换热器英语文献

Floating-head heat exchangersThe present invention relates to an improvement in floating-head type heat exchangers and particularly to means for providing fluid-tight contact between a floating tube sheet and a head flange such exchangers. More particularly, this invention is concerned with a special packing joint which provides an effective seal between the shell side and the tube side and the tube side of a floating-head type heat exchanger. The use of so-called”shell and tube”heat exchangers has gained widespread commercial acceptance. Such exchangers are useful in transferring heat between two liquids, such as for example, in oil refining operation, or between a liquid and a vapor, such as for example in steam power plants. The relative thermal expansions or contraction which frequently result from differences in the temperatures of the fluids flowing through the tubes and the shell, or from differences in the materials of construction of the tubes and the shells, have led to the development and use of the so-called”floating-head”heat exchangers. In this type of exchanger the tubes are rigidly attached to a stationary tube sheet which is fixed relative to the shell of the exchanger at one end,and are attached to a floating tube sheet at the other end. Hence, in operation,slidable movement obtains between the floating tube sheet and the shell and other fixed parts as the expanding and contracting tubes cause movement of the floating tube sheet, and stress and strains which may otherwise cause wear and failure in the exchanger are therefore avoided.In the shell and tube heat exchangers, one fluid usually enters the shell at one end and discharges from the opposite side of the shell at the other end, or at the same end, depending upon whether a singer-pass or a double-pass arragement of tubes is employed. The other fluid flows through the tubes from one end and is discharged at the other end(single pass flow), or the fluid may flow through part of the tubes at one end,re-routed through another part of the tubes, in which case such arrangement is referred to as multiple pass flow. Indouble-pass flow, for example, the liquid enters half the tubes at one end and flows, say from right to left, discharges into a receiving chamber, and then re-routed to the remaining half of the tubes through which it flows from left to right, and finally discharges from the heat exchanger.It can be readily appreciated that in this type of heat exchanger, provisions must be made to prevent leakage of fluid from the shell side to the tube side or vice versa. Several suchprovisions have been suggested and adapted to these exchanger but they are all disadvantageous in one way or another. Most frequently, the floating-head is either bolted or clamped onto the floating tube-sheet and the entire assembly is then mounted in the shell by means of conventional expansion rings or packing joints. This arragement, however, is expensive, cumbersome and difficult to install and to disassemble.Accordingly, this invetion comprehends and resides in the discovery of novel means for providing fluid-tight contact between the floating tube-sheet and the head flange protion of the floating head heat exchanger. The novel means employed herein comprises two rings, preferably metallic, with packing materials thereon, said rings being separated by compressible and resilient members, such as, for example, spring washers. These washers are arranged each over one of a multiplicity of circumferentially arranged pins extending longitudinally between the rings, fixed to one ring and slidably movable through aligned holes in the other ring. The pins serve to keep the washers in position.The novel means employed in the present invention and its adaptation to the floating-head heat exchanger are more readily comprehended with reference to the attached drawings wherein:FIGURE 1 is a partially sectionalized side elevation of a floating-head type heat exchanger embodying this invention;FIGURE 2 is an enlarged section showing the detials of a floating-head joint of FIGURE 1, and FIGURE 3 is an isometric free-body view of the two rings showing their relative positions with the washers.In these drawings, like numerals designate like parts.Referring to the drawings,there is shown, on thefloating-head side of the heat exchanger, a shell flange 11, gasket 13 and head flang 15, all connected together via bolt 17. Also shown on this side of