　　　本论文设计是对液体灌装生产线灌装阀的设计，以及液体的灌装方法在旋转灌装机上的布置与安排。该液体灌装阀的结构是阀体的上部有进液口，阀的下部有出液口，且出口与灌装头部位有着密封装置，通过挤压力与密封装置接触，阀体上部设有排气管，该排气管的一端从阀体内部穿过之后伸出液体出口与灌装头连接在一起而另外一端则伸出液体与外部联通从而排除气体，在液体灌装阀导通是阀体的上部进液口与下部出液口会形成一个液体通道使得液体可以从此通道进入瓶体，灌装头的最大尺寸与大于液道的最大尺寸，和滑动体的外径对齐，可以避免灌装过程中液体冲击力形成大量泡沫等对灌装精度和效率的影响，在排气通道在安装压紧螺母处设有螺纹,用来安装压紧螺母，阀可以通过调节螺母和压紧螺母调节弹簧的张力和阀体尺寸，用来调节灌装过程中一些小的光装误差，本发明在于灌装完成后能及时停止灌装,灌装精度高,节约能源,应用广泛，效率高，且液体灌装阀结构简单。

　　　关键词：排气管；弹簧；布置；弹性低压件

摘要III

AbstractIV

目录V

1 绪论1

1.1 本课题的研究内容和意义1

1.2 国内外的发展概况1

1.3 本课题应达到的要求2

2 灌装生产线整体结构设计3

2.1 方案的选取3

2.1.1 直线型灌装机3

2.1.2 旋转型灌装机3

2.1.3 自动化灌装机4

2.2 生产线各机构的设计5

2.2.1 灌装的供瓶机构5

2.2.2 灌装的供料机构6

2.2.3 灌装阀的升降机构7

2.3 本章小结8

3 罐装的基本原理和灌装阀的分析与设计9

3.1 罐装的基本原理9

3.1.1 灌装的基本方法9

3.1.2 定量方法12

3.2 灌装阀的分析与设计14

3.2.1 灌装阀的工作原理14

3.2.2 阀的各部分的设计与计算14

3.2.3 灌装时间的计算和过程16

3.2.4 灌装阀密封材料的选择19

3.2.5 灌装阀门启动结构设计与分类20

3.2.6 灌装阀弹簧的分析与设计21

3.3 本章小结30

4 灌装过程的调整及机器存在的问题31

4.1 灌装过程的调节31

4.1.1 料缸液位的调整31

4.1.2 灌装量的调整31

4.1.3 转速的调整31

4.1.4 机器存在的主要问题32

5 电气控制阀和调试维护33

5.1 电气的控制33

5.2 主要制动过程33

5.3 简单的生产线运动控制33

5.4 设备调试与维护33

5.4.1 整机要求33

5.4.2 灌装阀的调试与维护33

5.4.3 安全操作规则34

6 结论与展望35

6.1 结论35

6.2 展望35

致谢36

参考文献37

附录38

等压法灌装供料机构图38

压力法供料机构38

1 绪论

　　　对于灌装机的主要主成部分灌装阀来说，缺少了灌装阀自动灌装生产线就无法运转工作，且一个灌装阀的好坏决定了这个灌装生产线上运作的效率，它需要根据灌装工艺的要求以最快的速度联通或者切断与储液箱的联系，保证灌装工作的顺利的进行，且由于不同的液体的物理化学性质并不是相同的，故而导致了灌装工艺的不同，因此所使用的阀体也并不相同，不同的灌装使用不同得阀体。诸如饮料，酒类，液体化妆品之类。

　　　为了能够更好地实现灌装生产线的自动化，提高瓶装生产线的效率，解决自动化灌装生产的各种问题，保证产品的质量，特进行本课题关于液体灌装生产线灌装阀的设计研究。

1.1 本课题的研究内容和意义

　　　随着科学技术的日益发展与进步，人们的生活水平也是逐步的提高，由以前小农社会的自给自足发展到由大型公司生产人们去购买的商品经济，而在此过程中由于人口基数的增大同时还伴随着人工成本的增加，自动化的生产需求已迫在眉睫。

