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暖通空调论文3000字（暖通空调论文3000字范文）
发布时间：2023-04-25 02:59:50
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大家好！今天让创意岭的小编来大家介绍下关于暖通空调论文3000字的问题，以下是小编对此问题的归纳整理，让我们一起来看看吧。
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本文目录:
一、关于暖通空调设计的一些思考
1. 一个软件的安装一定要打开试用一下，没问题才是安装成功。
2. 标准、条文上的黑体粗字是必须执行的。
3. 冷冻水：供实、虚回
4. 画出来的图要能用
5. 坡度：供、回靠水泵送上去
6. 自动放气阀：位于给水管最高点，一般放在厨房、卫生间，因为是铝板容易拆。安装高度注意看图上标高
7.注意检查口
8. 厨房、卫生间吊的顶比石膏板低
9. 回水管：高 供水管：低
10. 预留套管位置是暖通与结构碰后的结果
11. 对水管的位置没有明确的规范要求，但要跟土木碰一碰，确保结构的没问题
12. 冷凝水管从梁下走，从地漏排走
13. 穿梁的预埋套管一般不可以离的太近，一般为200的间距，太近的话中间穿不了钢筋，结构的稳固性不好
14. 空调不能对着头吹
15. 管路少穿墙
16. 吊顶美观也很重要
17. 冷凝水坡度一般都是0.003（千分之三），也可以是0.005
18. 画的图一定要清晰地表达意思意图
19. 图纸说明：空调设计包括：依据、概况、参数（室内、室外、维护结构）、冷热负荷、空调设计系统（空调、自控）、环保
（1）施工说明
（2）图例
（3）所用标准图集
（4）主要设备材料表
（5）图纸目录
20. 可以通过看水系统图来研究系统结构
21. 要保持足够的劲头、手速（画图就要快速画好，不要慢吞吞）
22. 看管间距方法：[管径/2+保温层厚度（查规范）]*2
23. 画图时可以看看3维版，有更直观印象（画图时脑内要装换成真正的实物，这样根据实际去考虑规范规定的事）
24. 管路能穿剪力墙就不要穿梁
25. 风量按换气次数计算，若为双层地下车库则要按每辆车的量算
26. 热力入口在地下室
27. 大样图上阀门、保温层厚度在图集上有
28. 平面图上的阀门可以自由缩放、斜着放，表达出这个就行了(不能太小了，太小图上看不见)
29. 常用风盘制冷量：
FP-8：4.5KW
FP-6.3：3.5KW
FP-5：2.8KW
FP-3.5：2.0KW
制冷量计算时，把这层或这个系统的每个风盘制冷量代数相加，再乘以同时使用系数0.65，即可以的得到这层活这个系统的制冷量。
经验上一般小区不会同时开所有的空调，所以不用按空调制冷量算，一般取140W/m2已经很大了。按小区来取得话算50W/m2.
30. 标了标注的就是必须实际照做的
没有标注的是平面图，是概念图（按照那个摆，但具体位置不是）
31. 只接一条线的时候，比摩阻要控制在250以内，控制比摩阻是为了减小沿程损失。
正常比摩阻在100-300范围内，这是一个可能出现的范围，所有比摩阻值必须控制在250以下
32. 水系统布置：
（1） 布置原则（阀门种类，什么时候安）
（2） 水系统的承压能力
规范上有一般系统的承压能力，从而考虑是否需要竖向分区
（3） 水力计算
算水管水流量
算水管管径
水系统的沿程损失计算
水系统的管段局部损失
33. 系统工作压力=静压+动压
承压能力讲的是设备承压
34. 鸿业软件上的“分支计算”必须是断线，且有头有尾可以计算，所以一般重新画一个专门用于计算的立管系统
35. 梁图上穿梁的部分才要加套管
36. 画完图的最后要检查一下有没有问题
37. 穿梁图上考虑入户地暖管走地下室
