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中央空调水系统为何设计7℃-12℃供回水温度?都是洋首是瞻的谬误

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发表于 2018-1-25 17:24:50 | 显示全部楼层
本帖最后由 zhuangdijun 于 2018-1-29 09:03 编辑

专家

20年前他们都说我是空调专家
还说我是发明家
我照镜子
看见一个空调狗屁不通的董事经理得意地瞅着我
我告诉他:“你空调狗屁不通,就是发明了一个箱子而已。”
他对我不屑一顾。
我说:“你最多是一个木匠,绝对不是空调专家!”
他说:“是空箱子或是空调并不重要。”
“是木匠还是专家都无所谓。”
重要的是那个箱子能卖钱
有所谓的是它给企业赚了很多钱
镜子里的世界
钱赚多了狗屁不通也是专家
不会赚钱再通大不了是狗屁通而已
 楼主| 发表于 2018-1-26 13:11:51 | 显示全部楼层
3.2  空调机组在恶劣工况下运行的性能计算
考虑恶劣工况W^=(126.13, 33.31), DB=40℃, RH=70%, WB=34.58℃。室内负荷计算如下图: 162939nz31hv8z7tn0917v.png.thumb.png
表3的前两行显示,对于无新风渗透的集中新风空调系统而言,室内相对湿度对负荷的影响极小。在平衡状态下,室内热负荷+经处理新风的负荷 = 二级空调冷量 = 冷水机组冷量。经新风机组处理过的新风无法处理到室内等焓线上,尚需处理的负荷约4kW。通过反复估算,供水温度约10℃,室内空气状态约 28℃/58%时,上述三项都约为22kW。
        当室内空气状态为28℃/58% ,则 Δ=(18.136, 4.389/2.5)/1.278 =(14.191, 1.374)。应用近等相对湿度定律,恶劣工况下的室内平衡相对湿度可通过图3中的送风点 O1上加负荷向量计算,O1+ Δ = (37.99, 9.45) + (14.191, 1.374)  = (52.181, 10.824)。室内相对湿度=56.9%。

         详细计算公式为 Ni+1= F[0.262φ(W^) + 0.738Ni] + Δ    i=1,2,3……

风量=3834m3/h=1.278kg/s,负荷向量 Δ=(17.913,4.387/2.5)/1.278=(14.016, 1.373),空调机组 φ 和 F 的水流量分别为0.57l/s和0.87l/s。取 N1=(63.91,13.802), DB=28.4℃, RH=57%,新风送风点K= φ(W^)=(75.85,19.68), 见图5左图。
C1=0.262K+0.738N1 = (67.04, 15.344), DB=27.63℃, WB=22.7℃; 163116mfjxtx5t5xxixxh0.png.thumb.png
O1=F(C1)= (50.05, 12.5), 图5右图;
N2= O1+Δ=(50.05, 12.5)+( 14.016, 1.373)= (64.066, 13.873), DB=28.38℃, RH=57.4%。
N2 不等于N1 ,所以迭代继续。
C2=0.262K+0.738N2= (67.16, 15.397), DB=27.61℃, WB=22.73℃;
O2=F(C2)= (50.1, 12.52), 见图6左图;
N3= O2+Δ=(50.1, 12.52)+( 14.016, 1.373) = (64.116, 13.893), DB=28.37℃, RH=57.5%。  
N3 不等于 N2 ,所以迭代继续。
C3=0.262K+0.738N3= (67.19, 15.411), DB=27.61℃, WB=22.74℃;
O3=F(C3)= (50.12, 12.52), 见图6右图;
N4= O3+Δ=(50.12, 12.52)+( 14.016, 1.373)= (64.136, 13.893), DB=28.37℃, RH=57.5%。  
因此N4= N3,平衡室内状态点为DB=28.37℃, RH=57.5%。二级空调机组的冷量为21.27kW。以冷冻水温每增加1℃则冷水机组冷量增加3.25%估算,当出水温度为10℃时,冷水机组冷量=19.5*1.0975 =21.4kW。
163240taewpiaikea1vdgx.png.thumb.png
4  空调机组在梅雨季节的运行性能
4.1 空调机组在梅雨季节运行的性能计算
        取梅雨季节室外工况为W=(78.99, 20.286), DB=27℃, RH=90%,通过反复估算,当供水温度为7℃且室内空气状态约 21.6℃/60%, 则室内热负荷和空调机组冷量皆约为14.5kW。负荷向量 Δ=(14.520, 2.940/2.5)/1.278 = (11.362, 0.92)。
        表4显示在梅雨节的室内负荷。
163409pvni77r7hi447p30.png.thumb.png
当室内空气状态为21.6℃/61%。应用近等相对湿度定律,梅雨工况下的室内平衡相对湿度可通过在图6左图中的送风点 O2上加负荷向量计算,O2 + Δ = (50.1, 12.52) + (11.362, 0.92) = (61.462, 13.44)。相对湿度=60.9%。
准确计算取N1=(46.67, 9.787). DB=21.6℃, RH=61%。负荷向量Δ= (11.362, 0.92),当空调机组进水温度为 7℃,新风送风点 K= φ(W)=(48.06,12.12),见图7左图。
163609qhp299xh4h3ih2pg.png.thumb.png
O1=F (C1)= (35.37, 8.87), 见图7右图;
N2= O1+Δ=(35.37, 8.87) + (11.362, 0.92)= (46.732,  9.79), DB=21.65℃, RH=60.8%.
N2 不等于N1 ,所以迭代继续。 163730dp00gg22egna1o90.png.thumb.png
C2=0.262K+0.738N2=(47.08, 10.401), DB=20.48℃, WB=16.77℃;  
O2=F (C2)= (35.42, 8.88), 见图8左图;
N3= O2+Δ=(35.42, 8.88) + (11.362, 0.92)= (46.782,  9.8), DB=21.67℃, RH=60.8%.
N3 不等于N2 ,所以迭代继续。
C3=0.262K+0.738N3=(47.04, 10.394), DB=20.5℃, WB=16.78℃;
O3=F (C3)= (35.42, 8.88) = O2。
因此N4= N3,平衡室内状态点为DB=21.67℃, RH=60.8%。二级空调机组的冷量为21.27kW。