　　　为了满足国内的消费需求弥补国内生产力的不足，我国从上世纪八十年代开始每年都要进口大量的饮料、奶类制品以及酒类等包装机械，至今引进的势头仍然是呈现上涨势头。这些机械大部分是高速自动化的生产线，可靠性比较强，产量高，相当的部分设备是当今世界最为先进的机型。这些生产线的引进，使中国饮料、奶类制品以及酒类企业包装水平得以与发达的国家同步发展。与此同时，中国的包装机械的生产也取得了长足进步，部分灌装、封口一体设备已经达到了较高水平，包括塑料饮料瓶、酸奶杯、无菌包装等成型设备以及贴标机在内的包装生产线的水平也得到了相当程度的提升，基本可以满足中型企业的需要，部分已经可以替代进口设备，并且出口量逐年提高，灌装阀作为灌装机中不可缺少的部分，应用十分广泛，每年使用量相当巨大，国内在灌装速度和精度上和国际水平还存在着一定的差距，因此对灌装阀重新研究设计若是可以取得成功将大大推动中国包装机械行业的发展，提高包装行业在国际中的低位。

1.3 本课题应达到的要求

　　　鉴于饮料酒类一类的以及化妆品等生活用品已在人们的生活中占有越来越大的比例，从而大大的带动了灌装行业的迅速发展。

　　　故而在对灌装阀设计的过程中我们应该立足于当代现实情况，要能够满足自动化的生产且拥有一定的效率，同时在设计产品的过程中要全面的了解灌装的原理以及各过程的详细步骤，对于各部件的功能需要详细的了解，在保证产品的功能的基础之上尽量提高生产的效率，考虑灌装阀在灌装整体中的作用，对灌装机的全局整体机构和布置进行考虑，合理的对灌装阀进行布置以达到更高的生产效率。