38. 画图要心中有成算，手上看起来慢其实快。动手改起来要完全改完再该别的
39. 别人讲的都是暂时这样或者经验这样，要以规范和图集为准
40. 有时候不是对于错，只是个人的习惯问题
41. 标注只要表达清楚了就可以，没有规定一定在哪个方向。
42. 标注时考虑大小、位置、高度
43. 冷凝水的高度和位置一般不表示，因为太细，一般画出图即可，位置施工时会自己协商
44. 套管的具体位置按预埋套管图上的位置。
45. 即使有大样，平面图上阀门也要画全，实在画不下，注上详见大样
46. 一层会有指北针，是建筑图上给的
47. 水管水力计算：
（1） 根据选型风盘的功率，在乘上同时使用系数即得负荷值
（2） 在旁边地方画一个风盘，cx修改风盘的参数（可以一层风盘的负荷值都用这一个风盘负荷值来代替）
（3） 画一个给水，一条排水管，选自动设备连管
（4） 选 水管——分支计算，点最下方管出“初算”结果
（5） 让比摩阻降下来，按流速计算，改部分管径值（尤其是最末端管，放大些）
（6） 点重新计算，没什么问题就标注
48. 回头再检查一遍：想想工人拿到我的图怎样理解每个位置
想想我是不是都表达清楚了
随便取一小块，看看我知不知道这能不能安装在别的位置
49. 只有自动排气阀的立管 DN20
冷凝水管 de25 （de32） i=0.003
地漏de25
平面图上自动排气阀在给水管上 DN15（户内）
末端截止阀 DN25
泄水阀（管）DN50 （热力入口）
排污阀 DN50 （热力入口）
旁通阀 DN80
热力入口自动排气阀DN20
50. 每个FP 都要一个电动二通阀
热能表、自力式压差控制阀每户一个
51. 立管高度低于60m一般不用补偿
延长量=t*l*线性伸缩系数
线性伸缩系数取0.012
如果不作补偿，热胀冷缩，立管太长，容易把水表扯下来或是漏水
一般把延长量控制在2cm以下
施工温差在3℃左右
33（3m层高*11层）*55*0.012=21.78=2cm
52. 波纹补偿器可以放在楼层面上方，便于检修
53. 算负荷：
（1） 用负荷工具条中房间管理算出每个房间、外墙、窗大小，记下来
（2） 负荷计算中创建，该气象参数，一定要选在“新规范《GB50736-2012》气象参数”上
（3） 改维护材料结构（看节能书最下面汇总的K值，注意区分冬夏季，冬天的可以和节能书上不一致，夏天的一定要和节能书上的一致）
（4） 25#——楼层属性——选关联层、关键层、相同层
（5） 改完一定要按刷新数据（注意设层高）
（6） 每个房间改名称（体现功能）
相同房间要汇总，该房间面积，设备灯光不改（随意），人员取0.03人/m2，新风量取换气次数*h（即为单位面积新风）
54. 写在图纸材料表上的外墙、窗等材料取主要部分
55. 管线走线时注意顶板高、翻管等问题（还要有一定的预留空间）
56. 窗户LC2418指宽24 高18
57. Kv就是算出来的流量
Kvs流量系数，指阀门两端压差为0.1MPa,水密度为1g/cm2，阀门全开时的流量是调节阀的重要参数，反映调节阀的容量
58. 风机盘管水流量：根据风盘标称的供冷量除1.163再除5得出来标准水流量
59. CAD去水印的方法:
法一：导成pdf的cad。首先另存为dxf格式，再打开这个dxf格式的文件，点击打印——打印机（cad to pdf）——打印样式（monochrome.ctb）——图纸尺寸（若是加长版，在 特性 中自定义图纸尺寸）——打印范围（窗口）——居中打印——预览
法二：乱刀小软件。命令ap——最上方的框内选择BladeR18-x64.arx 文件——点击加载
——加载成功以后再打印，就没有印戳了
二、建筑设计的本科生毕业论文
随着社会经济的不断发展,人们的生活水平也有了很大的提高,对于住房条件的要求也越来越高,为了满足居民的住房需求,我国的建筑业加大了房屋建筑设计的规模和力度。下文是我为大家搜集整理的建筑设计的本科生毕业论文的内容，欢迎大家阅读参考!
