上海某辦公樓給排水系統(tǒng)畢業(yè)設(shè)計(jì),上海,辦公樓,排水系統(tǒng),畢業(yè)設(shè)計(jì) 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 Thus, the criterion which determines whether flow is viscous or turbulent is the quantity ρvl/μ, 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=μ/ρ). 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 ρvd/μ = 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. In 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, ΔQ, to balance the head in that circuit and give ∑KQn = 0. The heart of this method lies in the determination of ΔQ. For any pipe we may write Q = Q0 ＋ΔQ where Q is the correct discharge and Q0 is the assumed discharge. Then, for a circuit (7.6) It 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 ΔQ must be subtracted from clockwise Q0’s 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 ΔQ. 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, AGA’s 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. 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, today’s 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. Today’s 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 change 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/m?C and for copper 0.0168mm/m?C. Respective module of elasticity are for steel E = 207×1.06kN/m2 and for copper E = 103×106 kN/m2. As an example, assuming a base temperature for water conducting piping at 0?C, a steel pipe of any diameter if heated to 120?C 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 100?C, 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 “l(fā)yre” loop. The last named are seldom 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. 層流與紊流 有兩種完全不同的流體流動(dòng)形式存在，這一點(diǎn)在1883年就由Osborne Reynolds 用試驗(yàn)演示證明。在試驗(yàn)里，水通過(guò)玻璃管從水箱里放出。流量由出口處的閥門來(lái)控制，一股很細(xì)的染色流束由入口注入玻璃管內(nèi)。在較低的流速時(shí)，可以看到染色流束在玻璃管中保持著一條完整的遷流。這表明流體粒子以平行的層狀流動(dòng)。這種粘性流體的流動(dòng)就是我們所知的層流，流體各層的質(zhì)點(diǎn)以有序的方式移動(dòng)，并在連續(xù)的截面上保持著相同的相對(duì)位置。 打開(kāi)出口閥門，管子里的速度就提高。隨著速度提高，最后會(huì)達(dá)到這樣的程度，即染色流束起初開(kāi)始擺動(dòng)然后破碎，這樣顏色就擴(kuò)散在整個(gè)截面上，這表明流體粒子已不再有次序流動(dòng)卻在連續(xù)的截面上占有相對(duì)不同的位置。這種流體的流動(dòng)形式就是紊流，它的特點(diǎn)就是不斷產(chǎn)生無(wú)數(shù)大小不等的渦團(tuán)，質(zhì)點(diǎn)摻混使得空間各點(diǎn)的速度隨時(shí)間無(wú)規(guī)則地變化。與之相關(guān)聯(lián)，壓強(qiáng)也隨之無(wú)規(guī)則地變化。 