滑雪場(chǎng)造雪系統(tǒng)中炮式造雪機(jī)的整體設(shè)計(jì)含SW三維及8張CAD圖
滑雪場(chǎng)造雪系統(tǒng)中炮式造雪機(jī)的整體設(shè)計(jì)含SW三維及8張CAD圖,滑雪場(chǎng),系統(tǒng),中炮式造雪機(jī),整體,總體,設(shè)計(jì),sw,三維,cad
課題簡(jiǎn)介
人工造雪是相當(dāng)昂貴的。然而，當(dāng)對(duì)比成本與收益時(shí)，就會(huì)發(fā)現(xiàn)人工造雪給許多旅游點(diǎn)帶來(lái)了豐厚的利潤(rùn)。人工造雪能夠顯著地增加顧客的滑雪體驗(yàn)。機(jī)器造雪能夠在整個(gè)滑雪季中提供足夠的雪量。在滑雪季初期，滑雪場(chǎng)幾乎完全依賴人工造雪。機(jī)器造雪也易于在較長(zhǎng)時(shí)間內(nèi)保證雪的質(zhì)量不變，并且比自然雪更適于抵抗升華及來(lái)自光源的熱量的影響，滑雪者的裝備也能夠重復(fù)使用。山上的雪一旦失去其晶體結(jié)構(gòu)就會(huì)變成球形的，也就不再能被塑造成形。將人造雪覆蓋在已經(jīng)失去晶體結(jié)構(gòu)的雪體的表面可以重新激發(fā)其晶體結(jié)構(gòu)，從而使雪恢復(fù)生命力。而機(jī)器造雪的設(shè)計(jì)被造雪機(jī)的基本原理并沒(méi)有任何變化。將水注入一個(gè)專用噴嘴或噴槍，在那里接觸到高壓空氣，高壓空氣將水流分割成微小的粒子并噴入寒冷的外部空氣中，在落到地面以前這些小水滴凝固成冰晶，也就是人們看到的雪花。造雪機(jī)是一種可以迅速把大量液態(tài)水轉(zhuǎn)化成為高壓霧化冰晶的電氣裝置，主要用于人工造雪，布置人工滑雪場(chǎng)地、消防等方面。多數(shù)配有方便移動(dòng)的履帶式車輪，上載有直徑較寬（直徑約半米）粗短的噴雪炮筒。數(shù)配有方便移動(dòng)的履帶式車輪，上載有直徑較寬（直徑約半米）粗短的噴雪炮筒。造雪機(jī)原理是在-15℃的蒸發(fā)器上結(jié)成冰，通過(guò)冷卻的空氣輸送到滑雪道方式的不受大氣溫度影響的嶄新造雪系統(tǒng)。人工造雪機(jī)不受氣候的影響，只要能保持一定的水量就可以造雪。
具體任務(wù)、內(nèi)容及要求（包括設(shè)計(jì)計(jì)算、實(shí)驗(yàn)分析、繪圖質(zhì)量各類圖紙張數(shù)、外文翻譯、參考文獻(xiàn)及撰寫(xiě)外文摘要等要求）
設(shè)計(jì)的主要內(nèi)容及要求：
1、 查閱文獻(xiàn)、熟悉課題、撰寫(xiě)開(kāi)題報(bào)告；
2、 造雪機(jī)的工藝分析；
3、 造雪機(jī)的運(yùn)動(dòng)特性分析；
4、 造雪機(jī)及其控制電路的設(shè)計(jì)；
5、 俯仰運(yùn)動(dòng)的實(shí)現(xiàn)與機(jī)構(gòu)設(shè)計(jì)；
6、 確定造雪機(jī)設(shè)計(jì)；
7、 運(yùn)動(dòng)及動(dòng)力參數(shù)計(jì)算；
8、 根據(jù)課題要求設(shè)計(jì)整體尺寸及零部件強(qiáng)度計(jì)算；
9、 用CAD軟件繪造雪機(jī)的設(shè)計(jì)總裝配圖；
10、重要零件的2D圖。
編寫(xiě)一本不少于10000字的設(shè)計(jì)說(shuō)明書(shū)，撰寫(xiě)1500字以上的文獻(xiàn)綜述，獨(dú)立翻譯一篇2000字以上、與課題相關(guān)的外文參考文獻(xiàn)，所有設(shè)計(jì)內(nèi)容均由計(jì)算機(jī)及相應(yīng)軟件形成電子文檔并打印，繪圖量不少于折成A0號(hào)圖紙4張，參考文獻(xiàn)不少于10篇。
時(shí)間進(jìn)度安排
實(shí)習(xí)調(diào)研、查閱資料
第 1-2 周
上機(jī)運(yùn)算（繪圖）
第 10-13 周
方案確定
第 3-4 周
撰寫(xiě)說(shuō)明書(shū)（論文）
第 14 周
設(shè)計(jì)計(jì)算（實(shí)驗(yàn)）
第 5-9 周
上交論文（設(shè)計(jì)）
第 15 周
教研室主任簽章： 畢業(yè)論文（設(shè)計(jì)）領(lǐng)導(dǎo)小組組長(zhǎng)簽章：
教務(wù)處制表
外文文獻(xiàn)
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 methods
Chang 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; Yap?c? 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 ratio
6.5 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 (Yap?c? 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 (Yap?c?, 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 (Yap?c? 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_ cwCpeTin — TexTe(2)
where m_ cw is the mass flow rate of chilling water.