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灌装生产线灌装阀的设计.gif

内容简介：: 无锡太湖学院信机系机械工程及自动化专业毕业设计论文任务书一、题目及专题：1、题目灌装生产线上灌装阀的设计 2、专题二、课题来源及选题依据在现代灌装厂中，灌装机和灌装阀已经成为了一个厂的命脉。一个高效率高精度的灌装阀可以为一个灌装长带来强大的活力。灌装阀的设计需要了解灌装生产工艺的各项需求，且设计到不少专业知识。此设计难度中等，设计量合适，可以很大程度提升学生的专业水准。随着技术的发展，生活水准的提高，灌装技术在未来必将发挥出巨大的作用。三、本设计（论文或其他）应达到的要求：能够详细的了解灌装生产线的灌装工序，熟悉灌装的每一步的过程，以及各部件的功效；详细的了解灌装机的全局机构和布置；详细的了解灌装机的传动机构，以及拨瓶机构的传动机构；详细的了解灌装阀工作的原理，灌装阀的组成，以及灌装阀工作时的各个步骤；四、接受任务学生：机械97 班姓名杭建国五、开始及完成日期：自2012年11月7日至2013年5月25日六、设计（论文）指导（或顾问）：指导教师签名签名签名教研室主任学科组组长研究所所长签名系主任签名2012年11月12日英文原文Screw CompressorsThe direction normal to the helicoids, can be used to calculate the coordinates of the rotor helicoids and from x and y to which the clearance is added as:，，（2.19）where the denominator D is given as：（2.20） and serve to calculate new rotor end plane coordinates, x0n and y0n,with the clearances obtained for angles = /p and respectively. These and now serve to calculate the transverse clearance 0 as the difference between them, as well as the original rotor coordinates and .If by any means, the rotors change their relative position, the clearance distribution at one end of the rotors may be reduced to zero on the flat side of the rotor lobes. In such a case, rotor contact will be prohibitively long on the flat side of the profile, where the dominant relative rotor motion is sliding, as shown in Fig. 2.29. This indicates that rotor seizure will almost certainly occur in that region if the rotors come into contact with each other.Fig. 2.29. Clearance distribution between the rotors: at suction, mid rotors, and discharge with possible rotor contact at the dischargeFig. 2.30. Variable clearance distribution applied to the rotorsIt follows that the clearance distribution should be non-uniform to avoid hard rotor contact in rotor areas where sliding motion between the rotors is dominant.In Fig. 2.30, a reduced clearance of 65 m is presented, which is now applied in rotor regions close to the rotor pitch circles, while in other regions it is kept at 85 m, as was done by Edstroem, 1992. As can be seen in Fig. 2.31, the situation regarding rotor contact is now quite different. This is maintained along the rotor contact belt close to the rotor pitch circles and fully avoided at other locations. It follows that if contact occurred, it would be of a rolling character rather than a combination of rolling and sliding or even pure sliding. Such contact will not generate excessive heat and could therefore be maintained for a longer period without damaging the rotors until contact ceases or the compressor is stopped.2.6 Tools for Rotor ManufactureThis section describes the generation of formed tools for screw compressor hobbing, milling and grinding based on the envelope gearing procedure.2.6.1 Hobbing ToolsA screw compressor rotor and its formed hobbing tool are equivalent to a pair of meshing crossed helical gears with nonparallel and nonintersecting axes. Their general meshing condition is given in Appendix A. Apart from the gashes forming the cutter faces, the hob is simply a helical gear in which.Fig. 2.31. Clearance distribution between the rotors: at suction, mid of rotor and discharge with a possible rotor contact at the dischargeEach referred to as a thread, Colburne, 1987. Owing to their axes not being parallel, there is only point contact between them whereas there is line contact between the screw machine rotors. The need to satisfy the meshing equation given in Appendix A, leads to the rotor hob meshing requirement for the given rotor transverse coordinate points and and their first derivative.The hob transverse coordinate points and can then be calculated. These are sufficient to obtain the coordinate The axial coordinate , calculated directly, and are hob axial plane coordinates which define the hob geometry.The transverse coordinates of the screw machine rotors, described in the previous section, are used as an example here to produce hob coordinates. he rotor unit leadsare 48.754mm for the main and 58.504mm for the ate rotor. Single lobe hobs are generated for unit leads :6.291mm for the main rotor and 6.291mm for the gate rotor. The corresponding hob helix angles are 85 and 95. The same rotor-to-hob centre distance C = 110mm and the shaft angle = 50 are given for both rotors. Figure 2.32 contains a view to the hob.Reverse calculation of the hob screw rotor transformation, also given in Appendix A permits the determination of the transverse rotor profile coordinates which will be obtained as a result of the manufacturing process. These ay be compared with those originally specified to determine the effect ofFig. 