建筑设计的本科生毕业论文篇1
浅谈建筑设计中节能建筑设计
摘要：当今社会经济飞速发展，做为我国国民经济三大支柱产业之一的建筑业，在能源消耗中占的比重越来越大，在当下大力倡导节能环保的大环境下，节能建筑做为共同关注的重要问题被提上日程。本文阐述了建筑设计中节能设计的概念、现状和优势，并提出了节能建筑设计中的几点策略，充分利用自然能，降低不可再生能源消耗，促进我国建筑可持续发展。
关键词：节能建筑;设计;应用
随着我国经济快速增长，各项建设取得巨大成就的同时，我国也付出了巨大的资源和环境被破坏的代价，经济发展与资源环境被破坏的矛盾日趋尖锐，群众对环境污染问题反应强烈，能源的短缺已不容忽视，节约能源与环境保护已受到世界性的普遍关注，在我国亦不例外。目前，全世界有近30%的能源消耗在建筑物上，长此以往，将严重影响世界经济的可持续发展。因此，我们必须从可持续发展的战略出发，使建筑尽可能少地消耗不可再生资源，降低对外界环境的污染及破坏，并为使用者提供健康、舒适与自然和谐的工作及生活空间。
1节能建筑概念
节能建筑是指遵循气候设计和节能的基本方法，对建筑规划分区、群体和单体、建筑朝向、间距、太阳辐射、风向以及外部空间环境进行研究后，设计出的低能耗建筑，其主要指标有：建筑规划和平面布局要有利于自然通风，绿化率不低于35%;建筑间距应保证每户至少有一个居住空间在大寒日能获得满窗日照2小时等。目前节能建筑已逐渐成为国际建筑界的主流趋势。一个经常被忽略的事实是：建筑在能源消耗总量中，几乎占到了70%，这一比例远远高于运输和工业领域。在发展低碳经济的道路上，建筑的“节能”和“低碳”注定成为绕不开的话题。
2 节能建筑设计的现状和优势
2.1节能建筑研究及应用现状
节能建筑已逐渐成为国际建筑界的主流趋势。在中国，节能建筑思想也越来越受到重视，并已写进国家的发展规划中。目前对于节能建筑研究较多的是建筑外窗、玻璃幕墙的应用，而对外墙、屋顶以及楼地板的研究较为欠缺。另外，夏热冬冷地区的研究较寒冷地区、严寒地区的研究多，主要是因为夏热冬冷地区采暖和空调能耗均较高，节能设计需同时考虑围护结构的保温和隔热性能，而这两者是相互矛盾的，所以，要想达到既保温又隔热的目的，有很多困难需要解决。
2.2低碳节能建筑的优势分析
2.2.1采用地毯式的建筑能使能耗显著降低。据统计，建筑在建造和使用过程中可消耗50%的能源，并产生34%的环境污染物。节能建筑则大大减少了能耗，和既有建筑相比，它的耗能可降低70%～80%。所以低碳式建筑更有利于环境的保护。
2.2.2节能建筑产生出新的建筑美学。一般的建筑采用的是商品化的生产技术，建造过程的标准化、产业化，造成了大江南北建筑风貌大同小异、千城一面，而节能建筑强调的是突出本地的文化、本地的原材料，尊重本地的自然、本地的气候条件，这样在风格上完全是本地化的，并由此产生了新的建筑美学。节能建筑向大自然的索取最小，这样的建筑，让人在体验新建筑美感的同时，能更好地享受健康舒适的生活。
2.2.3节能建筑环保理念贯穿始终。传统建筑多是在建造过程或使用过程中，考虑到环境问题，而节能建筑强调的是从原材料的开采、加工、运输、使用，直至建筑物的废弃、拆除的全过程，节能、环保理念贯彻始终，强调建筑要对全人类、对地球负责。
3 推进节能建筑的措施
3.1 建筑规划的节能设计
3.1.1 合理选址
建筑选址主要是根据当地的气候、地质、水质、地形及周围环境条件等因素的综合状况来确定。建筑设计中，既要使建筑在其整个生命周期中保持适宜的微气候环境，为建筑节能创造条件，同时又要不破坏整体生态环境的平衡。
3.1.2 正确选择朝向
日照及朝向选择的原则是冬季能获得足够的日照并避开主导风向，夏季能利用自然通风并防止太阳辐射。然而建筑的朝向、方位以及建筑总平面的设计应考虑多方面的因素，建筑受到社会历史文化、地形、城市规划、道路、环境等条件的制约，要想使建筑物的朝向均满足夏季防热和冬季保温是困难的，因此，只能权衡各个因素之间的得失，找到一个平衡点，选择出这一地区建筑的最佳朝向和较好朝向，尽量避免东西向日晒。
3.2 建筑围护结构节能设计
建筑围护结构组成部分(屋顶、外墙、门和窗、遮阳等设施)的设计对建筑能耗与用户所处热舒适环境有根本的影响。一般增大围护结构的费用仅为总投资的 3%～6%，而节能却可达 20%～40%。通过改善建筑物围护结构的热工性能，在夏季可减少室外热量传入室内，在冬季可减少室内热量的流失，使建筑热环境得以改善，从而减少建筑冷、热消耗。
3.2.1 屋顶节能
屋顶是住宅第五立面，对建筑造型起着重要作用。住宅做斜坡顶屋面，可借助屋面坡度与日照斜率相接近的特点，可再降低住宅顶层的层高。在维持平屋面住宅日照间距的条件下，既取得了改变建筑轮廓、有效地解决了屋面防水和扩大屋顶部位使用空间的效果;也减少了住宅之间的日照间距，节约了建设用地。平屋顶可采用北向的退台，既获得露天活动空间，也可缩小日照间距。