The heat input to the vapor generator was calculated from
Q_ g ? m_ pehex — hinTg (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 to
Ln ? —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 COP–Tg 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 COP–Tg 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 ratio
6.56. 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 COP–Tg 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 (Yap?c? 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.
中文文獻(xiàn)
噴射器制冷系統(tǒng)作為一種熱驅(qū)動(dòng)系統(tǒng)，由于其設(shè)計(jì)簡(jiǎn)單，長(zhǎng)期以來(lái)一直是許多研究者關(guān)注的研究課題。這些系統(tǒng)有兩個(gè)主要缺點(diǎn)制冷效率低和使用機(jī)械驅(qū)動(dòng)泵。如果能消除這兩個(gè)缺點(diǎn)，特別是第一個(gè)缺點(diǎn)，該系統(tǒng)將在空調(diào)和制冷行業(yè)得到廣泛的應(yīng)用。生產(chǎn)這種熱驅(qū)動(dòng)制冷系統(tǒng)具有制冷效果，可以使用低品位的能源，如太陽(yáng)能、地?zé)崮芎陀酂?這些能源廣泛存在，而且成本低廉。如果這些系統(tǒng)中使用的液體泵能夠獨(dú)立驅(qū)動(dòng)，或者能夠開(kāi)發(fā)出沒(méi)有泵的噴射式制冷系統(tǒng)，那么這些系統(tǒng)可以不使用電能。此外，泵和一些與泵相關(guān)的額外設(shè)備將被消除。近年來(lái)對(duì)噴射器制冷系統(tǒng)的研究主要集中在提高系統(tǒng)的能量轉(zhuǎn)換性能上。為提高它們的效率而進(jìn)行的研究一般包括更好地設(shè)計(jì)噴射器、選擇適當(dāng)?shù)闹评鋭?、?yōu)化操作設(shè)備以及在制冷循環(huán)中增加各種設(shè)備，如預(yù)冷器和回?zé)崞鳌？茖W(xué)家們對(duì)這些研究課題進(jìn)行了數(shù)十年的密集研究，并使用一種或多種方法在系統(tǒng)性能系數(shù)方面取得了顯著的提高。
張和陳使用花瓣噴嘴來(lái)提高蒸汽噴射制冷系統(tǒng)的性能。實(shí)驗(yàn)結(jié)果表明，當(dāng)系統(tǒng)在較大的面積比下運(yùn)行時(shí)，花瓣型噴嘴系統(tǒng)的性能優(yōu)于錐型噴嘴系統(tǒng)。
開(kāi)發(fā)一種無(wú)液泵或只有熱驅(qū)動(dòng)的噴射器制冷系統(tǒng)是一個(gè)很有吸引力的課題。如果實(shí)現(xiàn)了這一點(diǎn)，這些系統(tǒng)將不包含任何移動(dòng)部件。為此，黃等人嘗試開(kāi)發(fā)了一種具有熱泵效應(yīng)的噴射器冷卻系統(tǒng)。設(shè)計(jì)了一種集液體泵和發(fā)電機(jī)于一體的多功能發(fā)生器冷卻系統(tǒng)，消除了系統(tǒng)中泵的功能。研究人員對(duì)該系統(tǒng)進(jìn)行了測(cè)試，認(rèn)為該系統(tǒng)的設(shè)計(jì)是可行的，可以連續(xù)生產(chǎn)通過(guò)定期切換系統(tǒng)的兩個(gè)多功能發(fā)生器，達(dá)到冷卻效果。
