TJ01-008@上海某办公楼给排水系统毕业设计
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毕业设计论文
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TJ01-008@上海某办公楼给排水系统毕业设计,毕业设计论文
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Laminar and Turbulent Flow Observation shows that two entirely different types of fluid flow exist. This was demon- strated by Osborne Reynolds in 1883 through an experiment in which water was discharged from a tank through a glass tube. The rate of flow could be controlled by a valve at the outlet, and a fine filament of dye injected at the entrance to the tube. At low velocities, it was found that the dye filament remained intact throughout the length of the tube, showing that the particles of water moved in parallel lines. This type of flow is known as laminar, viscous or streamline, the particles of fluid moving in an orderly manner and retaining the same relative positions in successive cross- sections. As the velocity in the tube was increased by opening the outlet valve, a point was eventually reached at which the dye filament at first began to oscillate and then broke up so that the colour was diffused over the whole cross-section, showing that the particles of fluid no longer moved in an orderly manner but occupied different relative position in successive cross-sections. This type of flow is known as turbulent and is characterized by continuous small fluctuations in the magnitude and direction of the velocity of the fluid particles, which are accompanied by corresponding small fluctuations of pressure. When the motion of a fluid particle in a stream is disturbed, its inertia will tend to carry it on in the new direction, but the viscous forces due to the surrounding fluid will tend to make it conform to the motion of the rest of the stream. In viscous flow, the viscous shear stresses are sufficient to eliminate the effects of any deviation, but in turbulent flow they are inadequate. The criterion which determines whether flow will be viscous of turbulent is therefore the ratio of the inertial force to the viscous force acting on the particle. The ratio vlc o n s tfo r c eV is c o u s fo r c eI n e r tia l Thus, the criterion which determines whether flow is viscous or turbulent is the quantity vl/, known as the Reynolds number. It is a ratio of forces and, therefore, a pure number and may also be written as ul/v where is the kinematic viscosity (v=/). Experiments carried out with a number of different fluids in straight pipes of different diameters have established that if the Reynolds number is calculated by making 1 equal to the pipe diameter and using the mean velocity v, then, below a critical value of vd/ = 2000, flow will normally be laminar (viscous), any tendency to turbulence being damped out by viscous friction. This value of the Reynolds number applies only to flow in pipes, but critical values of the Reynolds number can be established for other types of flow, choosing a suitable characteristic length such as the chord of an aerofoil in place of the pipe diameter. For a given fluid flowing in a pipe of a given diameter, there will be a critical velocity of flow corresponding to the critical value of the Reynolds number, below which flow will be viscous. ntsIn pipes, at values of the Reynolds number 2000, flow will not necessarily be turbulent. Laminar flow has been maintained up to Re = 50,000, but conditions are unstable and any disturbance will cause reversion to normal turbulent flow. In straight pipes of constant diameter, flow can be assumed to be turbulent if the Reynolds number exceeds 4000. Pipe Networks An extension of compound pipes in parallel is a case frequently encountered in municipal distribution system, in which the pipes are interconnected so that the flow to a given outlet may come by several different paths. Indeed, it is frequently impossible to tell by inspection which way the flow travels. Nevertheless, the flow in any networks, however complicated, must satisfy the basic relations of continuity and energy as follows: 1. The flow into any junction must equal the flow out of it. 2. The flow in each pipe must satisfy the pipe-friction laws for flow in a single pipe. 3. The algebraic sum of the head losses around any closed circuit must be zero. Pipe networks are generally too complicated to solve analytically, as was possible in the simpler cases of parallel pipes. A practical procedure is the method of successive approximations, introduced by Cross. It consists of the following elements, in order: 1. By careful inspection assume the most reasonable distribution of flows that satisfies condition 1. 