2.32. Rotor manufacturing: hobbing tool left, right milling toolmanufacturing errors such as imperfect tool setting or tool and rotor deformation upon the final rotor profile.For the purpose of reverse transformation, the hob longitudinal plane coordinatesand andshould be given. The axial coordinate is used to calculate , which is then used to calculate the hob transverse coordinates:，（2.21）These are then used as the given coordinates to produce a meshing criterionand the transverse plane coordinates of the “manufactured” rotors.A comparison between the original rotors and the manufactured rotors is given in Fig. 2.33 with the difference between them scaled 100 times. Two types of error are considered. The left gate rotor, is produced with 30um offset in the centre distance between the rotor and the tool, and the main rotor withFig. 2.33. Manufacturing imperfections0.2 offset in the tool shaft angle . Details of this particular meshing method are given by Stosic 1998.2.6.2 Milling and Grinding ToolsFormed milling and grinding tools may also be generated by placing in the general meshing equation, given in Appendix A, and then following the procedure of this section. The resulting meshing condition now reads as: (2.22)However in this case, when one expects to obtain screw rotor coordinates from the tool coordinates, the singularity imposed does not permit the calculation of the tool transverse plane coordinates. The main meshing condition cannot therefore be applied. For this purpose another condition is derived for the reverse milling tool to rotor transformation from which the meshing angle is calculated: (2.23)Once obtained, will serve to calculate the rotor coordinates after the “manufacturing” process. The obtained rotor coordinates will contain all manufacturing imperfections, like mismatch of the rotor tool centre distance, error in the rotor tool shaft angle, axial shift of the tool or tool deformation during the process as they are input to the calculation process. A full account of this useful procedure is given by Stosic 1998.2.6.3 Quantification of Manufacturing ImperfectionsThe rotor tool transformation is used here for milling tool profile generation. The reverse procedure is used to calculate the “manufactured” rotors. The rack generated 5-6 128mm rotors described by Stosic, 1997a are used as given profiles: x(t) and y(t). Then a tool rotor transformation is used to quantify the influence of manufacturing imperfections upon the quality of the produced rotor profile. Both, linear and angular offset were considered. Figure 2.33 presents the rotors, the main manufactured with the shaft angle offset 0.5 and the gate with the centre distance offset 40 m from that of the original rotors given by the dashed line on the left. On the right, the rotors are manufactured with imperfections, the main with a tool axial offset of 40 m and the gate with a certain tool body deformation which resulted in 0.5 offset of the relative motion angle . The original rotors are given by the dashed line.3Calculation of Screw Compressor PerformanceScrew compressor performance is governed by the interactive effects of thermodynamic and fluid flow processes and the machine geometry and thus can be calculated reliably only by their simultaneous consideration. This may be chieved by mathematical modelling in one or more dimensions. For most applications, a one dimensional model is sufficient and this is described in full. 3-D modelling is more complex and is presented here only in outline. A more detailed presentation of this will be made in a separate publication.3.1 One Dimensional Mathematical ModelThe algorithm used to describe the thermodynamic and fluid flow processes in a screw compressor is based on a mathematical model. This defines the instantaneous volume of the working chamber and its change with rotational angle or time, to which the conservation equations of energy and mass continuity are applied, together with a set of algebraic relationships used to define various phenomena related to the suction, compression and discharge of the working fluid. These form a set of simultaneous non-linear differential equations which cannot be solved in closed form.The solution of the equation set is performed numerically by means of the Runge-Kutta 4th order method, with appropriate initial and boundary conditions.The model accounts for a number of “real-life” effects, which may significantly influence the performance of a real compressor. These make it suitable for a wide range of applications and include the following: The working fluid compressed can be any gas or liquid-gas mixture for which an equation of state and internal energy-enthalpy relation is known, i.e. any ideal or real gas or liquid-gas mixture of known properties. The model accounts for heat transfer