the exchanger is a floating tube-sheet 19 to which is attached a multiplicity of tubes 21 through which one fluid madium flows. The other fluid medium enters the shell 23 of the exchanger at enrance 25, flows through the shell in contact with the outer surface of tubes 21 and leave the shell at exit 27.Forming a liquid-tight contact between the floating bute-sheet 19, the shell flange 11 and head flange 15 there is shown a unitary structure comprising two metallic rings 29 and 31 which are connected via two or more pins 33. There pins are attached at one end to one of said rings, say,ring 31 by welding or any other suitable means, and at the other end the pins areinserted in apertures in the ring 29 through which the pins are free to move in axial direction. A resilient and compressible member 35, such as,for example, a spring wsaher, is set over each pin, which member is responsive to the relative movements and expansions and contractions at the floating-head joints. Pins 33, apart from their function of connecting the two metal rings, also serve to hold the spring washers in the circumferential array shown.The unitary structure referred to above is placed in a recess 37(or a groove, or a notch) specially cut in the flanges, and the remaining space on either side of the rings is filled with packing materials 39 and 41 to fill up the recess. The diameter of the packing materials is preferably slightly larger than the outside diameter of the two metal rings to provide an effective seal as will hereinafter be explained.Rings 29 and 31 can be of metallic or plastic materials capable of withstanding the compressive forces exerted thereon by the spring washers 39 during the operation of the heat exchanger. The washers 35 may be of any suitable resilient and compressible materials capable of responding to the thermal expansions and contractions resulting from the differences in the temperatures of the fluids flowing through the shell andthe tube, or to differences in the materials of construction of the shell and the tubes. The number washers can very depending upon the compressive forces exerted in the floating-head joint.In assembling the heat exchangher, when bolts 17 are tightened, gasket 13 is compressed between the surfaces of the shell flange 11 and head flange 15. The compressive forces so set up are transmitted to the packing material 39 and 41 which packing material are therefore compressed away from each other as well as against the floating tube-sheet 19 the shell flange 11 and the head flange 15. Thus a fluid-tight contact is provided between the shell side and the tube side of the heat exchanger at floating-head joint. The compressive forces which are so transmitted to the packing matreial are in turn absorbed by the spring washers 35 which remain compressed in response to these compressive forces and which can return to their normal uncompressed position upon the removal of these forces.Thermal expansions and contractions at the floating-head joint,resulting from differences in temperature or differences in the materials of construction, as was previously discussed, cause relative movement of the floating tube-sheet 19 with respect to the shell flange 11 and head flange 15. The noveldevice permits the tube sheet to slide against the surfaces of the shell flange and the head flange and at the same time provides a seal between the shell side and the tube side of the exchanger.The device of this invention can be employed in single-pass as well as multiple-pass heat exchangers. It offers simplicity of installation as well as disassemblement of the exchanger and is less costly than the heretofore common types of installations.What is claimed is:1.In a floating-head heat exchanger having a flanged shelland a flanged head cover for said shell providingtherewith an annular recess at the juncture of the shelland cover, a tube bundle in said shell, a floating tubesheet slidably supporting one end of said tube bundleand having an annular surface facing said recess, a fluidtight sealing means disposed in said annular recesscomprising a pair of sealing ring members, spring meansbetween said sealing ring menbers resiliently biasingthe same apart and into engagement with the end wallsof said annular recess, said sealing ring members eachbeing in sealing engagement with the annular surface ofsaid tube sheet and the bottom of said recess andproviding a seal between the same and the shell and