毕业设计我帮你

压缩机 外文翻译
单螺杆式压缩机具有单一开槽转子，用旋转恒星齿密封叶片来限制气体的口袋，因为它们沿转子槽移动。再次，各种几何形状是可能的，但当前正在生产压缩机具有六个凹槽和分与11齿的转子。正常排列是2分，一个在转子的两侧。因此，每个转子槽被使用两次，在主转子每转一圈，并在转子的气体压力载荷被平衡了，导致轻得多轴承负荷比相应的双螺杆设计。星星由转子驱动，并且因为没有转矩被传递的网眼的润滑要求更轻也。油冷却和密封通常，油电路是类似的双螺杆的。
螺杆压缩机没有余隙容积，并且VE由于再膨胀作为在一个活塞机器没有任何损失。容积损失导致主要是从经由内置间隙的制冷剂回流至吸气泄漏。油用于密封，但油，其含有溶解的制冷剂，降低了泄漏通过制冷剂的释放和通过加热进入的气体VE两者。的VE减小随着压力比，但小于与一些类型的活塞。泄漏损耗是尖端速度的函数，从而使较小的机器需要操作更高的速度以维持效率。与同步电动机驱动器，此设置的大小的下实际限制
在所有的螺杆式压缩机，所述气体体积将减少到入口体积由出口端口被发现的时间的预先设定的比例，这被称为内置体积比。此时内螺钉气体被打开到冷凝器压力，气体将流动向内或向外通过排出口，如果压力不相等。
螺杆式压缩机的吸收功率将是在其最佳只有当工作压力比对应于内置的体积比。的上方和下方的压缩损失可以被可视化为一个指示图上附加区域，如图4.21。这导致具有强峰定义一个IE特性，如图4.18。过度压缩的气体导致效率的损失峰的左侧，而在右边有下压缩用气体回流入时的排放口不被覆盖的压缩腔。改变排出口的大小的变化的峰值的位置，并且这是由两条曲线如图4.18所示。的螺杆式压缩机，应选择具有体积比西装能够为应用程序。泄漏也有助于对效率的损失，但摩擦作用是相当小的。
螺杆压缩机的容量减少影响滑块覆盖了桶壁的一部分,允许气体通过回吸,所以不同的工作行程(图4.22)。通常是滑动的一部分桶的大小来调整卸货港同时,这样体积比至少是大约维持在部分负载。许多设计变化和控制方法存在。单螺杆类型通常会有两个滑动阀、升降阀有时被用来代替幻灯片。减少10%的最大容量是平常.
油分离、冷却和过滤螺杆压缩机增加原本简单的机器的复杂性。液体喷射有时用来代替外部油冷却器。一些商业螺杆压缩机有油装卸电路组装。在图4.23的吸入气体进入左边的吸入口连接,通过电动机,通过压缩机,进入右边的多级分离器和基因表达放电的连接。
4.12卷轴压缩机
尽管滚动机制已经知道多年来,一直专利于1905年在法国,直到上个世纪的后半部分它出现在商用压缩机中的第一。生产工艺,这时开发创新足量地使精密螺旋形式在经济上。
卷轴压缩机容积式压缩制冷剂的机器有两个配置螺旋形滚动成员如图4.24所示。一个滚动仍然固定在第二滚动在轨道上移动。注意,移动滚动不但是轨道旋转圆周运动。一般两到三个轨道,或曲轴革命,都必须完成压缩循环。
滚动与螺丝某些共同的特征。这两种类型都内置于体积比，因此，在滚动显示的IE曲线在形状上的那些螺钉的相似（图4.18）。没有余隙容积，因此没有再膨胀损失。然而，有一个在压缩腔的密封非常重要的区别。螺杆依靠转子和机壳，而滚动可以接触密封，即滚动接触对方的口袋边界建设之间的间隙。这是可能的，因为轨道运动引起低得多的速度比旋转运动，并且还对侧翼和所述卷轴的顶端负载可被控制。此外，还有在所述卷轴和吸入的中心的排放口区域之间没有直接的路径。这样做的结果是非常低的泄漏和热传递损失，得到更好的VE比大多数其他类型（图4.16）的特性。这使得涡旋件，以有效地工作在小得多的位移比螺丝（图4.2），与上大小限制由制造的经济有效地确定。
几乎所有的密封类型的生产卷轴和一个典型的完全外形如图4.25所示。这些压缩机有优势类似尺寸的活塞密封的空调应用和鼓励投资生产设施,构建全球数百万。
平面体积曲线使滚动在极端条件下提供更多的冷却和加热能力,压缩过程更平滑、更加安静,而且有许多运动部件少,保证非常高的可靠性。此外滚动耐故障条件如液体洪水,符合机制可以提供卸载开始和极压保护(Elson et al .,1991)。