2. Write condition 2 for each pipe in the form hL = KQn (7.5) where K is a constant for each pipe. For example, the standard pipe-friction equation would yield K = 1/C2 and n = 2 for constant f. Minor losses within any circuit may be included, but minor losses at the junction points are neglected. 3. To investigate condition 3, compute the algebraic sum of the head losses around each elementary circuit. hL = KQn. Consider losses from clockwise flows as positive, counterclockwise negative. Only by good luck will these add to zero on the first trial. 4. Adjust the flow in each circuit by a correction, Q, to balance the head in that circuit and give KQn = 0. The heart of this method lies in the determination of Q. For any pipe we may write Q = Q0 Q where Q is the correct discharge and Q0 is the assumed discharge. Then, for a circuit 0100 / Qhn hQKn QKQLLnn (7.6) ntsIt must be emphasized again that the numerator of Eq. (7.6) is to be summed algebraically, with due account of sign, while the denominator is summed arithmetically. The negative sign in Eq. (7.6) indicates that when there is an excess of head loss around a loop in the clockwise direction, the Q must be subtracted from clockwise Q0s and added to counterclockwise ones. The reverse is true if there is a deficiency of head loss around a loop in the clockwise direction. 5. After each circuit is given a first correction, the losses will still not balance because of the interaction of one circuit upon another (pipes which are common to two circuits receive two independent corrections, one for each circuit). The procedure is repeated, arriving at a second correction, and so on, until the corrections become negligible. Either form of Eq. (7.6) may be used to find Q. As values of K appear in both numerator and denominator of the first form, values proportional to the actual K may be used to find the distribution. The second form will be found most convenient for use with pipe-friction diagrams for water pipes. An attractive feature of the approximation method is that errors in computation have the same effect as errors in judgment and will eventually be corrected by the process. The pipe-networks problem lends itself well to solution by use of a digital computer. Programming takes time and care, but once set up, there is great flexibility and many man-hours of labor can be saved. The Future of Plastic Pipe at Higher Pressures Participants in an AGA meeting panel on plastic pipe discussed the possibility of using polyethylene gas pipe at higher pressures. Topics included the design equation, including work being done by ISO on an updated version, and the evaluation of rapid crack propagation in a PE pipe resin. This is of critical importance because as pipe is used at higher pressure and in larger diameters, the possibility of RCP increases. Several years ago, AGAs Plastic Pipe Design Equation Task Group reviewed the design equation to determine if higher operating pressures could be used in plastic piping systems. Members felt the performance of our pipe resins was not truly reflected by the design equation. It was generally accepted that the long-term properties of modern resins far surpassed those of older resins. Major considerations were new equations being developed and selection of an appropriate design factor. Improved pipe performance Many utilities monitored the performance of plastic pipe resins. Here are some of the long-term tests used and the kinds of performance change they have shown for typical gas pipe resins. nts Elevated temperature burst test They used tests like the Elevated Temperature Burst Test, in which the long-term performance of the pipe is checked by measuring the time required for formation of brittle cracks in the pipe wall under high temperatures and pressures (often 80 degrees C and around 4 to 5-MPa hoop stress). At Consumers Gas we expected early resins to last at least 170 hrs. at 80 degrees C and a hoop stress of 3 MPa. Extrapolation showed that resins passing these limits should have a life expectancy of more than 50 yrs. Quality control testing on shipments of pipe made from these resins sometimes resulted in product rejection for failure to meet this criterion. At the same temperature, todays resins last thousands of hours at hoop stresses of 4.6 MPa. Tests performed on pipe made from new resins have been terminated with no failure at times