said spring means permitting said sealing rings to react resiliently in response to longitudinal movement of said floating tube sheet.2.In a floating-head heat exchanger having a flanged shelland a flanged head cover for said shell providingtherewith an annular recess at the juncture of the shell and cover, a tube bundle in said shell, a floating tube sheet slidably supporting one end of said tube bundle and having an annular surface facing said recess, a fluid-tight sealing means disposed in said recesscomprising a pair of oppositely disposed spaced rings, resilient means mounted by and between said rings urging the same axially apart, packing members between each of said rings and the adjacent surfaces of said recess ,said packing members being of slightly larger diameter than the rings and bearing on the floating tube sheet and the opposed suefaces of said recess to provide a seal between the tube sheet and the shell and to provide a seal between the tube sheet and the shell and to permit said sealing members to react resiliently responsive to longitudinalmovement of said floating tube sheet.。

换热器工艺设计专业英语Heat Exchanger Process Design: A Technical Insight.Heat exchangers are critical components in various industrial processes, facilitating the efficient transfer of heat between two or more fluids. The design of these exchangers is a complex task that requires a thorough understanding of thermodynamics, fluid dynamics, and materials science. This article delves into the intricacies of heat exchanger process design, highlighting the key considerations and challenges involved.1. Understanding the Basics.Before delving into the design aspects, it's crucial to grasp the fundamental principles of heat exchange. Heat transfer occurs due to a temperature difference between two systems. In a heat exchanger, this difference is exploited to transfer thermal energy from one fluid (the hot fluid) to another (the cold fluid). The efficiency of the heattransfer depends on several factors, including the type of heat exchanger, the flow rates and temperatures of the fluids, and the properties of the heat transfer surface.2. Types of Heat Exchangers.There are several types of heat exchangers, each suitable for different applications. Some common types include:Shell and Tube Heat Exchangers: These are the most widely used, consisting of a shell with multiple tubes inside. The hot and cold fluids flow through the tubes and shell, respectively, enabling heat transfer.Plate Heat Exchangers: These use a series of thin plates to create flow channels for the fluids. They are often used in high-pressure and high-temperature applications.Finned Tube Heat Exchangers: Fins are attached to the tubes to increase the surface area and enhance heattransfer. These are commonly found in refrigeration systems.3. Design Considerations.When designing a heat exchanger, several crucialfactors must be taken into account:Material Selection: The choice of material for theheat exchanger is crucial, as it must withstand the operating temperatures and pressures while maintaining good heat transfer properties. Materials like copper, stainless steel, and titanium are commonly used in heat exchangers.Thermal Efficiency: The heat exchanger must bedesigned to maximize thermal efficiency, ensuring maximum heat transfer from the hot fluid to the cold fluid. This involves optimizing the flow rates, temperatures, and heat transfer surface area.Pressure Drop: The design should minimize pressuredrop across the exchanger, ensuring smooth fluid flow and reduced energy losses.Maintenance and Cleaning: Considerations for easy access and cleaning are essential to maintain the heat exchanger's performance over time.4. Challenges in Heat Exchanger Design.Designing an efficient heat exchanger presents several challenges:Balancing Thermal and Mechanical Loads: The design must strike a balance between providing sufficient heat transfer surface area and ensuring structural integrity.Dealing with Fouling: Over time, deposits can accumulate on the heat transfer surface, reducing heat transfer efficiency. Designing for easy cleaning and maintenance is crucial.Optimizing for Multiple Fluids: In complex systems, heat exchangers may need to handle multiple fluids with different properties and flow rates. Designing for optimalheat transfer in such cases can be challenging.5. Conclusion.Heat exchanger process design is a complex task that requires a deep understanding of thermodynamics, fluid dynamics, and materials science. By considering the type of heat exchanger, material selection, thermal efficiency, pressure drop, and maintenance requirements, designers can create efficient and reliable heat exchangers that meet the specific needs of industrial applications. Addressing the challenges in heat exchanger design, such as balancing thermal and mechanical loads, dealing with fouling, and optimizing for multiple fluids, is key to ensuring optimal performance over time.。