埃姆斯介紹了一種設(shè)計(jì)用于噴射器制冷系統(tǒng)的新方法。該方法假定在超音速噴射器的擴(kuò)散段內(nèi)，流動(dòng)動(dòng)量以恒定的速率變化。該理論方法產(chǎn)生了一個(gè)擴(kuò)散器幾何形狀，消除了設(shè)計(jì)工況下dif-熔斷器內(nèi)的熱力學(xué)激波過(guò)程。
埃姆斯等人(1999)使用兩個(gè)主噴嘴和三個(gè)帶有混合室的擴(kuò)散器，研究了噴射器幾何形狀對(duì)蒸汽噴射泵制冷機(jī)性能的影響。實(shí)驗(yàn)結(jié)果表明，一次壓力比和主要出口面積與喉道面積的比值保持不變。
Chunnanond和Aphornratana(2004)研究了帶有錐形混合室的蒸汽噴射制冷機(jī)噴嘴幾何形狀和位置對(duì)性能的影響?；谒麄兊臏y(cè)試,他們表示,減少系統(tǒng)壓力,使用較小的噴嘴喉部面積(因此更高的面積比)和使用噴嘴混合室可以提高再保險(xiǎn)——更好的安全性和冷卻能力,提供關(guān)鍵的冷凝器壓力卻降低了。
在噴射器制冷系統(tǒng)中，無(wú)論使用何種工質(zhì)制冷劑，都要考慮噴射器的類型和二次裝置;為了獲得最大的性能，需要根據(jù)幾何參數(shù)和流量參數(shù)對(duì)系統(tǒng)的運(yùn)行條件進(jìn)行優(yōu)化。這些系統(tǒng)的最佳操作條件主要隨噴射器面積比的變化而變化,換句話說(shuō)，當(dāng)根據(jù)實(shí)際考慮，如熱源、制冷劑和廢熱介質(zhì)溫度來(lái)確定操作溫度或壓力時(shí)，應(yīng)該設(shè)計(jì)一個(gè)噴射器，使其面積比和噴嘴位置在選定的操作條件下是最佳的。本研究的目的是在大范圍內(nèi)確定R-123噴射器制冷系統(tǒng)的優(yōu)化運(yùn)行參數(shù)噴射器面積比的范圍。為了達(dá)到這個(gè)目的，三個(gè)不同的主噴嘴和兩個(gè)不同的混合室被制造成六個(gè)面積比。該系統(tǒng)通過(guò)將這些元件安裝在噴射器上進(jìn)行測(cè)試，噴射器內(nèi)的噴嘴可以軸向移動(dòng)。
在研究的第一部分中，首先在不同的發(fā)生器溫度下測(cè)量了這些噴嘴的一次蒸汽流量，然后以噴射器吸力室的最小壓力為準(zhǔn)則，分別找出各面積比下的最佳噴嘴位置。在研究的第二部分，介紹了在給定的面積比下，確定了最小二乘操作條件。為此，以最小面積比為例，給出了包括操作溫度或壓力對(duì)系統(tǒng)的影響在內(nèi)的實(shí)驗(yàn)結(jié)果
本研究中。即通過(guò)三個(gè)不同的實(shí)驗(yàn)來(lái)尋找各面積比的最佳工作點(diǎn)。結(jié)果表明，在蒸發(fā)器溫度為10℃，冷凝器壓力為125 kPa的情況下，得到了最佳的比表面積溫度。
最后，在相同的蒸發(fā)器和冷凝器條件下，通過(guò)反復(fù)試驗(yàn)，分別確定了其余五種面積比的最佳發(fā)電機(jī)性能。結(jié)果表明，各面積比只存在一個(gè)最優(yōu)的發(fā)電機(jī)溫度。所有表明隨發(fā)生器溫度變化的圍周曲線均在圖上完整地表示出來(lái)，并與實(shí)驗(yàn)結(jié)果進(jìn)行了比較, 以溫度為函數(shù)，得到了最佳性能曲線。此外,優(yōu)化實(shí)驗(yàn),結(jié)果比較與優(yōu)化分析結(jié)果為同一制冷劑通過(guò)使用文獻(xiàn)中給出的噴射流模型。
在固定的蒸發(fā)器和冷凝器條件下，最優(yōu)曲線幾乎隨發(fā)生器溫度線性增加。在一次噴管和擴(kuò)壓器效率為0.90時(shí)，最佳-的實(shí)驗(yàn)曲線與理論曲線一致。