between the gas and the compressor rotors or its casing in a form, which though approximate, reproduces the overall effect to a good first order level of accuracy. The model accounts for leakage of the working medium through the clearances between the two rotors and between the rotors and the stationary parts of the compressor. The process equations and the subroutines for their solution are independent of those which define the compressor geometry. Hence, the model can be readily adapted to estimate the performance of any geometry or type of positive displacement machine. The effects of liquid injection, including that of oil, water, or refrigerant can be accounted for during the suction, compression and discharge stages. A set of subroutines to estimate the thermodynamic properties and changes of state of the working fluid during the entire compressor cycle of operations completes the equation set and thereby enables it to be solved.Certain assumptions had to be introduced to ensure efficient computation.These do not impose any limitations on the model nor cause significant departures from the real processes and are as follows: The fluid flow in the model is assumed to be quasi one-dimensional. Kinetic energy changes of the working fluid within the working chamber are negligible compared to internal energy changes. Gas or gas-liquid inflow to and outflow from the compressor ports is assumed to be isentropic. Leakage flow of the fluid through the clearances is assumed to be adiabatic.3.1.1 Conservation EquationsFor Control Volume and Auxiliary RelationshipsThe working chamber of a screw machine is the space within it that contains the working fluid. This is a typical example of an open thermodynamic system in which the mass flow varies with time. This, as well as the suction and discharge plenums, can be defined by a control volume for which the differential equations of the conservation laws for energy and mass are written. These are derived in Appendix B, using Reynolds Transport Theorem.A feature of the model is the use of the non-steady flow energy equation to compute the thermodynamic and flow processes in a screw machine in terms of rotational angle or time and how these are affected by rotor profile modifications. Internal energy, rather than enthalpy, is then the derived variable. This is computationally more convenient than using enthalpy as the derivedVariable since, even in the case of real fluids, it may be derived, without reference to pressure. Computation is then carried out through a series of iterative cycles until the solution converges. Pressure, which is the desired output variable, can then be derived directly from it, together with the remaining required thermodynamic properties.The following forms of the conservation equations have been employed in the model:中文翻译螺杆式压缩机几何的法线方向的螺旋，可以用来计算的坐标转子螺旋和的从x和y的间隙加入如：，， (2.19)其中分母D被给定为： (2.20)，服务来计算新的转子端的平面的坐标，和，得到的间隙角 =锌/ p和。这些，现在的差额计算的横向间隙0在它们之间，以及原来的转子坐标和。如果以任何方式，转子的改变它们的相对位置，该间隙的平侧面上分布在转子的一端，也可以减少到零的转子叶片。在这样的情况下，转子的接触将是令人望而却步长侧扁的档案中，其中占主导地位的相对滑动转子运动，如示于图。 2.29。这表明，转子扣押几乎肯定会如果转子进入彼此接触，发生在该区域。图.29。：吸力，中间转子和转子之间的间隙分布可能转子接触放电在放电图 2.30。可变间隙分布应用到转子如下的间隙分布应该非均匀，以避免在转子转子之间的滑动运动的地方是硬转子的接触占主导地位。另外，在图2.30 ，清除率降低65微米，这是现在应用在转子靠近转子节圆的区域，而在其他区域是保持在85微米所做的那样， 1992年由Edstroem 。正如在图中可以看出的。 2.31，现在的情况，转子的接触是完全不同的。这是保持靠近转子的节圆沿转子的接触带，并完全避免在其他位置。因此，如果发生接触，这将是一个滚动字符，而不是相结合的滚动和滑动，甚至是纯滑动。这样的接触不会产生过多的热量，因此可以保持一段较长时间，而不会损坏转子直到接触终止或使压缩机停止。2.6 为转子制造的工具本节描述了一代形成的工具螺杆压缩机滚齿，铣床和磨床的基础上信封资产负债程序。2.6.1 滚齿机工具螺杆压缩机转子和其形成的滚齿机工具相当于一个对相互啮合的交错轴斜齿轮与非平行不相交轴。他们的啮合条件一般除了见附录A。张裂缝形成刀具的面孔时，仅仅是一个螺旋齿轮的滚刀。图2.31。转子之间的间隙分布：在抽吸，中期的转子和可能转子接触放电在放电每个齿被称为作为一个线程， Colburne ，1987 。由于其自身的轴线不平行，它们之间的唯一的点接触，而有行螺杆机转子之间的接触。需要满足的啮合在附录A中给出的公式，导致转子A “滚刀啮合要求对于给定的转子横向坐标点X01和Y01和他们的第一次衍生。滚刀横向坐标点X02和Y02计算出来的。这是足够的，以获得坐标轴向坐标z2的，直接计算，和R2是滚刀轴向平面坐标它定义滚刀的几何形状。轴向坐标z2的，直接计算，和R2是滚刀轴向平面坐标它定义滚刀的几何形状。螺杆机转子的横向坐标，描述在前面的部分，被用来作为一个例子在这里产生滚刀坐标。转子单元导致p1的有48.754毫米的主和- 58.504毫米的门转子。单叶炉产生单位领导P2 ： 6.291毫米的主转子和闸转子- 6.291毫米为。相应的滚刀螺旋角分别为85 和95 。相同的转子滚刀中心距离C = 110毫米和轴角 = 50 给出了两个转子。图2.32中包含一个查看的炉灶。反向计算的炉灶 - 螺杆转子的改造，也给出了附录A，允许确定转子型线的横向坐标这将在制造过程中得到的结果。这些可能与最初指定的进行比较，以确定影响图。 2.32。转子制造业：滚齿刀具左，右铣刀制造的错误，如完美的工具设置或工具，转子变形后，最终的转子型线。如果在反向变换的目的，滚刀的纵向平面坐标R2和Z2和应给予。轴向坐标z2的使用计算 = z2/p2 ，然后将其用于计算滚刀横向的坐标：， (2.21)然后用这些作为给定的坐标，以产生一个啮合判据的横向平面上的坐标的“人造”转子。原来的转子之间的比较和所制造的转子给出图。 2.33与它们之间的区别缩放100倍。二被认为是类型的错误。左边的门转子，30微米抵消在该转子与该工具，和主旋翼之间的中心距离0.2 偏移刀具轴角。这个特殊的网格划分方法的详情给出由Stosic 1998。图。 2.33。制造缺陷2.6.2 铣削和磨削工具形成铣削和磨削工具也可以通过放置p2= 0产生啮合方程，一般在附录A中，然后按照本节的程序的。现在的啮合条件内容： (2.22)然而，在这种情况下，当一个人希望获得螺杆转子坐标从工具坐标，所施加的奇异性，不允许计算该工具的横向平面坐标。主要的啮合条件不能因此其应用。为此目的，另一个条件推导了扭转铣刀到转子的变换从该啮合角计算方法是: (2.23)一旦获得，后，将用来计算转子坐标“制造”的过程。得到的转子的坐标将包含所有制造不完善的地方，如不匹配的转子 - 刀具中心的距离，在转子中的误差 - 工具轴角度，

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