International Journal of Thermal Sciences46(2007)1311–1317/locate/ijtsPerformance analysis offinned tube and unbaffled shell-and-tubeheat exchangersJoydeep Barman,A.K.Ghoshal∗Department of Chemical Engineering,Indian Institute of Technology,Guwahati,North Guwahati781039,Assam,IndiaReceived15May2006;received in revised form26August2006;accepted6December2006Available online5February2007AbstractThis work considers an optimum design problem for the different constraints involved in the designing of a shell-and-tube heat exchanger consisting of longitudinallyfinned tubes.A Matlab simulation has been employed using the Kern’s method of design of extended surface heat exchanger to determine the behavior on varying the values of the constraints and studying the overall behavior of the heat exchanger with their variation for both cases of triangular and square pitch arrangements,along with the values of pressure drop.It was found out that an optimum fin height existed for particular values of shell and tube diameters when the heat transfer rate was the maximum.Moreover it was found out that the optimumfin height increased linearly with the increase in tube outer diameter.Further studies were also performed with the variation of other important heat exchanger design features and their effects were studied on the behavior of overall performance of the shell-and-tube heat exchanger.The results were thereby summarized which would proclaim to the best performance of the heat exchanger and therefore capable of giving a good idea to the designer about the dimensional characteristics to be used for designing of a particular shell and tube heat exchanger.©2007Elsevier Masson SAS.All rights reserved.Keywords:Fin height;Heat exchanger;Heat transfer rate;Longitudinalfins;Number of tube side passes;Pressure drop;Tube pitch layout1.IntroductionFins have long been recognized as effective means to aug-ment heat transfer.The literature on this subject is sizeable. Shell and tube heat exchanger with its tube eitherfinned or bare is extensively taught in the undergraduate level.Several text and reference books deal with the problems of longitudinalfinned tube in a shell and tube heat exchanger[1–4].It is well under-stood that with increase infin height of a longitudinalfin,heat transfer area increases to increase the heat transfer and at the same time the driving force decreases to decrease the heat trans-fer.However,one important design aspect,which probably is not discussed,is presented here.For a particular shell diameter, capacity of tube numbers is decided depending on tube size and pitch arrangement.In case offinned tube,height of thefin also plays an important role.Therefore,with increase infin height though surface area increases but number of tubes as well as *Corresponding author.E-mail addresses:joydeepb@iitg.ernet.in(J.Barman),aloke@iitg.ernet.in (A.K.Ghoshal).efficiency of thefin decreases.So,there might be an optimum condition of tube number andfin height for a particular tube arrangement and a particular shell diameter for which the heat transfer rate is the maximum[5].A Matlab coding has been de-signed to study the behavior of the overall performance of a heat exchanger on varying the important design features involved in it.The important constraints involved in the designing of a heat exchanger are studied here using the Matlab program.Several results and optimum conditions related to them are briefed out and tabulated in this literature to give a basic idea to the de-signer about the requirements and limitations to be included while designing afinned tube and unbaffled shell-and-tube heat exchanger.In the present article,Kern’s method of design[2]of ex-tended surface heat exchanger is applied for a shell-and-tube heat exchanger problem.Optimum conditions offin height and number of tubes in cases of triangular pitch and square pitch arrangements are found out along with the values of