4.2 梅雨季节低档风速运行
在梅雨季节,以室内温度 21.6℃运行肯定是不合理的。常规做法是室内以低档风速运行,即二级空调机组风速降低到3834*0.50=1917m3/h=0.639kg/s。新风和二级空调机组的冷冻水流量维持在 0.57 l/s 和0.87l/s 。
发表于 2018-1-27 11:21:42 | 显示全部楼层
邂逅

第一次的邂逅
他又是前门奔到后厅
又是后庭奔到前门
热情主动积极诚恳
有呼必迎
有求必应
在大伙的眼里
他是我最忠实的朋友
他的忠实让我忘记了
我是房子的主人
发表于 2018-1-27 20:29:22 | 显示全部楼层
本帖最后由 zhuangdijun 于 2018-1-29 06:29 编辑

第二次的邂逅
我教他打猎
他依旧有呼必迎
然而却不学打枪
也不会射箭
衔着一只野鸡
吹嘘捉到十只
逮到一只田鼠
叫到十里外都听得见
他不钟意学射击
他钟意的是
我手中的猎物
发表于 2018-1-28 11:00:48 | 显示全部楼层
本帖最后由 zhuangdijun 于 2018-1-29 06:30 编辑

第三次的邂逅
我走我的路
快节奏的怒吼吓我一跳
我两步一回头
战战兢兢地往前走
直到怒吼的节奏放缓
惊魂稍定
我恍然顿悟
原来我没有意识路边的一只死猫
或他身后还有一小撮后脚短的哥们
我更是毫无意识
我经过的是他家门口
发表于 2018-1-29 06:31:35 | 显示全部楼层
本帖最后由 zhuangdijun 于 2018-1-29 06:37 编辑

第四次的邂逅
他竟然又变成谦卑得出奇
热情主动
都露在灿烂的笑脸
三步当两步冲上来握我的手
然后是
温顺得就像一只小猫咪
是万有引力造成这180度的转变?
是我突如其来的魅力?
我并没开法拉利
或是开兰博基尼
只能是他的顿悟
虽然都说江山易改本性难移
但肯定是顿悟
不容置疑
终于我转过来了
原来是灯下黑
我终于看到了他的顿悟
我看到了
他早已盯上的
我手中那块带肉的猪骨头
发表于 2018-1-31 08:55:35 | 显示全部楼层
本帖最后由 zhuangdijun 于 2018-1-31 13:18 编辑

