各种设备，如预冷器和回热器。科学家们对这些研究课题进行了数十年的密集研究，并使用一种或多种方法在系统性能系数方面取得了显著的提高。张和陈使用花瓣喷嘴来提高蒸汽喷射制冷系统的性能。实验结果表明，当系统在较大的面积比下运行时，花瓣型喷嘴系统的性能优于锥型喷嘴系统。开发一种无液泵或只有热驱动的喷射器制冷系统是一个很有吸引力的课题。如果实现了这一点，这些系统将不包含任何移动部件。为此，黄等人尝试开发了一种具有热泵效应的喷射器冷却系统。设计了一种集液体泵和发电机于一体的多功能发生器冷却系统，消除了系统中泵的功能。研究人员对该系统进行了测试，认为该系统的设计是可行的，可以连续生产通过定期切换系统的两个多功能发生器，达到冷却效果。埃姆斯介绍了一种设计用于喷射器制冷系统的新方法。该方法假定在超音速喷射器的扩散段内，流动动量以恒定的速率变化。该理论方法产生了一个扩散器几何形状，消除了设计工况下dif-熔断器内的热力学激波过程。埃姆斯等人(1999)使用两个主喷嘴和三个带有混合室的扩散器，研究了喷射器几何形状对蒸汽喷射泵制冷机性能的影响。实验结果表明，一次压力比和主要出口面积与喉道面积的比值保持不变。Chunnanond和Aphornratana(2004)研究了带有锥形混合室的蒸汽喷射制冷机喷嘴几何形状和位置对性能的影响。基于他们的测试,他们表示,减少系统压力,使用较小的喷嘴喉部面积(因此更高的面积比)和使用喷嘴混合室可以提高再保险更好的安全性和冷却能力,提供关键的冷凝器压力却降低了。在喷射器制冷系统中，无论使用何种工质制冷剂，都要考虑喷射器的类型和二次装置;为了获得最大的性能，需要根据几何参数和流量参数对系统的运行条件进行优化。这些系统的最佳操作条件主要随喷射器面积比的变化而变化,换句话说，当根据实际考虑，如热源、制冷剂和废热介质温度来确定操作温度或压力时，应该设计一个喷射器，使其面积比和喷嘴位置在选定的操作条件下是最佳的。本研究的目的是在大范围内确定R-123喷射器制冷系统的优化运行参数喷射器面积比的范围。为了达到这个目的，三个不同的主喷嘴和两个不同的混合室被制造成六个面积比。该系统通过将这些元件安装在喷射器上进行测试，喷射器内的喷嘴可以轴向移动。在研究的第一部分中，首先在不同的发生器温度下测量了这些喷嘴的一次蒸汽流量，然后以喷射器吸力室的最小压力为准则，分别找出各面积比下的最佳喷嘴位置。在研究的第二部分，介绍了在给定的面积比下，确定了最小二乘操作条件。为此，以最小面积比为例，给出了包括操作温度或压力对系统的影响在内的实验结果本研究中。即通过三个不同的实验来寻找各面积比的最佳工作点。结果表明，在蒸发器温度为10，冷凝器压力为125 kPa的情况下，得到了最佳的比表面积温度。最后，在相同的蒸发器和冷凝器条件下，通过反复试验，分别确定了其余五种面积比的最佳发电机性能。结果表明，各面积比只存在一个最优的发电机温度。所有表明随发生器温度变化的围周曲线均在图上完整地表示出来，并与实验结果进行了比较, 以温度为函数，得到了最佳性能曲线。此外,优化实验,结果比较与优化分析结果为同一制冷剂通过使用文献中给出的喷射流模型。在固定的蒸发器和冷凝器条件下，最优曲线几乎随发生器温度线性增加。在一次喷管和扩压器效率为0.90时，最佳-的实验曲线与理论曲线一致。图1为实验装置示意图。该试验装置的主要部件为蒸汽发生器、带活动喷嘴的喷射器、冷凝器、接收罐、膨胀阀、蒸发器、副冷却器和滑片泵;制冷剂液体在装置中循环。制冷系统的二次元件是热水锅炉、循环泵、预热器、流经它们的水流、测控装置。实验设置和过程和喷射制冷系统的工作原理被详细地描述在文献中。读数被监测，并由计算机记录。所有用于测量的设备都与测量范围内的数据采集系统校准。蒸汽发生器的压力测量的准确性与土10 kPa,而其他系统中压力测量的精度为土1 kPa。蒸发器的水流速测量的流量为0.03 L / min。的蒸汽发生器的温度测量的温度为 0.5摄氏度。在测量温
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