pressure drop.Other results concerning the various constraints of a heat exchanger like number of passes,tube outer diameter and tube pitch layout were also studied and compared in this literature.1290-0729/$–see front matter©2007Elsevier Masson SAS.All rights reserved. doi:10.1016/j.ijthermalsci.2006.12.0051312J.Barman,A.K.Ghoshal/International Journal of Thermal Sciences46(2007)1311–1317 Nomenclaturea s,a tfluidflow area.............................m2 A o,A i tube surface area...........................m2c s,c t specific heat capacity...............J kg−1K−1d thickness of eachfin.........................m de equivalent diameter for heat transfercalculations................................m D,D1inner and outer diameter of tube..............m D2inner diameter of shell.......................m D b tube bundle diameter........................m De s equivalent diameter for pressure dropcalculations................................m f s,f t friction factor offluidh f heat transfer coefficient offins......W m−2K−1 h f i heat transfer coefficient of outside tube surface andfins with respect to the inner tubesurface............................W m−2K−1 h i heat transfer coefficient of inside tubesurface............................W m−2K−1 H f height of eachfin...........................m G s,G t mass velocity offluid...............kg m−2s−1 K s,K t thermal conductivity...............W m−1K−1 L length of each tube..........................m n number of tube side passes N f number offins per tubeN T total number of tubesP t tube pitch..................................m P w wetted perimeter............................m Pr s,Pr t Prandtl numberQ overall heat transfer rate per unit LMTD..W K−1 Re s,Re t Reynolds numbers s,s t specific gravity offluidU overall heat transfer coefficient......W m−2K−1 w tube sidefluid’s massflow rate...........kg s−1 W shell sidefluid’s massflow rate...........kg s−1 P s, P t pressure drop............................Pa Greek symbolsμs,μt,μw viscosity..............................Pa s ηffin efficiencySubscriptsffini inside of tubeo outside of tubes shell sidet tube sidew wall2.The mathematical program model of Kern’s method and solution procedure:determination of tube bundle diameter and maximum number of tubesA shell-and-tube heat exchanger with an internal shell diam-eter,D2,consisting offinned tubes of outer diameter,D1,inner diameter,D,length,L,withfins of height,H f,and thickness, d,is considered here.Total number offins per tube is N f and total number of tubes is N T.Tube bundle diameter is D b:D b=(D1+2H f)×(N T/K)(1/M)(1) The constants,K and M,are determined from Table1for dif-ferent tube passes and tube pitch layouts for a tube pitch,P t=1.25(D1+2H f)[1](2) Tube bundle diameter isfirst calculated by iterative process for bare tubes;henceforth maximum number offinned tubes,N T,is calculated from the derived tube bundle diameter from Eq.(1).3.Shell side calculationsTheflow area,a s,wetted perimeter,P w,equivalent diame-ter,d e,mass velocity,G s,Reynold’s number,Re s and Prandtl number,Pr s are calculated using Eqs.(3)–(8)as follows:a s=πD22/4−N TπD21/4+N f×d×H f(3) P w=N T(πD1−N f×d+2N f×H f)(4) d e=4a s/P w(5) G s=W/a s(6) Re s=d e×G s/μs(7) Pr s=c s×μs/K s(8) The heat transfer coefficient for the outside tube andfin surfaces can be calculated using Sieder–Tate correlation[4],Eqs.(9)and(10)as shown below:h f=1.86(K s/d e)×(Re s×Pr s×d e/L)1/3(9) for laminarflow;Table1Values of constants,K and M[1]Triangular pitch Square pitchNo.of passes1246812468K0.3190.2490.1750.07430.03650.2150.1560.1580.04020.0331 M 2.142 2.207 2.285 2.499 2.675 2.207 2.291 2.263 2.617 2.643J.Barman,A.K.Ghoshal /International Journal of Thermal Sciences 46(2007)1311–13171313h f =0.027(K s /d e )×Re 0.8s ×Pr 1/3s×(μs /μw )0.14(10)for turbulent flow.4.Tube side calculationsThe tube side flow area,a t ,mass velocity,G t ,Reynolds number,Re t and Prandtl number,Pr t ,for tube side fluid are calculated from Eqs.(11)–(14).With the values of viscosity,μt ,specific heat capacity,c t and thermal conductivity,K t ,for tube side fluid and using the Sieder–Tate correlation,the heat transfer coefficient of inside tube surface,h i ,can be calculated.a t =πN T ×D 2 /4n (11)G t =w/a t (12)Re t =DG t /μt (13)Pr t =c t ×μt /K t(14)5.Fin efficiency calculationsThe process is assumed as a steady state one and there isa continuous flow of fluid in the axial direction (both in the shell and tube side).Therefore,for a particular value of radial location,the temperature for any location in the axial direction would be almost same.Further,the angular directional variation of temperature is also neglected.Thus,the problem is reduced to a one-dimensional heat conduction problem.Hence,the fin efficiency is represented as ηf and calculated using Eqs.