虽然无油注入到压缩过程是需要的，轴承和推力面的润滑是非常重要的。油可供给到利用通过沿轴的长度的偏移钻井产生的离心力的上驱动轴承等的表面上。采用无级调速变频驱动能力控制是可能的许多卷轴。最近采用间歇，频繁滚动分离的方法已经出台（现代汽车，2002年）。当卷轴轴向地隔开的容量为零。电机继续以正常速度而是以非常低的功率和气体从高压侧倒流防止通过排出阀运行。当卷轴都汇集正常泵送恢复。固定涡旋件的轴向运动是由一个液压致动活塞供电，响应于来自控制器的脉冲宽度调制信号。总循环时间通常为20秒，该循环时间内所加载的周期的持续时间是无限可变的。因为循环时间相对较短，该系统的热惯量有阻尼的波动，使得效果是非常相似的连续操作以降低容量的效果。
空调滚动“起飞”，促使引入许多变体，其中最重要的是制冷或低温的版本。作为与螺杆，使用更小的排出口使压缩机适用于较低温度的应用中的更高的压力比进行优化。通过引入一个排出阀，类似于那些在往复式压缩机采用，在压缩下的影响降到最低限度。液体注入用于冷却在必要时和节能循环可以用于提高容量和效率（见第3章）。这些发展使制冷涡旋与活塞类型在各种各样的商业应用竞争。
一个空调的大小上上的大小限制为单个空调涡旋已大大扩大了与引进的双涡旋具有滚动组安装在水平轴（皮雷纳，2007）的每端。关于产能减少50％是由空转的滚动集之一来实现。
4.13滑动和旋转叶片式压缩机
偏心转子和滑动叶片之间的体积将与角位置而变化，以提供正位移压缩机（图4.26）的一种形式。较大的车型有八个或多个刀片，而且不需要入口或出口阀。叶片被保持在与外壳由离心力紧密接触，并密封是通过注射沿叶片的长度的润滑油的改进。旋片机有没有间隙体积，但它们在应用由设立在叶片的尖端的推力的应力是有限的。虽然它们已被用于在低排放压力诸如化合物周期的第一阶段中，他们不再广泛适用于制冷剂压缩。
滑片或滚动活塞式压缩机有一个或两个刀片，不转动，而是由弹簧对偏心，旋转滚轮举行。这些压缩机需要排气阀。这种类型的已开发大量用于家用电器，包装空调器及类似的应用，高达15千瓦的冷却占空比（见图4.27）
4.14动态压缩器
动力压缩机的能量传递到气体通过速度或离心力的作用，然后将其转换为压力能。最常见的类型是离心式压缩机。吸入气体进入轴向成转子，其具有弯曲叶片的眼睛，并且从叶片圆周抛出切向。
通过这样的机器给气流通的能量依赖于气体的速度和密度。由于密度已经是固定的工作条件下，离心式压缩机的设计性能将由转子尖端速度决定。由于所使用的气体的密度低，尖端速度高达300米/秒是常见的。在2900转/分的电动机的速度，一单级机将需要在直径的叶轮2微米。为了减少这一个更易于管理的大小，驱动器从标准速电机齿轮或在供频率变化来获得更高的电机速度。驱动电机是一体的压缩机组件，可以是开放或密闭式的。在单级离心压缩机的空调的职责，转子速度通常为约10000转/分。
气体可被压缩在两个或更多个阶段。叶轮是在同一轴上，从一个阶段直接传递到下一个赋予与气体紧凑串联布置。压缩的步骤不是很大，并且如果两阶段的情况下，该气体可以通过从第一到第二，没有任何间冷却。
离心机可以用于工业用氨和其它致冷剂被建成，并且这些可以具有多达七个压缩级。随着高尖端速度在使用中，它是不实际建立一个小机器，最小的可用离心压缩机制冷职责的能力为约260千瓦。半封闭压缩机是由7000千瓦和开放带动机器达到21000千瓦。有没有部件需要润滑，除了主轴承。其结果是，该机器可以运行几乎无油。
这种尺寸的系统需要大直径的制冷剂的吸入和排出管道连接的完整的系统的组件。其结果，与除了大规模的工业厂房，它们几乎总是建立作为液体冷却，水冷包与冷凝器和蒸发器的完整的工厂制造的包的一部分。为离心式包装冷水机组的主要制冷剂是R134a的。
该离心分离机的泵送特性不同于容积式压缩机，因为，在过高的排放压力，气体可以向后溜过转子。这一特性使得该离心压缩机敏感的冷凝条件，给予更高的职责，性能更好的系数，如果压头下降，而严重缺陷的性能，如果压头上升。这也将随容量减少叶片的角度。过大的压力会导致逆流状态，之后是一个第二的一小部分通过后的升压流作为头压力下降。蒸气浪涌，交替正向和反向气流，在叶轮投掷额外应力和驱动马达。这样运行的条件是，以尽可能避免的，通过用适当的低水头压力与由良好维护的聚光系统的设计。评级曲线表明失速或喘振极限..