圖1為實(shí)驗(yàn)裝置示意圖。該試驗(yàn)裝置的主要部件為蒸汽發(fā)生器、帶活動(dòng)噴嘴的噴射器、冷凝器、接收罐、膨脹閥、蒸發(fā)器、副冷卻器和滑片泵;制冷劑液體在裝置中循環(huán)。制冷系統(tǒng)的二次元件是熱水鍋爐、循環(huán)泵、預(yù)熱器、流經(jīng)它們的水流、測(cè)控裝置。實(shí)驗(yàn)設(shè)置和過(guò)程和噴射制冷系統(tǒng)的工作原理被詳細(xì)地描述在文獻(xiàn)中。讀數(shù)被監(jiān)測(cè)，并由計(jì)算機(jī)記錄。所有用于測(cè)量的設(shè)備都與測(cè)量范圍內(nèi)的數(shù)據(jù)采集系統(tǒng)校準(zhǔn)。蒸汽發(fā)生器的壓力測(cè)量的準(zhǔn)確性與土10 kPa,而其他系統(tǒng)中壓力測(cè)量的精度為土1 kPa。蒸發(fā)器的水流速測(cè)量的流量為0.03 L / min。的蒸汽發(fā)生器的溫度測(cè)量的溫度為 0.5攝氏度。在測(cè)量溫度、體積和時(shí)間周期誤差的基礎(chǔ)上，確定了流量的不確定度:
本研究采用的噴射器模型如圖2所示。在該模型混合室入口為圓形入口，混合室面積恒定，擴(kuò)散器為圓錐形。噴射器的設(shè)計(jì)基于恒定區(qū)噴射流模型在文獻(xiàn)(Yap?c?Ersoy, 2005;Ersoy, 1999)。實(shí)驗(yàn)中使用的六個(gè)噴射器的結(jié)構(gòu)和尺寸如表1所示。利用不同的三個(gè)超音速噴嘴和兩個(gè)帶擴(kuò)壓器的混合室，得到了這六種結(jié)構(gòu)。
性能系數(shù)(COP)是評(píng)價(jià)噴射器再制冷系統(tǒng)性能的最重要參數(shù)。它表示相對(duì)于輸入系統(tǒng)的能量而言的冷卻能力。與蒸汽發(fā)生器相比，制冷劑泵對(duì)系統(tǒng)的能量輸入過(guò)低。因此，在不考慮液壓泵功率的情況下，由下式計(jì)算出噴射器制冷系統(tǒng)的性能系數(shù)。
制冷量由(Qe/ Qg)/4(1)決定
_ e?m_ cwCpeTin - TexTe
在m_ cw為冷水的質(zhì)量流量。
蒸汽發(fā)生器的熱輸入是由
Q_ g?m_ pehex - - - - - - hinTg
為了確定每個(gè)噴嘴在不同發(fā)生器溫度下的一次蒸汽流量，進(jìn)行了實(shí)驗(yàn)，分別得到了一次蒸汽流量與發(fā)生器溫度的函數(shù)關(guān)系。實(shí)驗(yàn)中使用的三個(gè)噴嘴的一次蒸汽質(zhì)量流量隨發(fā)生器溫度的變化如圖3所示。利用最小二乘方法，得到了m_ p——Tg 對(duì)每個(gè)噴嘴的實(shí)驗(yàn)數(shù)據(jù)分別擬合曲線。建立了該曲線方程，并將其應(yīng)用于不同發(fā)電機(jī)溫度下的一次質(zhì)量流量預(yù)測(cè)。
相對(duì)噴嘴位置Ln/ dm 在0.5 < L的范圍內(nèi)n/ dm < 2根據(jù)Nahdi等人(1993)對(duì)制冷劑R-11的實(shí)驗(yàn)結(jié)果。在初步的試驗(yàn)中，在Pc ?125 kPa, Pg ?752.4 kPa (98 oC)和r ?9.97,使用制冷劑r - 123。因此，噴嘴位置被調(diào)整到ln ?5毫米(上游的圓柱形混合cham - ber)在每個(gè)實(shí)驗(yàn)和發(fā)電機(jī)最優(yōu),溫度測(cè)定首先分別通過(guò)保持蒸發(fā)器和冷凝器條件不變。在此基礎(chǔ)上，測(cè)量了在最適宜的發(fā)電機(jī)溫度下，吸力室壓力隨噴嘴位置的變化。最后對(duì)調(diào)整后的噴嘴距離是否在最佳范圍內(nèi)進(jìn)行了控制。實(shí)驗(yàn)結(jié)果如圖4所示。在確定最佳噴嘴位置時(shí)，如文獻(xiàn)(Hamner, 1978)中，以吸入室的最小壓力為準(zhǔn)則。
從上圖中可以清楚地看出，所有噴嘴調(diào)整后的最佳噴嘴位置范圍都在-10到5mm之間。
當(dāng)冷凝器和蒸發(fā)器的溫度/壓力已知時(shí)，就有可能找到最合適的時(shí)機(jī) \