(15)and (16).ηf =tanh (mH f )/(mH f )(15)wherem =(2h f /K f d)1/2(16)6.Heat transfer calculationsHeat transfer coefficient of outside surface and fins with respect to the inner surface of tubes,h f i and heat transfer coef-ficient of inside surface,h i ,are given as below using Eqs.(17),(21)and (22):h f i =(H f ×P ×N f ×ηf ×N T +A o )h f /A i(17)A o and A i are the outside bare tube surface area and inside surface area of tubes respectively,where P is the perimeter of a fin,as given by Eqs.(18)–(20).A o =(πD 1−N f d)×N T L (18)A i =πDN T L (19)P =2(L +d)(20)The heat transfer coefficient for the inside tube surface can be calculated using Sieder–Tate correlation [4],Eqs.(21)and (22)as shown below:h i =1.86(K t /D)×(Re t ×Pr t ×D/L)1/3(21)for laminar flow;h i =0.027(K t /D)×Re 0.8t ×Pr 1/3t×(μt /μw )0.14(22)for turbulent flow.Thus the overall heat transfer coefficient,U ,with respect to the inside tube surface is given by Eq.(23):U =(h f i ×h i )/(h f i +h i )(23)Finally,the heat transfer rate with respect to the inside tube sur-face area,Q per degree LMTD is calculated using Eq.(24)as follows:Q =U ×A i(24)7.Pressure drop calculationsEquivalent diameter for pressure drop calculations in case of shell side fluid will be different from the diameter used for heat transfer calculations.This diameter is given by Eq.(25):De s =4a s /(P w +πD 2)(25)The pressure drops for shell side and tube side fluid, P s and P t respectively are calculated using Eqs.(26)–(29)as follows:P s = f s ×G 2s ×L / 5.22×1010×De s ×s s (26)f s =16/Re s(27)for laminar flow andf s =0.0035+0.24/Re 0.42s(28)for turbulent flow.Here,Re s =(De s ×G s )/μsP t = f t ×G 2t×L ×n / 5.22×1010×D ×s t (29)where f t is the tube side friction factor and can be calculated as shown above,Eqs.(27)and (28),using tube side Reynolds number,Re t .s s and s t are the specific gravities of shell side and tube side fluids respectively [2].8.Solution basisAn exemplary problem discussed below is used to study the objectives as discussed.Hot fluid (3.8kg s −1)in shell-side is to be cooled by a cold fluid (6.4kg s −1)in tube side.Inner di-ameter of the shell and length of the shell are kept constant as 0.5and 4.88m respectively.Inner and outer diameters of the tube are varied.Number of fins with thickness 9×10−4m per tube is 20and is kept constant for all the calculations.Ther-mal conductivity of the fin material is 45W m −1K −1.Hot and cold fluids are oxygen gas and water respectively.The values for thermal conductivity,viscosity and heat capacity of oxygen gas and water are calculated at an average temperature of 353and 305K respectively.9.Results and discussionsThe Kern’s method of designing of shell and tube heat ex-changers with extended surfaces was used for the designing of the heat exchanger concerned in this paper.The equations in-volved in this method are all simple and well established,and1314J.Barman,A.K.Ghoshal /International Journal of Thermal Sciences 46(2007)1311–1317were incorporated in a Matlab program specially coded for the purpose of this paper.This program is simply a step-wise cal-culation and does not involve any iteration or any optimization technique that may lead to some numerical errors.However,the program was thoroughly checked and thereafter run to arrive at the reasonable conclusions as reported in the manuscript.The results tabulated in Tables 2–5,and results shown graph-ically in Figs.1and 2were found out for a tube outer diameter of 0.0254m.Tables 2and 3present the maximum number of finned tubes of different fin heights for triangular pitch and square pitch arrangements respectively,which can be accom-modated in the shell of inner diameter 0.5m.They also reflect the obvious nature of variations of the shell-side and tube-side pressure drops with variation of fin height keeping one tube pass only.It is well understood that as the number of tubes decreases with the increase in fin height,the tube side fluid flow area is decreased thereby increasing the pressure drop.On the other hand,the shell side flow area increases leading to decrease in pressure drop,which is also shown through