由于离心机太大频繁停止和重新启动控制，某种形式的容量减少的必须是内置的。一般的方法是进行节流或偏转吸入气体流入叶轮。大多数的模型也能够减少泵送容量降至总流量的10％-15％。低成本逆变器的可用性已导致使用变速驱动器，提供增加离心式压缩机的效率。然而，这并不能完全取代的需要可变进口导叶，因为从低流量系统头的要求（Brasz，2007）所产生的早期（低头）激增。离心式压缩机具有可变几何形状的一个例子示于图4.28。最近进入压缩机领域是变速离心式压缩机用直流驱动电机和磁轴承图4.29。这开辟了无油系统的可能性。
喷气压缩机是一个动态的压缩机在大小尺度的另一端。目前它是未发生的研究工作和商业引入受试者。其作为提高吸收循环的一种方法的使用已被成功地测试（埃姆斯，2005）。
五章 油制冷剂电路
5.1介绍
油的主要目的是用于压缩机的润滑;还密封和冷却喷油类型。由压缩机制造商指定的机油应尽可能使用。各种矿物油和合成润滑剂类型可用图5.1给出了一些提示，以他们的西服的能力与制冷剂的类型。空白表明油的类型，一般不适合。有必要强调的是，示出的油的类型的每一个代表一个家庭的产品，它可以被混合以得到所需的粘度，并且有必要使用在家庭内的适当等级或产品。例如，多元醇酯（POE）油被示为适合的烃，但是当用与用于氢氟烃相比，更高的粘度等级一般将需要的烃。此外，特定的添加剂包，以提高润滑性或充当抑制剂可以存在于品牌的产品，这就是为什么在压缩机制造商应与问候改变指定油为特定压缩机协商。
5.2需求和特点
油的性质，必须适当考虑其行为中的系统，即从系统油返回到压缩机中，而油在热传递过程中的蒸发器和冷凝器的作用。压缩机设计者的任务的一部分，是为了确保润滑剂健身两个压缩机和系统。
在工作环境中的润滑剂总是的油和制冷剂，因此，它的组成和性质通过是压力和温度的依赖性溶解度特性支配的混合物。在CFC，HCFC和氨制冷剂/矿物油的组合是由多年的经验支持;它们的性能是公知的。压缩机设计者已经利用粘度和含氯制冷剂，以在运动部件的设计良好的效果的优良边界润滑（润滑性）性质的组合。与HFC制冷剂的出现的是需要从矿物油到合成油移动，以便确保混溶性的制冷剂，并从系统因而足够的回油。 POE油被选择用于大多数应用中，根据它们的性质，成本和可用性。公共所有企业是从有机酸和醇，其结合以产生酯和水制成。酯的制剂是由原始酸结构决定的。正如它的名字所暗示的，是POE是由酸的混合物衍生的酯的混合物。
在制冷剂回路，其物理相互作用与制冷剂本身润滑油的行为是主导因素，特别是在一般的电路和蒸发器的设计。应当注意的是，氨的大部分润滑剂中的溶解度是非常低的。
制冷剂在油中的溶解度的程度是理想的，因为在蒸发器中的混合物的粘度降低，允许它变得更多的移动，这有助于输送回到压缩机。对压缩机运转的最重要的属性是用于轴承的润滑的溶液的粘度。与蒸发器的温度低侧压力变化的制冷剂的浓度也改变，这反过来又影响粘度。对卤烃制冷剂的典型行为如图5.2所示。
在1巴对应于蒸发发言权-40℃的低压少量是制冷剂被溶解，这对粘度的影响可以忽略。与在较高的蒸发压力，比方说6巴，相当于10℃的油吸收的10％的制冷剂，有效地降低了粘度的一半，该基油和轴承承载能力降低。这些影响进行了研究用的粘度/温度/压力图的辅助下，其中的格式示于图5.3。此图仅仅是为了说明的特征的一般形式，并基于制冷剂R134a和POE油。具体的数据可以在ASHRAE制冷手册中找到。
当液体润滑剂和液体制冷剂被混合在一起，并使其沉降，均匀的混合物可以形成。在这种情况下，在一对被认为是在当前的温度和压力下互溶。可替代地，两个单独的阶段可形成;其中之一是富油溶液，另一个是制冷剂的富溶液。在大多数情况下，较重的制冷剂的富溶液是在底部。这可能会导致问题的系统中压缩机是在寒冷的地方和制冷剂停机期间凝结在曲轴箱。在启动时，油泵将趋向于绘制一个非常低粘度制冷剂 - 富混合物。曲轴箱加热器和抽气周期用于避免这个问题。这是不与氨的情况下，因为它通常不与润滑剂混合除少量与油趋于积聚在蒸发器的底部，在那里它可以被排空。
润滑剂所需性质可以总结如下。许多的特征是由制冷剂的影响，所以该油属性不能单独考虑。
1.足够的润滑粘度的轴承的温度和压力，和足够的润滑性的滑动接触。
2.稳定性，从而使化学反应或分解不会发生在的条件会遇到。通常情况下，最高温度和压力是在压缩的排出。耐氧化度用FL灰点测试。
3.润滑剂必须是湿气和污染自由，只要是可能的。
4. 润滑剂必须与系统中所用的材料相容。特定点是柔性非金属的橡胶和塑料部件，例如密封。铜，不能使用与氨。
5. 在低温侧的溶液的粘度应该是足够的回油足够低。
6. 固体不应该被沉淀。矿物油可以有时析出蜡在低温下;这是确定了絮凝点测试。