當(dāng)一條流束中的某個(gè)流體粒子的運(yùn)動(dòng)被擾亂，則它的慣性會(huì)使它移向新的方向，但周圍流體的粘滯力會(huì)使它與其余流束的運(yùn)動(dòng)保持一致。在粘性流體中，粘性切應(yīng)力足以抵消任何偏差的影響，但在紊流中是不夠的。因此，確定流動(dòng)是粘滯性的還是紊流性的標(biāo)準(zhǔn)就是作用在粒子上的慣性力和粘性力之比： 這樣，用來(lái)判斷流動(dòng)是粘滯性的還是紊流性的標(biāo)準(zhǔn)就是ρvl/μ，也就是雷諾數(shù)。這是力之間的比，因此理論上也可以寫成ul/v（v=μ/ρ，流體的運(yùn)動(dòng)粘滯系數(shù)）。 在不同管徑的直管里用許多不同流體所進(jìn)行的試驗(yàn)已經(jīng)證實(shí)，如雷諾數(shù)是通過(guò)使L等于管徑并且使用平均速度v來(lái)計(jì)算，那么在低于臨界值ρvd/μ = 2000的條件下流動(dòng)一般是層流（粘滯流動(dòng)），任何紊流的傾向都會(huì)由于粘滯摩擦而受到抑制。這個(gè)雷諾數(shù)的值僅適用于管道中的流體，但雷諾數(shù)的臨界值可以用來(lái)確定其他形式的流動(dòng)，例如選擇合適的弦桿翼剖面來(lái)代替管道直徑。對(duì)于已知直徑的管道中的流體而言，會(huì)有一個(gè)臨界流速vc，以及對(duì)應(yīng)的雷諾數(shù)，如果低于這個(gè)數(shù)，則表明流體是粘滯流動(dòng)。 在管道中，雷諾數(shù)值大于2000的情況下，流體不一定就變?yōu)槲闪?。層流可以維持到Re = 50,000,但是條件并不穩(wěn)定，任何干擾都會(huì)使其它又變?yōu)橐话愕奈闪?。在直徑一定的直管中，如果雷諾數(shù)超過(guò)4000那么流體就有可能變?yōu)槲闪鳌? 管網(wǎng) 平行復(fù)合管道的延伸是市政分配系統(tǒng)中常見(jiàn)的一種情況，在這種情況下管道相互連接，使得通向某一出口的流體可以來(lái)自不同的路徑。的確，通過(guò)觀察往往很難說(shuō)清楚流體將流經(jīng)哪一個(gè)管路。但是，不管管網(wǎng)有多復(fù)雜，其中的流體都必須確保連續(xù)性與能量的基礎(chǔ)關(guān)系。如下所述： 1． 流入接合處的流體必須與流出的等量； 2． 在每根管中的流體都必須滿足流體在單管中的管道摩擦定律； 3． 在任何閉合回路中，水頭損失的代數(shù)和必須為0。 管網(wǎng)一般來(lái)講由于太過(guò)復(fù)雜而難以分析解決，但在簡(jiǎn)單一些的情況下是可以 的，例如平行管。Cross 介紹了一種實(shí)用的程序，采用的是連續(xù)性近似法。它由以下的原理組成，包括： 1． 通過(guò)仔細(xì)的觀察采取最合理的流體分配方案以滿足條件1； 2． 對(duì)每根管道以方程hL = KQn來(lái)判斷是否滿足條件2，式中K是每根管的特性 常數(shù)。例如，標(biāo)準(zhǔn)管道摩擦方程中的K = 1/C2以及n = 2。任何環(huán)路中較小的沿程水頭損失可能是包括的，但局部水頭損失可以忽略不計(jì)。 3． 為了研究條件3，計(jì)算每個(gè)基本環(huán)路中水頭損失的代數(shù)和?！苃L = ∑KQn。假 設(shè)順時(shí)針?lè)较蛄鲃?dòng)的損失為正，逆時(shí)針的則為負(fù)，那么在第一次試驗(yàn)中，它們的和只有在非常幸運(yùn)的情況下才會(huì)為零。 4． 通過(guò)一個(gè)修正值ΔQ來(lái)調(diào)整每條環(huán)路中的流體，使該管路中的水頭平衡，并 給出∑KQn = 0。這個(gè)方法的核心取決于ΔQ的確定。對(duì)于任何管道我們有： Q = Q0 ＋ΔQ 式中Q是準(zhǔn)確的流量而Q0是假定的流量。那么，對(duì)于一個(gè)環(huán)路而言： (7.6) 必須再次強(qiáng)調(diào)的是方程(7.6)的分子和分母都是采用了適當(dāng)?shù)挠?jì)算符號(hào)確定的。方程(7.6)中的負(fù)號(hào)表明，當(dāng)順時(shí)針?lè)较虻沫h(huán)路上有過(guò)量的水頭損失時(shí)，ΔQ必須從順時(shí)針?lè)较虻腝0中減去，并增加到逆時(shí)針?lè)较蛏先ァＨ绻槙r(shí)針?lè)较虻沫h(huán)路上水頭損失不足時(shí)，情況正好相反。 5．在每條環(huán)路都給予了一個(gè)最初的修正值后，由于環(huán)路之間的相互影響，損失仍不平衡（一些兩條環(huán)路共有的管道就有兩個(gè)單獨(dú)的修正值，每個(gè)值對(duì)應(yīng)一條環(huán)路）。重復(fù)這樣的程序，獲得第二個(gè)修正值，乃至第三、第四個(gè)等等，直到修正值可以忽略不計(jì)。 方程(7.6)的兩種形式都可以用來(lái)找出ΔQ。由于K值同時(shí)出現(xiàn)在第一種形式的分子和分母上，相應(yīng)實(shí)際的K值就可以用來(lái)確定分配量。結(jié)合水管的管道摩擦力圖表，第二種方程形式使用起來(lái)最簡(jiǎn)便。 