exceeding 5,700 hrs. These results were performed on samples that were squeezed off before testing. Such stresses were never applied in early testing. When extrapolated to operating conditions, this difference in test performance is equivalent to an increase in lifetime of hundreds (and in some cases even thousands) of years. Environmental stress crack resistance test Some companies also used the Environmental Stress Crack Resistance test which measured brittle crack formation in pipes but which used stress cracking agents to shorten test times. This test has also shown dramatic improvement in resistance brittle failure. For example, at my company a test time of more than 20 hrs. at 50 degrees C was required on our early resins. Todays resins last well above 1,000 hrs. with no failure. Notch tests Notch tests, which are quickly run, measure brittle crack formation in notched pipe or molded coupon samples. This is important for the newer resins since some other tests to failure can take very long times. Notch test results show that while early resins lasted for test times ranging between 1,000 to 10,000 min., current resins usually last for longer than 200,000 min. All of our tests demonstrated the same thing. Newer resins are much more resistant to the growth of brittle crack than their predecessors. Since brittle failure is considered to be the ultimate failure mechanism in polyethylene pipes, we know that new materials will last much longer than the old. This is especially reassuring to the gas industry since many of these older resins have performed very well in the field for the past 25 yrs. with minimal detectable change in properties. While the tests showed greatly improved performance, the equation used to establish the pressure rating of the pipe is still identical to the original except for a ntschange in 1978 to a single design factor for all class locations. To many it seemed that the methods used to pressure rate our pipe were now unduly conservative and that a new design equation was needed. At this time we became aware of a new equation being balloted at ISO. The methodology being used seemed to be a more technically correct method of analyzing the data and offered a number of advantages. Thermal Expansion of Piping and Its Compensation A very relevant consideration requiring careful attention is the fact that with temperature of a length of pipe raised or lowered, there is a corresponding increase or decrease in its length and cross-sectional area because of the inherent coefficient of thermal expansion for the particular pipe material. The coefficient of expansion for carbon steel is 0.012 mm/mC and for copper 0.0168mm/mC. Respective module of elasticity are for steel E = 2071.06kN/m2 and for copper E = 103106 kN/m2. As an example, assuming a base temperature for water conducting piping at 0C, a steel pipe of any diameter if heated to 120C would experience a linear extension of 1.4 mm and a similarly if heated to copper pipe would extend by 2.016 mm for each meter of their respective lengths. The unit axial force in the steel pipe however would be 39% greater than for copper. The change in pipe diameter is of no practical consequence to linear extension but the axial forces created by expansion or contraction are con- siderable and capable of fracturing any fitments which may tend to impose a restraint ; the magnitude of such forces is related to pipe size. As an example, in straight pipes of same length but different diameters, rigidly held at both ends and with temperature raised by say 100C, total magnitude of linear forces against fixed points would be near enough proportionate to the respective diameters. It is therefore essential that design of any piping layout makes adequate com- pensatory provision for such thermal influence by relieving the system of linear stresses which would be directly related to length of pipework between fixed points and the range of operational temperatures. Compensation for forces due to thermal expansion. The ideal pipework as far as expansion is concerned, is one where maximum free movement with the minimum of restraint is possible. Hence the simplest and most economical way to ensure com- pensation and relief of forces is to take advantage of changes in direction, or where this is not part of the layout and long straight runs are involved it may be feasible to introduce deliberate dog-leg offset changes in direction at suitable intervals. As an alternative, at calculated intervals in a straight