7.  高电阻是必要的封闭式电动机。
8.  发泡特性，必须加以考虑。
9. 可用性以可接受的成本是必要的。
通过在上述的一些点评论方式，应当注意的是，在实际工作环境中的润滑特性只能由实际经验和/或寿命试验来证明。这要归功于压缩机制造商和系统安装的是，转换为HFC制冷剂和POE油一直非常顺利，无故障的过程中，工程工作。化学稳定性必须是足够的水分和空气的存在，虽然其目的是总是排除从系统中这些污染物。痕迹仍然存在于实践，这是处理如下。过度发泡是不理想的，当它是通过快速释放的制冷剂时，压缩机起动引起的，曲轴箱压力降低，因为它往往会产生的油损失到系统。正常运转期间一些泡沫可以协助油分布在压缩机和降低声级。
5.3水分和空气污染
在过去与水分的主要问题是形成冰在关键区域，如膨胀阀，但它也造成腐蚀和损坏电机绕组。润滑油在确定一制冷系统的污染物的效果方面发挥重要作用。因为它是当它进入所述压缩机（不像汽车发动机被迅速通过燃料，水，碳和大气粉尘污染的）的油应该保持为清洁。因此，压缩机油的条件是该系统的物理和化学清洁的直接指示。润滑油应保持在密闭容器中，以排除大气中的水分。油氨排水系统不会再次使用，除非可以适当过滤，并保持干燥。过热或在密闭或半密闭型压缩机马达的绕组的电气故障会产生污染物，包括氢卤酸。护目镜和橡胶手套应该处理这些犯罪嫌疑人时，油被磨损。如果证明是酸，油必须除去并小心处置，并且系统彻底清理。
水分发生反应POE润滑剂和这引起有机酸。这些比卤酸弱得多，这需要考虑到当酸度水平使用酸检测试剂盒进行测定。然而这首诗的任何故障是不可取的，建议水分含量保持低于50ppm的POE系统。因为公共所有企业具有高亲和力的水分，必须保持油于大气的暴露绝对最小值。湿气无法通过真空程序完全去除，因此适当的干燥过滤器总是建议。图5.4示出了安装和调试的过程中的典型的预期含水量方差。一种压缩机，它是预充电用的POE油被连接的第一到工厂组装单元，例如压缩机包，安装在现场观察正确疏散和密封程序，并最终与运行中的水分含量随后通过降低朝向50ppm的干燥。
适当的撤离将删除含气量到可接受的最低限度。氧化等的高温化学反应更可能导致在湿气的存在损害。
当系统从CFC和HCFC转换到HFC制冷剂中，油也通常需要改变到POE。原始设备制造商的建议前应尝试这种转换的追捧。
5.4油分离器
在一个往复机械的压缩行程中，气体变得更热，部分在气缸壁上的油将通过用放电气体。会发生与所有润滑压缩机的种类一些油结转，并在小型独立系统，很快发现自己的路径返回到压缩机。一个长的空闲时间之后启动可导致大量的油结转一小段时间因发泡。与较大的更复杂的系统与远程蒸发器的油最好是适合的油分离器中的排出管线，以减少结转到系统（见图5.5）。
 
The single screw compressor has a single grooved rotor, with rotating star tooth seal vanes to confine the pockets of gas as they move along the rotor flutes. Once again, various geometries are possible, but compressors currently being manufactured have a rotor with six flutes and stars with eleven teeth. The normal arrangement is two stars, one on either side of the rotor. Each rotor flute is thus used twice in each revolution of the main rotor, and the gas pressure loading on the rotor is balanced out, resulting in much lighter bearing loads than for the corresponding twin screw design. The stars are driven by the rotor, and because no torque is transmitted the lubrication requirements in the mesh are lighter also. Oil cooling and sealing is usual and the oil circuit is similar to that of the twin screw.