近似法最吸引人的一個(gè)特點(diǎn)就是計(jì)算上的誤差與判斷誤差有相同的效果，而最終它們會(huì)在過(guò)程中被加以改正。 管網(wǎng)問(wèn)題非常適合于采用計(jì)算機(jī)來(lái)解決。編制程序需花費(fèi)大量的時(shí)間和精力，但是一旦完成，就有很大的機(jī)動(dòng)靈活性，許多耗人費(fèi)時(shí)的勞動(dòng)就可省去。 更高壓力下塑料管道的前景 美國(guó)煤氣協(xié)會(huì)AGA的一個(gè)針對(duì)塑料管道的專案小組的成員討論了在較高壓力下使用聚乙烯輸氣管的的可能性。討論的主題包括有設(shè)計(jì)方程（其中包括國(guó)際科學(xué)組織ISO在更新版本上完成的工作），以及對(duì)PE管樹(shù)脂上裂縫快速擴(kuò)展的評(píng)估。這一點(diǎn)非常重要，因?yàn)楫?dāng)管道在較高壓力下使用、而管徑更大的情況下，鋼筋混凝土管的可能性增加了。 若干年以前，AGA的塑料管道設(shè)計(jì)任務(wù)小組檢查了設(shè)計(jì)方程，以確定是否能在塑料管道系統(tǒng)中使用更高的工作壓力。小組成員認(rèn)為管道樹(shù)脂的性能并沒(méi)有通過(guò)設(shè)計(jì)方程反映出來(lái)。一般認(rèn)為新的樹(shù)脂塑管在耐用性上遠(yuǎn)遠(yuǎn)勝過(guò)過(guò)去的樹(shù)脂塑管，因此主要考慮的問(wèn)題是新方程的發(fā)展以及合適的設(shè)計(jì)要素的選擇。 改良的管道性能 許多設(shè)備用來(lái)監(jiān)測(cè)塑料管道樹(shù)脂的性能。在這里講述一下一些針對(duì)典型的輸氣管道樹(shù)脂進(jìn)行過(guò)的耐久性測(cè)試，以及幾種性能上的變化。 溫升爆裂測(cè)試 他們使用像溫升爆裂測(cè)試之類的測(cè)試。在這一測(cè)試中管系的耐久性能通過(guò)高溫和高壓下管壁形成脆裂所需的時(shí)間來(lái)校核（通常是80攝氏度和4-5MPa的環(huán)壓下）。在供應(yīng)燃?xì)鈺r(shí)我們希望老的樹(shù)脂塑管在80攝氏度、3MPa的環(huán)壓下至少可以堅(jiān)持使用170個(gè)小時(shí)。推斷表明通過(guò)了這些極限的樹(shù)脂預(yù)期其壽命應(yīng)該能超過(guò)50年。裝運(yùn)時(shí)對(duì)這些樹(shù)脂塑管質(zhì)量檢測(cè)，有時(shí)會(huì)由于沒(méi)有達(dá)到這一標(biāo)準(zhǔn)而對(duì)該產(chǎn)品拒絕使用。 在相同溫度條件下，今天的樹(shù)脂塑管在4.6MPa環(huán)壓下可持續(xù)使用數(shù)千小時(shí)。測(cè)試表明用新樹(shù)脂制造的管道可使用超過(guò)5700小時(shí)而沒(méi)有任何損壞。這些結(jié)果是在臨測(cè)試前檢出的（樹(shù)脂）抽樣得出的。這種壓力從未在早期的測(cè)試中使用過(guò)。根據(jù)工作條件推斷，測(cè)試性能上的區(qū)別與數(shù)百年的壽命增長(zhǎng)是相等的（某些情況下甚至是數(shù)千年）。 環(huán)壓下的防裂測(cè)試 也有些公司進(jìn)行了環(huán)壓下的防裂測(cè)試，用來(lái)測(cè)量管道中脆裂的形成，并加大了壓力來(lái)減短測(cè)試的時(shí)間。 這個(gè)試驗(yàn)表明了在防止脆裂上的驚人的改進(jìn)。例如，在我的公司里對(duì)于我們的早期樹(shù)脂塑管進(jìn)行試驗(yàn)需要20小時(shí)以上的時(shí)間和50攝氏度的溫度。而現(xiàn)在的樹(shù)脂塑管能夠良好地持續(xù)1000小時(shí)以上而沒(méi)有損壞。 槽口測(cè)試 可以快速進(jìn)行的槽口測(cè)試，用來(lái)測(cè)量帶有槽口的管道或?qū)ｉT澆鑄的試驗(yàn)管中脆裂的形成。這對(duì)新的樹(shù)脂塑管非常重要，因?yàn)槠渌脑囼?yàn)需要很長(zhǎng)的時(shí)間才能使管道發(fā)生損壞。槽口測(cè)試的結(jié)果說(shuō)明早期的樹(shù)脂塑管持續(xù)的試驗(yàn)時(shí)間在1000到10000分鐘之間，而現(xiàn)在的樹(shù)脂塑管則通常可持續(xù)超過(guò)200000分鐘。 我們所有的試驗(yàn)證實(shí)了相同的結(jié)果。更新的樹(shù)脂塑管比起它們的前輩，對(duì)脆裂的防止有著更好的效果。由于認(rèn)為脆裂是聚乙烯管道中結(jié)構(gòu)的最終損壞，因此我們知道新的材料比起舊的來(lái)能夠持續(xù)使用更久。這對(duì)于燃?xì)夤I(yè)特別可靠，因?yàn)樵S多這些舊的樹(shù)脂塑管在過(guò)去的25年時(shí)間里表現(xiàn)得非常好，而它們的性能只在最小范圍內(nèi)進(jìn)行了些改變。 測(cè)試表明了管道性能很大的改進(jìn)，過(guò)去用來(lái)建立管道壓力等級(jí)的方程式仍然與原有的相同，除了1978年對(duì)于一個(gè)針對(duì)所有等級(jí)的設(shè)計(jì)因素的改變。 從許多方面來(lái)看，如今將管道按壓