Figs.1and 2for tri-angular pitch and square pitch arrangements respectively.The variations of the heat transfer rates for both the pitches with variations of fin height are reported in Tables 2and 3respec-Fig.1.Variation of heat transfer rate and shell-side pressure drop with increase in fin height for triangular pitch arrangement,one tube side pass and for tube outer diameter,0.0254m.tively.The nature of the variations is shown through Figs.1and 2for triangular pitch and square pitch arrangements respec-tively.It is observed from the figures that there exists an opti-mum fin height (0.4572×10−2m for triangular pitch and 0.4826×10−2m for square pitch arrangement),which gives the highest heat transfer rate.Corresponding to these optimum fin heights,optimum number of adjustable finned tubes is 78and 60respectively.Under these optimum conditions,heat transfer rates are 7798.4and 5843.0W K −1,tube side pressure drops are 0.2985and 0.4723kPa and shell side pressure drops are 1.3217and 0.8343kPa for triangular and square pitch arrange-ments respectively.Tables 4and 5show the corresponding values of optimum fin height,total number of tubes,heat transfer rate and pres-sure drop for different values of tube side passes.We notice from these tables (Tables 4and 5)that for a constant shell inner diameter,with increase in the number of tube-side pass the maximum heat transfer rate corresponding to the optimum value of fin height decreases.It is also noticed that as the total number of tubes decreases the tube side pressure drop values in-creases largely which is a major drawback from economicandFig.2.Variation of heat transfer rate and shell-side pressure drop with increase in fin height for square pitch arrangement,one tube side pass and for tube outer diameter,0.0254m.Table 2Capacity of finned tubes of 0.0254m outer diameter in the shell,pressure drops and heat transfer rate values for triangular pitch arrangement and for one tube side passHeight of fin,H f ×102,m 0.2540.3810.43180.45720.5080.55880.6350.762Total number of tubes,N T10286817873696355Shell side—pressure drop, P s ,kPa 1.4341 1.3692 1.3383 1.3217 1.2893 1.2569 1.2093 1.1335Tube side—pressure drop, P t ,kPa0.18820.2530.28270.29850.3330.36950.42950.546Heat transfer rate per unit LMTD,Q ,W K −17469.37767.17797.37798.47777.47730.57622.47371.2Table 3Capacity of finned tubes of 0.0254m outer diameter in the shell,pressure drops and heat transfer rate values for square pitch arrangement and for one tube side pass Height of fin,H f ×102,m 0.2540.3810.40640.43180.45720.48260.5080.5334Total number of tubes,N T8268666462605850Shell side—pressure drop, P s ,kPa 0.87630.86050.85430.8480.84110.83430.82740.8191Tube side—pressure drop, P t ,kPa0.27650.37570.39850.4220.44610.47230.49850.5268Heat transfer rate per unit LMTD,Q ,W K −15522.45797.55820.45835.15842.45843.05837.65826.8J.Barman,A.K.Ghoshal /International Journal of Thermal Sciences 46(2007)1311–13171315optimization point of views.As expected,the shell side pressure drop decreases with decrease in tube number but the decrease is much less in comparison to the increase for the tube side pressure drop.So,in this case,the tube side pressure drop val-ues bear more importance while selecting the number of passes.Hence,from the tabulated data obtained it can be said that one tube side pass is the best choice for the finest results of heat ex-changer performance unless a constraint related to the number of tubes is faced when higher values of tube side pass could be considered.Moreover,it was also noticed that for a particular fin height,the total number of adjustable tubes varies for the pitch arrangements.As the number of tubes that could be ad-justed in a square pitch arrangement were less in number than in triangular pitch arrangement so even the most optimum value of fin height in case of square pitch arrangement could not pro-duce the same heat transfer rate as compared to the other.But the shell side pressure drop is higher in magnitude in triangu-lar pitch than in square pitch arrangement,whereas the relation is just the opposite in case of tube side pressure drop values.So,in the absence of any pressure drop constraints,thetriangu-Fig.3.Variation of optimum fin height with outer diameter of tubes.lar pitch arrangement with the optimum value of fin height will prove to be the best choice.The other tables,i.e.,Tables 6–9give the values of different important parameters such as a s ,A i ,A o ,Re s ,Pr s ,h f ,ηf ,h f i ,h i and U used and determined during the calculations.Fig.3shows the variation of optimum fin height with the change of tube outer diameters for a fixed number of tube side passes and for triangular pitch arrangement.The relation between them is found to be linear and can be expressed by Eq.