Screw compressors have no clearance volume, and there is no loss of VE due to re-expansion as in a piston machine. Volumetric losses result mainly from leakage of refrigerant back to the suction via in-built clearances. Oil is used for sealing, but leakage of oil, which contains dissolved refrigerant, reduces VE both by release of refrigerant and by heating the incoming gas. The VE decreases with increasing pressure ratio, but less than with some piston types (Figure 4.16). Leakage losses are a function of tip speed, so that smaller machines need to operate a higher speed to maintain efficiency. With synchronous motor drives, this sets a lower practical limit on the size (Figure 4.2).
In all screw compressors, the gas volume will have been reduced to a pre-set proportion of the inlet volume by the time the outlet port is uncovered, and this is termed the built-in volume ratio. At this point the gas within the screws is opened to condenser pressure, and gas will flow inwards or outwards through the discharge port if the pressures are not equal.
The absorbed power of the screw compressor will be at its optimum only when the working pressure ratio corresponds to the built-in volume ratio. The over and under compression losses can be visualized as additional areas on an indicatory diagram as in Figure 4.21. This results in an IE characteristic having a strongly defined peak, as in Figure 4.18. To the left of the peak over-compression of the gas results in loss of efficiency, whilst to the right there is under-compression with back flow of gas into the compression pocket when the discharge port is uncovered. Changing the size of the discharge port changes the position of the peak, and this is illustrated by the two curves in Figure 4.18. A screw compressor should be chosen to have a volume ratio suit-able for the application. Leakage also contributes towards efficiency loss, but friction effects are quite small.
Capacity reduction of the screw compressor is effected by a sliding block covering part of the barrel wall, which permits gas to pass back to the suction, so varying the working stroke (Figure 4.22 ). It is usual for the sliding part of the barrel to adjust the size of the discharge port at the same time, so that the volume ratio is at least approximately maintained at part load. Many design variations and control methods exist. The single screw type will generally have two sliding valves; lifting valves are sometimes used instead of slides. Reduction down to 10% of maximum capacity is usual.
The oil separation, cooling and filtering for a screw compressor add to the complexity of an otherwise simple machine. Liquid injection is sometimes used instead of an external oil cooler. Some commercial screw compressors have the oil-handling circuit built into the assembly. In Figure 4.23 the suction gas enters at the suction connection on the left, passes over the motor, through the compressor, into the multi-stage separator on the right and finally back to the discharge connection.
4.12 SCROLL COMPRESSORS
Although the scroll mechanism has been known for many years, having been patented in France in 1905, it was not until the latter part of the last century that it first appeared in commercially available compressors. Manufacturing technology had by this time developed sufficiently to enable the precision spiral forms to be made economically.
Scroll compressors are positive displacement machines that compress refrigerants with two terfitting spiral-shaped scroll members as shown in Figure 4.24 .One scroll remains fixed whilst the second scroll moves in orbit inside it. Note that the moving scroll does not rotate but orbits with a circular motion. Typically two to three orbits, or crankshaft revolutions, are required to complete the compression cycle.
The scroll has certain common features with the screw. Both types have a built in volume ratio and therefore the scroll exhibits IE curves similar in shape to those of the screw (Figure 4.18 ). There is no clearance volume and hence no re-expansion loss. However, there is a very important difference in the sealing of the compression pockets. The screw relies on clearance between rotors and casing whereas the scroll can be built with contacting seals, i.e. the scrolls touch each other at the pocket boundaries. This is possible because the orbiting motion gives rise to much lower velocities than rotating motion, and also the load on the flanks and tips of the scrolls can be controlled. Additionally there is no direct path between the discharge port area at the centre of the scrolls and the suction. The result of this is very low leakage and heat transfer losses, giving better VE characteristic than most other types (Figure 4.16). This enables the scroll to function efficiently in much smaller displacements than the screw (Figure 4.2), with theupper size limit being effectively determined by the economics of manufacture.
Almost all production scrolls are of the hermetic type and a typical con-figuration is shown in Figure 4.25. These compressors have advantages over similar sized piston hermetics in air-conditioning applications and this has encouraged investment in production facilities, building millions worldwide.
The flat volumetric curve enables the scroll to deliver more cooling and heating capacity at extreme conditions, the compression process is smoother and quieter, and there are many fewer moving parts, ensuring very high reliability. Additionally the scroll has excellent resistance to fault conditions such as liquid flood back, and compliance mechanisms can deliver unloaded starting and extreme pressure protection (Elson et al., 1991).