3.2.2 墙体节能
墙体是建筑外围护结构的主体，其功能主要是承重、防水、防潮、隔热、保温。其所用材料的保温性能直接影响建筑的耗热量，一般情况下，单一墙体材料往往难以同时满足保温、隔热要求，因而在节能的前提下，应进一步推广空心砖墙及其复合墙体技术。其一般做法是，用砖或钢筋混凝土作承重墙，并与绝热材料复合。
3.3 建筑材料节能设计
合理选用建筑节能材料也是全面建筑节能的一个重要方面。建筑材料的选择应遵循健康、高效、经济、节能的原则。一方面，随着科技的发展，大量的新型高效材料不断被研制并应用到建筑设计中去，更好地起到节能效果。另一方面，要结合当地的实际情况，发掘出一些地方节能材料，更好地应用到建筑节能中去。 3.4 利用新能源
可再生能源在暖通空调系统中的应用包括：太阳能的应用、自然通风的应用、地下水的应用、地热(冷)的应用等。
3.4.1 太阳能的应用地球拦截的太阳能辐射相当于目前全球电力消费量的1500倍，而在现有技术、经济条件下可供开发利用的太阳能，只占理论资源量的很小一部分。太阳能在暖通空调中的应用主要有太阳能采暖和太阳能制冷两个方面。
①太阳能采暖
太阳能采暖用电作为辅助能源，驱动用太阳能加热的水在管道中循环流动向房间供热。
②太阳能制冷
太阳能制冷主要包括太阳能压缩式制冷、太阳能吸收式制冷和太阳能吸附式制冷。太阳能压缩式制冷研究的重点是如何将太阳能有效地转换成电能，再用电能去驱动压缩式制冷系统。太阳能吸附式制冷是将系统中的加热器和冷却器去掉，将太阳能集热器与吸附床合二为一，冷却功能则利用夜间室外空气的自然冷却来完成。
3.4.2 自然风的应用
自然风的供冷是可再生能源在暖通空调应用中的重要组成部分。当室外空气的焓值和温度低于室内时，在供冷期内就可以利用室外风所带有的自然冷量来全部或部分满足室内冷负荷的需要。通常，这种情况出现在供冷期的过渡季和夜间，可采用的方法为新风直接供冷和夜间通风蓄冷。由于利用了自然风提供建筑所需要的冷量，与常规空调系统相比，在运行中不用电或少用电，既节约能源，又减少对环境的污染，同时也改善了室内空气品质。
3.4.3 地下水的应用
地下水由于地层的隔热作用，其温度受气温影响很小。在暖通空调中，有些地下水可以直接作为冷源，更是热泵良好的低位热源。所以水源热泵有着良好的节能前景。
4 结束语
保护环境、有效利用自然能源、削减能源负荷是新时期实现可持续发展的重要要求之一，建筑设计中应用节能技术是对可持续发展这一理念的最好回应，节能建筑将成为今后建筑设计的主打方向，建筑节能工程作为建设领域的新方向已成为我们既定的基本国策，我们应深刻认识到节能设计的重要性，从自身出发、从实际出发，设计出与实际生活和社会相适应的设计，努力使建筑能耗最低化，大力发展节能建筑，提高能源利用率，为加快建设资源节约型，环境友好型社会做贡献。
参考文献：
[1]刘加平，武六元. 建筑节能与建筑设计中的新能源利用[J]能源工程，2001
[2]周炜.小议建筑节能设计[J]陕西建筑，2008
建筑设计的本科生毕业论文篇2
浅析建筑设计与城市设计
摘要：城市是历史发展的产物,是集人类文明与传统于一身的聚集体,其结构庞大复杂,内容包罗万象，建筑是城市的重要组成部分，本文浅析其二者设计之间的相互关系。
关键词：城市;建筑;设计
城市是人们的家,如何让自己的家变得更美好,人们希望创造一种住着舒适、用着方便、看着美观的充满生机的独特的城市空间。对现状的无奈与对未来美好的渴望给人们提供了思考研究与创造的机会。建筑师、规划师和景观师纷纷研究各种理论与设计方法,以期能为城市添彩,创造更为舒适、更为人性化的城市空间。
一、 城市设计与城市规划
谈起城市设计与城市规划的关系,首先引用著名建筑师沙里宁在《论城市》一书中对城市设计的含义归纳:“城市设计是三维空间,而城市规划是二维空间,两者都是为居民创造一个良好的有秩序的生活环境。”
城市规划以一个城市的宏观发展为目标,它更多的考虑城市的工业化,商业化,现代化,要飞速发展,要为经济服务,提高城市运作的效率,所以要求有更高,更快,更先进,更现代化,更信息化的“硬”环境;而“城市设计”理念的出现,则是“人本主义”对高速工业化的反叛,它更应该注重人文的,文化的,美学的,自然的“软”环境。
但是两者也有共通性,城市设计既为城市规划提供思路和形象化的发展目标,也为建筑设计提供前提和轮廓,城市设计具有更多的立体性、可操作性和示意性,其主体就是空间环境设计。无论是建筑群的组合还是城市的空间设计,都有一种内在的秩序或结构作为联系的纽带。城市设计由注重城市的肌理、构图注重人的存在与活动,越来越体现出对主体城市的认识。从城市发展史中可以看到,人的主观活动往往起决定作用,在现代城市规划与设计过程中设计结果与规划结果并不一定完全吻合,所以它们之间需要相互反馈、相互调整。
二、 城市设计与建筑设计