我花了三个月的时间整了三篇论文,原想到国外学报去发表,两周前一位北京所的高工来访,说好像看到日本人在日本国内做7-15℃大温差空调系统的。翻译了一篇成中文,在本论坛和“创新新语”发布。截上未翻译的部分7-15℃末端设计和合肥所测试对比的,说明i-d向量空间理论的实际可行和简单有效的节能应用。能理论和全凭经验的差别是开灯和抹黑操作的差别。

table4.PNG

5  Standard 7-12℃ Fan-coil Units and its 7-15℃ Counterpart:
Thus far we have designed a 7-15℃ air-conditioning unit which out performs the conventional 7-12℃ unit. However, to be accurate the above discussed air-conditioning units are academic or only meant for the purpose of theoretical discussion. In practice take the standard 7-12℃ fan-coil units for example, they have been in existence for generations and is designed with indoor requirement of 25℃/50%. It is improper to design 7-15℃ fan-coil units with indoor relative humidity less than 50% which is already too dry and wasting energy.
For practical purposes our strategy is to design the 7-15℃ fan-coil units with indoor relative humidity between 50-55% with comfort as well as energy and cost saving as objectives.
         
           5.1  Design and Performance
In what follows we choose 7-12℃ and 7-15℃ FP68 fan-coil units designed by Pingri Technologies and manufactured by BG Corporation for analysis. Their designs are shown in figure 9.

           
Figure 9: design of conventional 7-12℃ and 7-15℃ FP68  fan-coil units F12 and F15

To calculate the performance of 7-12℃ FP68 fan-coil unit at the requirement of the Nanjing office in Section 1 and at the design ambient of Nanjing we estimate indoor air at 24.2℃/53% by trial and error. So from figure 10 total heat load=19.889kW.  To select the number of FP68 fan-coil units required for the project, the common practice to choose total rated capacity of fan-coil units = 1.2*19.889kW=23.867kW. Thus number of FP68 fan-coil units required = 23.867/3.989=5.983. Although number of units cannot be fractional it suffice to calculate the total air volume for the purpose of theoretical calculation.
Now air-volume = 5.983*685m³/h=4098m³/h=1.366kg/s. heat load vector = (19.889, 4.879/2.5)/1.366 = (14.56, 1.429). Choose F12(N) from left diagram of figure 9, F12(N) + Δ = (38.02, 9.114)+ (14.56, 1.429) = (52.58, 10.543), DB=25.47℃, RH=52%. By near constant indoor relative humidity principle relative humidity is around 52%.   

figure 10.PNG

To calculate the performance of 7-15℃ FP68 fan-coil unit at the environment of the Nanjing office in Section 1 and at the design ambient of Nanjing, air-volume = 5.983*670m³/h=4009m³/h=1.336kg/s. Estimate indoor air to be 23.85℃/54% by trial and error, heat load vector = (20.1, 4.884/2.5)/1.336 = (15.045, 1.462). Choose F15(N) from right diagram of figure 9, F15(N) + Δ = (37.93, 9.28)+(15.045, 1.462) = (52.975, 10.742), RH=53.3%. By near constant indoor relative humidity principle relative humidity is around 53.3%.
Calculation of performance of 7-12℃ and 7-15℃ FP68 fan-coil units are shown in figure 11.

Equilibrium indoor air of 7-12℃ and 7-15℃ FP68 fan-coil units at design ambient of Nanjing and indoor heat loads of the office in Section 1 are DB=24.23℃, RH=52.6% and DB=23.84℃, RH=54.1% respectively.

5.2  Test Results
Test results of the 7-15℃ FP68 fan-coil unit from Hefei General Machinery & Electrical Products Inspection Institute is shown in figure13. Capacity=4.256kW which is larger than the calculated value of 4.097kW shown in figure 9. The reason being the chilled water flow rate used for testing is 457l/h (derived from pressure drop of 5.3kPa) and inlet air NT=(55.71, 11.152), DB=27.01℃, WB=19.51% whereas design flow rate is only 442l/h and inlet air is DB=27℃, WB=19.5%. With inlet air NT capacity of unit is 4.124kW.
Capacity deviation due to increased of chilled water flow rate is 457/442-1=3.4%. Deduct 3.4% from the test result, capacity=4.256/1.034= 4.116kW. This differs from calculated value of 4.12kW by less than 0.2%.  The test result is better than expectation.
Discharged air O is DB=13.64℃, WB=12.98℃ or O=(36.57, 9.027). By principle of near constant indoor relative humidity, indoor relative humidity derived from test result at Δ= (15.018, 1.46) is O + Δ=(36.57, 9.027) + (15.018, 1.46) = (51.588, 10.487), DB=24.65℃, RH=54.3%. Indoor relative humidity matches calculated value very well.