Whilst no oil injection into the compression process is needed, bearing and thrust surface lubrication is vital. Oil can be fed to the upper drive bearings and other surfaces using the centrifugal forces generated by an offset drilling along the length of the shaft. Capacity control using variable speed inverter drive is possible for many scrolls. More recently a method using intermittent and frequent scroll separation has been introduced (Hundy, 2002). When the scrolls are separated axially the capacity is zero. The motor continues to run at normal speed but with very low power and back flow of gas from the high-pressure side is prevented by a discharge valve. When the scrolls are brought together normal pumping is resumed. The axial movement of the fixed scroll is powered by a hydraulically actuated piston, in response to a pulse width modulated signal from a controller. The total cycle time is typically 20 seconds, and the duration of the loaded period within that cycle time is infinitely variable. Because the cycle time is relativety short, the thermal inertia of the system has the effect of damping the fluctuations so that the effect is very similar to continuous operation at reduced capacity.
The ‘take-off ’ of air-conditioning scrolls has prompted the introduction of many variants, the most important of which is the refrigeration or low-temperature version. As with the screw, use of a smaller discharge port enables the compressor to be optimized for the higher pressure ratios applicable to lower temperature applications. By introducing a discharge valve, similar to those employed in reciprocating compressors, the effects of under compression can be minimized. Liquid injection is used for cooling where necessary, and the economizer cycle can be used to boost capacity and efficiency (see Chapter 3).These developments have enabled the refrigeration scroll to compete with piston types in a wide variety of commercial applications.
The upper size limit for a single air-conditioning scroll has been considerably extended with the introduction of a dual scroll that has a scroll set mounted on each end of a horizontal shaft (Pirenne, 2007). About 50% capacity reduction is achieved by idling one of the scroll sets.
4.13 SLIDING AND ROTARY VANE COMPRESSORS
The volumes between an eccentric rotor and sliding vanes will vary with angular position, to provide a form of positive displacement compressor (Figure 4.26). Larger models have eight or more blades and do not require inlet or outlet valves. The blades are held in close contact with the outer shell by centrifugal force, and sealing is improved by the injection of lubricating oil along the length of the blades. Rotary vane machines have no clearance volume, but they are limited in application by the stresses set up by the thrust on the tips of the blades. Whilst they have been used at low discharge pressures such as the first stage of a compound cycle, they are no longer widely applied in refrigerant compression.
Sliding vane or rolling piston compressors have one or two blades, which do not rotate, but are held by springs against an eccentric, rotating roller. These compressors require discharge valves. This type has been developed extensively for domestic appliances, packaged air-conditioners and similar applications, up to a cooling duty of 15 kW (see Figure 4.27).
4.14 DYNAMIC COMPRESSORS
Dynamic compressors impart energy to the gas by velocity or centrifugal force and then convert this to pressure energy. The most common type is the centrifugal compressor. Suction gas enters axially into the eye of a rotor which has curved blades, and is thrown out tangentially from the blade circumference.
The energy given to gas passing through such a machine depends on the velocity and density of the gas. Since the density is already fixed by the working conditions, the design performance of a centrifugal compressor will be decided by the rotor tip speed. Owing to the low density of gases used, tip speeds up to 300 m/s are common. At an electric motor speed of 2900 rev/min, a single-stage machine would require an impeller 2 m in diameter. To reduce this to a more manageable size, drives are geared up from standard-speed motors or the sup- ply frequency is changed to get higher motor speeds. The drive motor is integral with the compressor assembly and may be of the open or hermetic type. On single-stage centrifugal compressors for air-conditioning duty, rotor speeds are usually about 10 000 rev/min.
Gas may be compressed in two or more stages. The impellers are on the same shaft, giving a compact tandem arrangement with the gas from one stage passing directly to the next. The steps of compression are not very great and, if two-stage is used, the gas may pass from the first to the second without any inter-cooling.
Centrifugal machines can be built for industrial use with ammonia and other refrigerants, and these may have up to seven compression stages. With the high tip speeds in use, it is not practical to build a small machine, and the smallest available centrifugal compressor for refrigeration duty has a capacity of some 260 kW. Semi-hermetic compressors are made up to 7000 kW and open drive machines up to 21 000 kW capacity. There are no components which require lubrication, with the exception of the main bearings. As a result, the machine can run almost oil free.
Systems of this size require large-diameter refrigerant suction and discharge pipes to connect the components of the complete system. As a result, and apart from large-scale industrial plants, they are almost invariably built up as liquid-cooling, water-cooled packages with the condenser and evaporator complete as part of a factory-built package. The main refrigerant for packaged water chillers of the centrifugal type is R134a.
The pumping characteristic of the centrifugal machine differs from the positive displacement compressor since, at excessively high discharge pressure, gas can slip backwards past the rotor. This characteristic makes the centrifugal compressor sensitive to the condensing condition, giving higher duty and a better coefficient of performance if the head pressure drops, while heavily penalizing performance if the head pressure rises. This will vary also with the angle of the capacity reduction blades. Excessive pressure will result in a reverse flow condition, which is followed a fraction of a second later by a boosted flow as the head pressure falls. The vapour surges, with alternate forward and reverse gas flow, throwing extra stress on the impeller and drive motor. Such running conditions are to be avoided as far as possible, by designing with an adequately low head pressure and by good maintenance of the condenser system. Rating curves indicate the stall or surge limit..