建筑是组成城市的基本细胞,精制而富有特色的建筑最能展示城市的艺术性。建筑的设计手法现在基本有3 种:模仿、再生、创新。功能成为建筑设计的主题,形式只是外皮的建筑创作过程正在被建筑师们推敲。越来越多的迹象表明,许多建筑师正在研究建筑的基本组成元素,然后在某种法则的指导下,进行建筑的重组,从而展现崭新的建筑形式。建筑的外皮也成为单独研究的一个课题,其保温、承重、生态、维护等诸多功能被分层研究,再进行组合,形成有独特内涵的外皮或立面形式。这种解构主义的创作手法更立意于建筑的本原,创造出理性而非感性的建筑。这种建筑形式先思而后建,比施工图设计的建筑形式更利于城市整体的艺术环境。
城市设计是一门正逐步完善和发展的综合性学科,是一门在实践中安排城市发展规划与建筑设计、景观设计相对关联的实用性学科,它具有相对独立的基本原理和方法,它主要解决的是城市的面和线问题。建筑设计是在城市规划的前提下,根据建设任务要求和工程技术条件进行全面设想,并根据其功能具体确定建筑物的空间组合形式和详细尺寸,构造及材料做法。它也具有相对独立的基本原理和方法,主要解决的是城市的点和面问题。同时城市设计主要是通过建筑设计、景观设计来实现的。城市设计的内容也能够细微到桌椅、灯具甚至标志物,但与建筑设计仍有质的区别。城市设计对城市是从整体形象把握,即使具体到任何细小局部时,设计师依然将每个细部作为城市空间体系中的一个部分进行设计,而建筑设计只是关心在特定空间的某一建筑,却很少关心它的邻居,缺乏对城市空间的总体认识和把握。
在城市设计中不但要注重城市的功能分区,交通流线,而且还要注重建筑物的体量、尺度、比例、色彩、造型、材料、空间等。必须强调“城市设计最基本的特征是将不同的物体(包括建筑物)进行联合,使之成为一个有机整体,设计者不仅必须考虑物体本身的设计,而且还要考虑一个物体与其他物体之间的关系”。这就要协调好二者之间的关系,城市设计以城市和建筑群体空间环境作为主要对象,而一个好的城市设计则在于整体环境的和谐、优美,不仅仅是单纯的建筑单体设计。沙里宁在《论城市》中提出城市体形环境设计的三条原则,其中第二条就是“相互协调的原则”。西特在《城市建设艺术》一书中总结中世纪欧洲城市建设艺术中强调的“互协调要素”,并加以发展,指出自然界虽然千变万化,但又是相互协调的,因此,人类建设新城也应该遵守这条原则。在沙里宁的实践中,把建筑设计、户外空间以及园林绿化等融为一体,形成一个完整和谐的整体。
而我们的城市,最缺的就是关系,建筑与环境之间没有关系,建筑物与建筑物之间没有关系。单独看,有些还不错,放在一起就是乱七八糟。我认为这不是单纯建筑的问题,而是城市设计与建筑设计相协调的问题。
三、 结语
中国的许多城市有上千年的历史,积淀着浓厚的历史文化底蕴。然而现今的体制使许多建筑师成为克隆的高手,现代的城市建设已经让人们辨别不出南北方的差异,内陆与沿海的不同,千年的文化被百年的新城整合成一个模板。河北省许多地区的三级甲等医院门诊楼都是按同一份图纸盖出来的,只不过城市不同而已。KPF 的高科技与细腻,拉尔夫的楼梯间遍布大江南北,漂亮是漂亮,但缺少了味道。建筑本来是一种展现个性魅力的艺术创作,但现在成了表现城市共性的主要元素。
城市是一个国家精神文明与物质文明的缩影。在经济全球化大潮中,一个国家能否在激烈的国际竞争中取得优势,关键在于这个国家的大城市是否具有竞争实力。纵观当今世界,竞争不仅是经济力量的竞争,更是文化精神的竞争。一个新兴的经济型城市,如无文化底蕴,至多是一架经济机器,发展动力显然不足。国际上的大都市,巴黎、伦敦、纽约等,之所以能百多年经久不衰,就在于它们都有着深厚的文化积淀,又不断地与时俱进地提高自己的文化品位,引领时代新潮流。因此,我们应该吸收其精华、去其糟粕,切实处理好三者之间的关系,以找回我们遗失在快速城市化浪潮里的文明。
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三、建筑设备论文3000字
写作思路：根据题目要求，以建筑设备作为主题，详细记录哪些必要的设备，最后进行总结。正文：
1、建筑设备工程施工技术
为了让建筑设备工程更好运行和发挥作用，首先就应该把握施工技术要点，做好设备安装。但在日常工作中，一些施工人员的技术水平较低，责任心不强，导致设备安装存在问题与不足，制约建筑设备工程更好运行和发展作用。今后应该转变这种情况，把握施工技术要点，促进建筑设备工程质量提高，也为建筑工程施工效率提高创造便利。
1.1施工准备工作。为促进设备安装顺利完成，首先应该做好准备工作，建筑设备采购之前，采购员应该严格按照图纸要求进行，对永久性使用的设备，严格按照规定制定采购计划，报相关部门审批，然后按照要求采购。
收货时按照要求对设备进行验收，检查设备是否满足施工规范要求，材料是否合格，零配件是否满足要求，采购的设备是否存在问题，具备合格证和质量保证书。如果是成套大型设备，更应该做好设备验收工作，在有监理工程师和业主在场的情况下，经检查无误之后进行验收，确保设备质量，为接下来进行安装和设备运行创造良好条件。