為了確定臨界凝汽器壓力，進(jìn)而控制噴射器在混合室堵塞工況下的工作狀態(tài)，研究了凝汽器性能隨凝汽器壓力的變化規(guī)律g ?83 oC和Te ?10攝氏度。由圖6所示曲線可知，臨界凝汽器壓力約為126kpa，因此噴射器的工作點(diǎn)非常接近臨界工況，即系統(tǒng)在給定的面積比下，以最大的冷卻能力運(yùn)行。
圖7顯示了在相同的面積比下，性能指數(shù)隨蒸發(fā)器溫度的變化情況。實(shí)際上，最后的面積比實(shí)驗(yàn)是為了驗(yàn)證在上述操作條件下的性能值。這樣，在一個(gè)操作點(diǎn)上的經(jīng)驗(yàn)就重復(fù)了三次。由圖7可知，在T時(shí)刻的性能值e ?10度約為29%。另一方面，在這種工況下，性能值隨蒸發(fā)器溫度的升高而升高是預(yù)期的結(jié)果。
對(duì)其余五個(gè)噴射器重復(fù)上述實(shí)驗(yàn)步驟,實(shí)驗(yàn)確定了六種面積比的曲線，并以不同的標(biāo)記如圖8所示。從這些曲線可以清楚地看出，在固定的蒸發(fā)器和冷凝器溫度下，各面積比均存在一個(gè)最優(yōu)的發(fā)電機(jī)溫度。如圖所示，最佳性能曲線為與c峰值相切的直線g 曲線。最優(yōu)COP隨發(fā)生器溫度的升高呈近似線性增長(zhǎng)。COP的值opt 面積比由29%提高到41%左右，面積比由29%提高到6.56~11.45。從圖7的結(jié)果可以看出，COP較高opt 可通過(guò)提高蒸散器溫度來(lái)獲得。另一個(gè)重要的結(jié)果應(yīng)該在這里表示出來(lái)，那就是copt的斜率g 對(duì)于給定的面積比，在較低的溫度下比最佳的發(fā)電機(jī)溫度過(guò)高。因此，如果某一面積比下發(fā)電機(jī)的最優(yōu)運(yùn)行溫度與運(yùn)行溫度之差大于給定值，即10℃左右，則系統(tǒng)無(wú)法實(shí)現(xiàn)其功能。
在相同的工況下，最優(yōu)實(shí)驗(yàn)結(jié)果與最優(yōu)理論結(jié)果如圖9所示。最優(yōu)理論結(jié)果確定三種不同效率的噴射元素,使用文獻(xiàn)中提供的方法(Yap?c?Ersoy, 2005)。為了得到這些理論結(jié)果，將超音速噴嘴和擴(kuò)壓器的效率保持在圖9所示的值不變。當(dāng)效率為90%時(shí)，實(shí)驗(yàn)數(shù)據(jù)與理論數(shù)據(jù)吻合較好。
圖10顯示了最佳實(shí)驗(yàn)面積比和理論面積比隨發(fā)電機(jī)溫度的變化。實(shí)驗(yàn)結(jié)果和理論計(jì)算結(jié)果表明，最優(yōu)面積比隨回火溫度的增加呈近似線性增長(zhǎng)。換句話說(shuō)，較低的發(fā)生器溫度意味著較小的噴射器面積比。為了在固定的發(fā)生器溫度下獲得更高的COP，應(yīng)該將噴射器面積比與噴射器元件的效率一起增加。此外，理論計(jì)算結(jié)果與實(shí)驗(yàn)結(jié)果吻合較好，效率為90%。理論面積比略高于實(shí)驗(yàn)面積比。
在給定的蒸發(fā)器和冷凝器條件下，以R-123為工質(zhì)，采用帶柱形混合室的噴射器，從實(shí)驗(yàn)和理論上確定了最佳的發(fā)電機(jī)溫度與ejec- tor面積比的函數(shù)關(guān)系。對(duì)于給定的噴射器面積比，存在一個(gè)opti- mum發(fā)生器溫度，在此溫度下，從噴射器制冷系統(tǒng)獲得最大COP。當(dāng)發(fā)電機(jī)溫度低于與面積比相適應(yīng)的最佳溫度時(shí)，系統(tǒng)COP會(huì)急劇下降。在螺柱范圍內(nèi)，最優(yōu)面積比隨發(fā)電機(jī)溫度的升高幾乎呈線性增加。在理論計(jì)算中，將主噴嘴和擴(kuò)壓器的效率取為90%時(shí)，優(yōu)化實(shí)驗(yàn)結(jié)果與優(yōu)化理論結(jié)果吻合較好。
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