Since centrifugal machines are too big to control by frequent stopping and restarting, some form of capacity reduction must be inbuilt. The general method is to throttle or deflect the flow of suction gas into the impeller. With most models it is possible to reduce the pumping capacity down to 10–15% of full flow. The availability of low-cost inverters has led to the use of variable speed drive which offers increased centrifugal compressor efficiency. However this cannot totally replace the need for variable inlet guide vanes because of early (low head) surge arising from low flow system head requirements (Brasz, 2007). An example of a centrifugal compressor with variable geometry is shown in Figure 4.28. A recent entry to the field of compressors is a variable speed centrifugal compressor with a DC drive motor and magnetic bearings Figure 4.29. This opens up the possibility of oil free systems.
The jet compressor is a dynamic compressor at the other end of the size scale. At present it is a subject of research work and commercial introduction has not occurred. Its use as a way of enhancing the absorption cycle has been successfully tested (Eames, 2005).
5.1 INTRODUCTION
Chapter | Five Oil in Refrigerant Circuits
The primary purpose of the oil is for compressor lubrication; also sealing and cooling for oil-injected types. The oil specified by the compressor manufacturer should be used whenever possible. Various mineral and synthetic types of lubricant are available and Figure 5.1 gives some indication as to their suit-ability with refrigerant type. A blank indicates that the oil type is generally unsuited. It is necessary to emphasize that the oil types shown each represent a family of products, which can be blended to give the required viscosity, and it is necessary to use an appropriate grade or product within the family. For example,              polyolester (POE) oils are shown as suitable for hydrocarbons, but a higher viscosity grade will generally be required for hydrocarbons when compared with that used for HFCs. Moreover, specific additive packages to enhance lubricity or to act as inhibitors may be present in branded products, and this is why the compressor manufacturer should be consulted with regards to changing the specified oil for a particular compressor.
5.2 REQUIREMENTS AND CHARACTERISTICS
The properties of the oil must take proper account of its behaviour in the sys-tem, namely oil return from the system to the compressor, and the effect of oil on the heat transfer process in the evaporator and condenser. Part of the compressor designer’s task is to ensure lubricant fitness for both the compressor and the system.
In the working environment the lubricant is always a mixture of oil and refrigerant and therefore its composition and properties are governed by solubility characteristics that are pressure and temperature dependent. The CFC, HCFC and ammonia refrigerant/mineral oil combinations are backed by many years of experience; their properties are well known. Compressor designers have utilized the combination of viscosity and the excellent boundary lubrication (lubricity) properties of the chlorine-containing refrigerants to good effect in the design of the moving parts. With the advent of HFC refrigerants came the need to move from mineral oils to synthetic oils in order to ensure miscibility with the refrigerant and hence adequate oil return from the system. POE oils were chosen for most applications, based on their properties, cost and availability. POEs are made from organic acids and alcohols, which combine to produce esters and water. The formulation of the ester is determined by the original acid structure. As its name implies, a POE is a mixture of esters derived from a mixture of acids.
The behaviour of lubricating oil in a refrigerant circuit and its physical interaction with the refrigerant itself are dominant factors in the design of circuits in general and evaporators in particular. It should be noted that the solubility of ammonia in most lubricants is very low.
A degree of solubility of refrigerant in oil is desirable because viscosity of the mixture in the evaporator is reduced, allowing it to become more mobile, which aids transport back to the compressor. The most important property for compressor operation is the viscosity of the solution for bearing lubrication. As the low-side pressure changes with evaporator temperature the refrigerant concentration changes also and this in turn affects viscosity. A typical behaviour for halocarbon refrigerants is shown in Figure 5.2.
At low pressure of 1 bar corresponding to evaporation at say -40°C a small amount is refrigerant is dissolved and this has negligible effect on viscosity. With a higher evaporation pressure at, say 6 bar, corresponding to 10°C the oil absorbs 10% refrigerant which effectively reduces the viscosity to half that of the base oil and the bearing load carrying capacity is reduced. These effects are studied with the aid of viscosity/temperature/pressure diagrams, the format of which is shown in Figure 5.3. This diagram is just intended to illustrate the general form of the characteristic and is based on refrigerant R134a and POE oil. Specific data may be found in the ASHRAE Refrigeration Handboo.