pipe run specially designed expansion loops or “U” bends should be inserted. Depending upon design and space availability, expansion bends within a straight pipe run can feature the so called double offset “U” band or the horseshoe type or “lyre” loop. The last named are ntsseldom used for large heating networks; they can be supplied in manufacturers standard units but require elaborate constructional works for underground installation. Anchored thermal movement in underground piping would normally be absorbed by three basic types of expansion bends and these include the “U” bend, the “L” bend and the “Z” bend. In cases of 90 changes indirection the “L” and “Z” bends are used. Principles involved in the design of provision for expansion between anchor points are virtually the same for all three types of compensator. The offset “U” bend is usually made up from four 90 elbows and straight pipes; it permits good thermal displacement and imposes smaller anchor loads than the other type of loop. This shape of expansion bend is the standardised pattern for prefabricated pipe-in-pipe systems. All thermal compensators are installed to accommodate an equal amount of expansion or contraction; therefore to obtain full advantage of the length of thermal movement it is necessary to extend the unit during installation thus opening up the loop by an extent roughly equal the half the overall calculated thermal movement. This is done by “cold-pull” or other mechanical means. The total amount of extension between two fixed points has to be calculated on basis of ambient temperature prevailing and operational design temperatures so that distribution of stresses and reactions at lower and higher temperatures are controlled within permissible limits. Pre-stressing does not affect the fatigue life of piping therefore it does not feature in calculation of pipework stresses . There are numerous specialist publication dealing with design and stressing calculations for piping and especially for proprietary piping and expansion units; comprehensive experience back design data as well as charts and graphs may be obtained in manufacturers publications, offering solutions for every kind of pipe stressing problem. As an alternative to above mentioned methods of compensation for thermal expansion and useable in places where space is restricted, is the more expensive bellows or telescopic type mechanical compensator. There are many proprietary types and models on the market and the following types of compensators are generally used. The bellows type expansion unit in form of an axial compensator provides for expansion movement in a pipe along its axis; motion in this bellows is due to tension or compression only. There are also articulated bellows units restrained which combine angular and lateral movement; they consist of double compensator units restrained by straps pinned over the center of each bellowsor double tied thus being restrained over its length. Such compensators are suitable for accommodating very pipeline expansion and also for combinations of angular and lateral movements. 层流与紊流 有两种 完全不同的流体流动 形式存在，这一点在 1883 年就由 Osborne ntsReynolds 用试验演示证明。在试验里，水通过玻璃管从水箱里放出。流量由出口处的阀门来控制，一股很细的染色流束由入口注入玻璃管内。在较低的流速时，可以看到染色流束在玻璃管中保持着一条完整的迁流。这表明流体粒子以平行的层状流动。这种粘性流体的流动就是我们所知的层流，流体各层的质点以有序的方式移动，并在连续的截面上保持着相同的相对位置。 打开出口阀门，管子里的速度就提高。随着速度提高，最后会达到这样的程度，即染色流束起初开始摆动然后破碎，这样颜色就扩散在整个截 面上，这表明流体粒子已不再有次序流动却在连续的截面上占有相对不同的位置。这种流体的流动形式就是紊流，它的特点就是不断产生无数大小不等的涡团，质点掺混使得空间各点的速度随时间无规则地变化。与之相关联，压强也随之无规则地变化。 当一条流束中的某个流体粒子的运动被扰乱，则它的惯性会使它移向新的方向，但周围流体的粘滞力会使它与其余流束的运动保持一致。在粘性流体中，粘性切应力足以抵消任何偏差的影响，但在紊流中是不够的。因此，确定流动是粘滞性的还是紊流性的标准就是作用在粒子上的惯性力和粘性力之比： vlc o n s tfo r c eV is c o u s fo r c eI n e r tia l 这样，用来判断流动是粘滞性的还是紊流性的标准就是 vl/， 也就是雷诺数。这是力之间的比，因此理论上也可以写成 ul/v（ v=/，流体的运动粘滞系数）。 在不同管径的直管里用许多不同流体所进行的试验已经证实，如雷诺数是通过使 L 等于管径并且使用平均速度 v 来计算，那么在低于临界值 vd/ = 2000 的条件下流动一般是层流（粘滞流动），任何紊流的倾向都会由于粘滞摩擦而受到抑制。这个雷诺数的值仅适用于管道中的流体，但雷诺数的临界值可以用来确定其他形式的流动，例如选择合适的弦杆翼剖面 来代替管道直径。对于已知直径的管道中的流体而言，会有一个临界流速 vc，以及对应的雷诺数，如果低于这个数，则表明流体是粘滞流动。 在管道中，雷诺数值大于 2000 的情况下，流体不一定就变为紊流。层流可以维持到 Re = 50,000,但是条件并不稳定，任何干扰都会使其它又变为一般的紊流。在直径一定的直管中，如果雷诺数超过 4000 那么流体就有可能变为紊流。 管网 平行复合管道的延伸是市政分配系统中常见的一种情况，在这种情况下管道相互连接，使得通向某一出口的流体可以来自不同的路径。的确，通过观察往往很难说清 楚流体将流经哪一个管路。但是，不管管网有多复杂，其中的流体都必须确保连续性与能量的基础关系。如下所述： 1 流入接合处的流体必须与流出的等量； 2 在每根管中的流体都必须满足流体在单管中的管道摩擦定律； 3 在任何闭合回路中，水头损失的代数和必须为 0。 管网一般来讲由于太过复杂而难以分析解决，但在简单一些的情况下是可以 nts的，例如平行管。 Cross 介绍了一种实用的程序，采用的是连续性近似法。它由以下的原理组成，包括： 1 通过仔细的观察采取最合理的流体分配方案以满足条件 1； 2 对每根管道以方程 hL = KQn 来判断是否满足条
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