1.2设备安装技术。安装时要把握每个技术要点，保证安装施工质量提高。重视对设备的位置度、严密性和强度进行严格控制，最好进行设备耐压试验，有效保障设备质量。遵循设计规范要求和技术标准安装设备，把握每个技术要点，避免设备安装时出现故障，促进设备安装质量提高。
重视对地脚螺栓安装质量控制，由于其安装工艺复杂，难度较大，质量控制比较困难，并且容易出现倾斜现象，可能导致较大的误差，如果质量控制不到位，容易使得设备出现整体故障，影响其正常运行和工作。
另外，设备摩擦容易使轴承产生大量热量，导致该问题出现的原因是润滑油太少，润滑油洁净度不够，轴承间的缝隙调整不当等。为预防这些问题出现，首先就要合理调整轴承间的缝隙，做好润滑工作，确保润滑油质量，避免设备出现过热现象，确保建筑设备安装工程质量。
1.3试运转技术。建筑设备安装完成之后，要进行试运转，在空载、满载、正常状态下运转，做好相关数据记录工作，掌握设备运行情况。试运转时应该做好设备各项性能和数据指标的记录工作，对设备性能进行全面分析，对存在的缺陷及时改进和完善。这样一来，就能实现对设备性能的有效保障，为建筑施工顺利进行奠定基础。
1.4提高施工人员技术。注重对技术水平高，基础知识扎实的技术人员引进工作，重视提高建筑设备工程施工技术人员专业素质，使他们更好适应各项工作需要。加强对他们的管理和培训，提高施工人员素质，能熟练的操作建筑设备，促进工程建设效益提高。
2、建筑设备工程施工管理
除了把握施工技术要点，确保建筑设备工程更好运行和发挥作用之外，还应该加强施工管理，提高质量和安全管理水平，推动建筑设备工程施工效率提高。
2.1材料管理。采购建筑设备各项材料之前，对供应商基本情况进行调查，做好对各项设备的采购工作，确保材料质量合格，为更好开展工作奠定基础。同时还要对材料进行检验，有效保障材料质量合格，为建筑设备安装和施工顺利进行奠定基础，促进工程建设质量提高。
2.2质量管理。设备工程施工之前，要提高工程质量控制水平，做好图纸设计工作，对各项设备安装进行科学合理安排，制定科学合理的施工方案，做好技术交底工作，明确施工任务和要求，促进建筑设备工程施工顺利进行。
重视施工现场巡视工作，做好各项设备处理工作，及时发现存在的问题与不足，采取有效的措施处理和应对，保障工程质量提高。
2.3现场管理。施工单位要配备工作人员，做好对施工现场的巡视，完善现场检查，加强对施工机械和设备、施工人员的检查，确保设备正常作用发挥。重视设备日常维护工作，及时添加润滑油，促进设备使用性能最佳发挥。
对设备存在的缺陷也要及时处理和应对，避免施工现场因设备质量不合格而出现质量事故，提高建筑设备施工效率。
2.4安全管理。完善施工安全管理规章制度，明确施工人员管理责任，推动安全管理的规范化和制度化进程。完善现场安全巡视工作，及时排除存在的安全隐患，将安全事故消灭在萌芽状态，推动建筑施工顺利进行。建立安全事故应急处理机制，及时处理和应对突发事件，尽量降低安全事故带来的损失。
2.5试验检测。及时对各项设备进行试验检测，全面掌握设备综合性能，对存在故障的设备立即采取措施处理，从而确保设备质量和性能提高，使其更好运行和发挥作用，促进建筑设备工程运行效率提高，更好发挥相应的作用。
3、结束语
要想促进建筑设备工程更好运行和发挥作用，首先就得把握施工技术要点，加强施工管理，确保设备性能良好，促进运行效率提高。另外还要重视设备工程的日常维修和保养，使其处于良好的性能和工作状态，推动建筑工程施工效率提高，保障工程建设质量和效益。
四、谁能给几篇暖通空调类的英文文献，要附带汉语全文翻译的
testing of an air-cycle refrigeration system for road transport
Abstract
The environmental attractions of air-cycle refrigeration are considerable. 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 unit. This unique plant operated satisfactorily, delivering sustainable cooling for refrigerated trailers using a completely natural and safe working fluid. 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 high. 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 units.