figure12.PNG figure12a.PNG

figure13.PNG

Summary and discussion:
By viewing indoor heat loads as a two dimensional vector Δ on the i-d vector space, controlling of indoor relative humidity is reduced to regulating the magnitude of Δ. Increase the magnitude of Δ by slight decrease of air volume, we are able to redesign a 7-12℃ chilled water air-conditioning system using 7-15℃ chilled water which out performs its 7-12℃ counterpart in the sense of lower indoor relative humidity and temperature.
Table 4 compares the performance of the 7-15℃ and 7-12℃ designs. It is transparent that the former out-performs the latter. To be exact, both the indoor temperatures and relative humidity of the former are lower than those of the latter in design as well as adverse and rainy ambient.
Test results of 7-15℃ fan-coil unit in figure 13 also matches the calculated value extremely well. In particular, by means of near constant indoor relative humidity principle, it can be estimated from discharged air of testing that indoor relative humidity for the office in Nanjing is 54.3% at design heat loads which is very close to the calculated value of 54.1%.
We claim that we have cleared the myth that when chilled water temperature is increased indoor relative humidity may be out of control.  We believe that it is this misunderstanding of control of indoor relative humidity that stops the design of comfort air-conditioning system beyond 7-12℃ and 5-13℃ and this phobia have incurred at least an extra 6.5% of energy for generations.
To design a 7-15℃ air-conditioning unit to out perform its 7-12 ℃ counterpart is purely academic. In practice the objective of 7-15℃ comfort air-conditioning design should be comfort, energy efficiency and cost. Take for example the  7-12℃ chilled water fan-coil units, traditionally they are designed  with indoor requirement of  25℃/50% which is too dry especially for north western region of China and not energy efficient. Hence it is definitely improper to design the 7-15℃ fan-coil units with indoor relative humidity lower than 50%. The Pingri 7-15℃ fan-coil units are designed with indoor relative humidity slightly higher than its 7-12 ℃ counterpart in order to achieve the objectives of a more comfortable, more energy efficient and cost effective new product that is acceptable by the market.
The advantages of 7-15℃ design as compared to 7-12℃ or 5-13℃ designs is obvious. Compared to 5-13℃ design inlet chilled water temperature of 7-15℃ design is higher by 2℃ meaning the air-conditioning system can save energy by about 6.5%.
Compared to 7-12℃ design the advantage of 7-15℃ design is more obvious. By common sense it is the aggregate of the advantages of 5-13℃ design over 7-12℃ design on top of the 6.5% energy saving described above. The most significant advantage of small flow rate design is chilled water flow rate is reduced from 694l/h to 442l/h or by 36%. Thus operation of chilled water pumps saves energy by nearly 36% bearing in mind that most chilled water pumps operate for much longer hour than chillers.
        About twenty small projects of 7-15℃ comfort air-conditioning systems of fan-coil units and modular air-cooled chillers have been installed in Jiaxing area of Zhejiang Province. Besides energy saving customers’ feedback is that its cooling ability at adverse ambient is apparently superior to its 7-12 ℃ counterpart
  
References:
[1] Gatley, D.P. “Psychrometric chart celebrates 100th anniversary.” ASHRAE Journal 46(11) 2004: 16 – 20
[2] Carrier, WH. Rational Psychrometric Formulae.ASME-Transactions.1911 Bd 33: pp.1005-1053
[3] Dhar M, Soedel W.,  Transient analysis of a vapor compression refrigeration system, Proceeding of 25th International Congress of  Refrigeration, Venice, Italy, 1979.
[4] Bendapudi S, and Braun, J.E.A.  Review of literature on dynamic models of vapor compression equipment, ASHRAE 2002: 1043-RP, TC 4.11.
[5] Ding G.L., Recent Developments in simulation techniques for vapor-compression refrigeration systems, International of Refrigeration 2007: 30 1119-1133.
[6] Zhuang D.J., Analytical Calculation of Performance of Comfort Air-conditioning System With No Infiltration of Fresh Air
[7] Zhuang D.J., i-d vector space and dehumidification problem in rainy season of Nanjing

figure11.PNG
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