When liquid lubricants and liquid refrigerants are mixed together and allowed to settle out, a homogeneous mixture may be formed. In this case the pair are said to be miscible at the prevailing temperature and pressure. Alternatively, two separate phases may form; one of which is an oil-rich solution, the other being a refrigerant-rich solution. In most cases the heavier refrigerant-rich solution is at the bottom. This can cause problems with systems where the compressor is in a cold location and refrigerant condenses in the crankcase during shutdown. On start up, the oil pump will tend to draw a very low-viscosity refrigerant- rich mixture. Crankcase heaters and pump down cycles are used to avoid this problem. This is not the case with ammonia as it generally does not mix with lubricants except in small amounts and the oil tends to accumulate at the bottom of evaporators where it can be drained.
Desired properties for lubricant may be summarized as follows. Many of the characteristics are influenced by the refrigerant and so the oil property cannot be considered in isolation.
1. Adequate lubrication viscosity at the temperatures and pressures in the bearings, and adequate lubricity for sliding contacts.
2. Stability, so that chemical reactions or decomposition do not occur at the conditions to be encountered. Normally the highest temperature and pressure is at the discharge of compression. Resistance to oxidation is measured with a fl ash point test.
3. The lubricant must be moisture and contamination free, as far as impossible.
4. The lubricant must be compatible with the materials used in the system. Particular points are flexible non-metallic rubber and plastic components such as seals. Copper cannot be used with ammonia.
5. The viscosity of the solution on the low-temperature side should be low enough for adequate oil return.
6. Solids should not be precipitated. Mineral oils can sometimes precipitate waxes at low temperature; this is identified with a floc point test.
7. High electrical resistance is necessary for enclosed motors.
8. Foaming characteristics must be considered.
9. Availability at an acceptable cost is essential.
By way of comment on some of the above points, it should be noted that the lubrication characteristics in the actual working environment can only be proven by actual experience and/or life tests. It is a tribute to the engineering efforts of compressor builders and system installers that the changeover to HFC refrigerants and POE oils had been a very smooth and trouble-free process. Chemical stability has to be adequate in the presence of moisture and air although the aim is always to exclude these contaminants from the system. Traces are nevertheless present in practice and this is dealt with below. Excessive foaming is undesirable when it is caused by rapid refrigerant release when the compressor is started and the crankcase pressure reduces because it tends to give rise to loss of oil to the system. Some foaming during normal running can assist oil distribution in the compressor and reduce sound level.
5.3MOISTURE AND AIR CONTAMINATION
In the past the main problem with moisture was ice formation in critical areas such as expansion valves, but it also caused corrosion and damage to motor windings. Lubricants play an important role in determining the effect of contaminants on a refrigeration system. The oil should remain as clean as it is when it entered the compressor (unlike that of the automobile engine which is quickly contaminated by fuel, water, carbon and atmospheric dust). The condition of the compressor oil is therefore a direct indication of the physical and chemical cleanliness of the system. Lubricating oil should be kept in tightly sealed containers to exclude atmospheric moisture. Oil drained from ammonia systems is not used again unless it can be properly filtered and kept dry. Overheating or an electrical fault in the winding of a hermetic or semi-hermetic compressor motor will produce contaminants, including the halogen acids. Eye goggles and rubber gloves should be worn when handling such suspect oil. If shown to be acid, the oil must be removed and carefully disposed of and the system thoroughly cleaned out.
Moisture reacts with POE lubricants and this gives rise to organic acids. These are much weaker than the halogen acids and this needs to be taken into account when the acidity level is measured using an acid test kit. Nevertheless any breakdown of the POE is undesirable and it is recommended that the moisture level be kept below 50 ppm in POE systems. Because POEs have a high affinity for moisture it is essential to keep exposure of the oil to atmospheric air to an absolute minimum. The moisture cannot be fully removed by vacuum procedures and hence appropriate filter–driers are always recommended. Figure 5.4 illustrates a typical expected moisture content variance during the process of installing and commissioning. A compressor which is pre-charged with POE oil is connected first to a factory assembled unit, such as a compressor pack, installed on site observing correct evacuation and sealing procedures, and finally with subsequent running the moisture content is reduced towards 50 ppm by the drier.
Proper evacuation will remove the air content to an acceptable minimum. Oxidation and other high-temperature chemical reactions are more likely to cause damage in the presence of moisture.
When systems are changed over from CFCs or HCFCs to HFC refrigerants, the oil also generally needs to be changed to POE. The original equipment manufacturers recommendation should be sought prior to attempting such a conversion.
5.4 OIL SEPARATORS
During the compression stroke of a reciprocating machine, the gas becomes hotter and some of the oil on the cylinder wall will pass out with the discharge gas. Some oil carry-over will occur with all lubricated compressor types, and in small self-contained systems it quickly finds its way back to the compressor. Start up after a long idle period can result in a large amount of oil carry- over for a short period due to foaming. With larger more complex systems with remote evaporators oil it is desirable to fit an oil separator in the discharge line to reduce carry-over to the system (see Figure 5.5).