Keywords: Air conditioner; Refrigerated transport; Thermodynamic cycle; Air; Centrifuge compressor; Turbine expander COP,
Nomenclature
PR
Compressor or turbine pressure ratio
TA
Heat exchanger side A temperature (K)
TB
Heat exchanger side B temperature (K)
Tinlet
Inlet temperature (K)
Toutlet
Outlet temperature (K)
ηcomp
Compressor isentropic efficiency
ηturb
Turbine isentropic efficiency
ηheat exchanger
Heat exchanger effectiveness
1. Introduction
The current legislative pressure on conventional refrigerants is well known. 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 option. Consequently, as soon as alternative systems, such as non-HFC refrigerants or air-cycle systems are considered, the issue of increased energy consumption arises immediately.
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 transport. 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 transport. NormalAir Garrett developed and commercialised an air-cycle air conditioning pack that was fitted to high speed trains in Germany in the 90s. As part of an European funded programme, a range of applications for air-cycle refrigeration were investigated and several demonstrator plants were constructed. 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 refrigeration.
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 packaging. 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 fluid. 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 manufacture. 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 optimal. In practical terms, this meant using the chassis and panels for an existing SL200 unit along with the standard diesel engine and circulation fans. The turbomachinery used for compression and expansion was adapted from commercial turbochargers.
2. Thermodynamic modelling and design of the demonstrator plant
The thermodynamics of the air-cycle (or the reverse ‘Joule cycle’) are adequately presented in most thermodynamic textbooks and will not be repeated here. For anything other than the smallest flow rates, the most efficient machines available for the necessary compression and expansion processes are turbomachines. Considerations for the selection of turbomachinery for air-cycle refrigeration systems have been presented and discussed by Spence et al. [3].
a typical configuration of an air-cycle system, which is sometimes called the ‘boot-strap’ configuration. 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 compressor. Instead, the work recovered by the turbine during expansion is utilised in the secondary compressor. The two-stage compression also permits intercooling, which enhances the overall efficiency of the compression process. An ‘open system’ where the cold air is ejected directly into the cold space, removing the need for a heat exchanger in the cold space. 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 developed. The compression and expansion processes were modelled using appropriate values of isentropic efficiency, as defined in Eqs.The heat exchange processes were modelled using values of heat exchanger effectiveness as defined in The model also made allowance for heat exchanger pressure drop. 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 outlet. 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 unity. 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 cycle. However,illustrates, there is a broad range of pressure ratio and duty over which the system can be operated with only moderate variation of COP.
The class of turbomachinery suitable for the demonstrator plant required speeds of around 50 000 rev/min. To simplify the mechanical arrangement and avoid the need for a high-speed electric motor, the two-stage compression system shown was adopted. 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 exchangers. 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 plant.
The two-stage system in incorporated an intercooler between the two compression stages. By dispensing with the intercooler and its associated duct work a larger aftercooler could be accommodated with improved effectiveness and reduced pressure loss. Analysis suggested that the improved performance from a larger aftercooler could compensate for the loss of the intercooler.
shows the impact of the recuperator effectiveness on the COP of the plant, which is clearly more significant than that of the other heat exchangers. As well as boosting cycle efficiency, increased recuperator effectiveness also moves the peak COP to a lower overall system pressure ratio. The impact of pressure loss in the recuperator is the same as for the intercooler and aftercooler shown in. The model did not distinguish between pressure losses in different locations; it was only the sum of the pressure losses that was significant. 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 loss.
The specific cooling capacity of the air-cycle increases with system pressure ratio. Consequently, if a higher system pressure ratio was used the required cooling duty could be achieved with a smaller flow rate of air. shows the mass flow rate of air required to deliver 7,5 kW of cooling power for varying system pressure ratios.
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 rotors. and were based on efficiencies of 81 and 85% for compression and expansion, respectively. While such efficiencies are attainable with optimised designs, they would not be realised using compromised turbocharger components. For the design of the demonstrator plant efficiencies of 78 and 80% were assumed to be realistically attainable for compression and expansion.
Lower turbomachinery efficiencies corresponded to higher cycle pressure ratios and flow rates in order to achieve the target cooling duty. The cycle design point was also compromised to help heat exchanger performance. The pressure losses in duct work and heat exchangers increased in proportion with the square of flow velocity. Selecting a higher cycle pressure ratio corresponded to a lower mass flow rate and also increased density at inlet to the aftercooler heat exchanger. 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 exchanger. 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 space. 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 0,41. The system pressure ratio at the design point was 2,14 and the corresponding mass flow rate of air was 0,278 kg/s.
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 reduced. Also, operating at part-load corresponded to lower flow velocities and anticipated improvements in heat exchanger performance. 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 cycle.
3. Prime mover and primary compressor
The existing diesel engine was judged adequate to power the demonstrator plant. 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 controller.From the thermodynamic model, the pressure ratio for the primary compressor was 1,70. The centrifugal compressor required a shaft speed of around 55 000 rev/min. Other alternatives were evaluated for primary compression with the aim of obtaining a suitable device that operated at a lower speed. Other commercially available devices such as Roots blowers and rotary piston blowers were all excluded on the basis of poor efficiency.
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/min. 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 clutch.
4. Cold air unit
The secondary compressor and the expansion turbine were mounted on the same shaft in a free rotating unit. 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 installation. Consequently, both the secondary compressor and the turbine stage were specially chosen and developed to deliver suitable performance.
Most turbochargers use plain oil fed journal bearings, which are low-cost, reliable and provide effective damping of shaft vibrations. However, plain bearings dissipate a substantial amount of shaft power through viscous losses in the oil films. 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 power. Also, the clearances in plain bearings require larger blade tip clearances for both the compressor and the turbine with a consequential efficiency penalty. Given the pressurised inlet to the secondary compressor, the limited thrust capacity of the plain bearing arrangement was also a concern. A CAU utilising high-speed ball bearings, or air bearings, was identified as a preferable arrangement to plain bearings. Benefits would include greatly reduced bearing power losses, reduced turbomachinery tip clearance losses and increased thrust load capacity. However, adequate resources were not available to design a special one-off high speed ball bearing system. Consequently, a standard turbocharger plain bearing system was used.
The secondary compressor stage was a standard turbocharger compressor selected for a pressure ratio of 1,264. Secondary compressor and turbine selection were linked because of the requirement to balance power and match the speed. Since most commercial turbines are sized for high temperature (and consequently low density) air at inlet, a special turbine stage was developed for the application. Cost considerations precluded the manufacture of a custom turbine rotor, so a commercially available rotor was used. 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 velocity. The exhaust diffuser exited into a specially designed exhaust silencer.The performance of the turbine stage was measured before the unit was incorporated into the complete demonstrator plant. The peak efficiency of the turbine was established at 81%.
5. Heat exchangers
Due 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 exchangers. 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 cost. 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 plant.
Several different tube and fin heat exchangers were tested and used to validate a computational model. Once validated, the model was used to assess a wide range of possible heat exchanger configurations that could fit within the Thermo King SL200 chassis. 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 challenging. 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 chassis. The final selection comprised two separate aftercooler units, while the single recuperator was a large, triple pass unit. Based on laboratory tests and the heat exchanger model, the anticipated effectiveness of both the recuperator and aftercooler units was 80%.
6. Instrumentation
A range of conventional pressure and temperature instrumentation was installed on the air-cycle demonstrator plant. Air temperature and pressure was logged at inlet and outlet from each heat exchanger, compressor and the turbine. 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 counter. No air flow measurement was included on the demonstrator plant. Instead, the air flow rate was deduced from the previously obtained turbine performance map using the measurements of turbine pressure ratio and rotational speed.
7. System testing
During some preliminary tests a heat load was applied and the functionality of the demonstrator plant was established. 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 units. The calorimeter was ideally suited for accurately measuring the refrigeration capacity of the air-cycle demonstrator plant. 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 standards.
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 assessed. 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 point. The demonstrator plant operated satisfactorily, allowing sufficient measurements to be obtained at each of the three operating conditions. The recorded performance is summarised .
In total, the unit operated for approximately 3 h during the course of the various tests. 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 stage. 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 testing.
8. Discussion of measured performance
From the calorimeter performance measurements, the primary objective of the project had been achieved. 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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