(30):H f =0.0852×D 1+0.0025(30)Thus by using this equation,an approximate value of optimum fin height for the highest heat transfer rate can be pre-calculated for a tube of particular diameter.Fig.4shows a comparison of the performance of the heat exchanger for one and two tube side passes for triangular pitch arrangement.It is found out that the performance of the heat exchanger based on the heat transfer rate values for two-tube side passes could never meet up with the results for one tube side pass.Also after inspecting thepres-parison of heat transfer rates with fin heights for one and two tubes side passes.Table 4Optimum fin height for maximum heat transfer rate,for different tube-side passes and corresponding values of total number of finned tubes and pressure drops for triangular pitch arrangement and for tube outer diameter as 0.0254m Number of tube side passes,n 12468Optimum fin height,H f ×102,m0.45720.45720.45720.43180.38Heat transfer rate per unit LMTD,Q ,W K −17798.47639.46523.84383.53166.4Total number of finned tubes,N T 7872624740Shell-side pressure drop, P s ,kPa 1.3217 1.12040.84170.50830.3567Tube-side pressure drop, P t ,kPa0.29852.29120.032298.8108297.322Table 5Optimum fin height for maximum heat transfer rate,for different tube-side passes and corresponding values of total number of finned tubes and pressure drops for square pitch arrangement and for tube outer diameter as 0.0254m Number of tube side passes,n 12468Optimum fin height,H f ×102,m0.48260.45720.48260.40640.4064Heat transfer rate per deg.LMTD,Q ,W K −158435476.75286.92911.72503.0Total number of finned tubes,N T 6056513632Shell-side pressure drop, P s ,kPa 0.8340.70310.63640.32520.2773Tube-side pressure drop, P t ,kPa0.47233.557827.9617160.236441.3241316J.Barman,A.K.Ghoshal/International Journal of Thermal Sciences46(2007)1311–1317Table6Values of various parameters involved in determining the important variables of Table2Height offin,H f×102,m0.2540.3810.43180.45720.5080.55880.6350.762 Shell sideflow area,a s,m2 1.41 1.485 1.511 1.523 1.546 1.567 1.596 1.637 Inside tube surface area,A i,m231.326.3724.7123.9422.521.1819.4116.9 Outside tube surface area,A o,m231.10126.2124.5623.79322.3621.0519.2816.8 Shell side 3.988 3.612 3.521 3.483 3.423 3.377 3.33 3.294 Reynolds number,Re s×10−4Shell side0.7010.7010.7010.7000.7010.7010.7010.701 Prandtl number,Pr sFin heat transfer111.8108.26106.96106.34105.14104.01102.42100.05 coefficient,h f,W m−2K−1Fin efficiency,ηf0.98820.97460.9680.96450.95710.94920.93640.9133 Heat transfer coefficient of291.03365.342392.96406.34432.25457.06492.27545.82fins and outside tube surfacewith respect toinside tube surface,h f i,W m−2K−1Inside tube surface heat 1.325 1.52 1.601 1.642 1.726 1.811 1.943 2.17 transfer coefficient,h i×10−3,W m−2K−1Overall heat transfer coefficient238.63294.55315.52325.75345.68364.97392.74436.11 with respect toinside tube surface,U,W m−2K−1Table7Values of various parameters involved in determining the important variables of Table3Height offin,H f×102,m0.2540.3810.40640.43180.45720.48260.5080.5334 Shell sideflow area,a s,m2 1.532 1.595 1.606 1.6162 1.626 1.636 1.645 1.654 Inside tube surface area,A i,m225.0320.9820.28319.6218.9818.3817.8117.26 Outside tube surface area,A o,m224.8820.8520.15719.518.8718.2717.717.16 Shell side 4.987 4.54 4.484 4.434 4.392 4.355 4.323 4.296 Reynolds number,Re s×10−4Shell side0.7010.7010.7010.7010.7010.7010.7010.701 Prandtl number,P r sFin heat transfer98.3796.3295.9195.595.0994.794.393.92 coefficient,h f,W m−2K−1Fin efficiency,ηf0.98960.97730.97440.97130.96810.9650.96130.9577 Heat transfer coefficient of256.3325.67338.81351.72364.38376.81389.0400.96fins and outside tube surface withrespect to insidetube surface,h f i,W m−2K−1Inside tube surface heat 1.585 1.825 1.875 1.926 1.977 2.028 2.081 2.133 transfer coefficient,h i×10−3,W m−2K−1Overall heat transfer coefficient220.62276.36286.96297.4307.67317.78327.73337.52 with respect toinside tube surface,U,W m−2K−1sure drop values(Table4),it can be well concluded that the best option would be to select a heat exchanger with one tube side pass if there is no tube number constraint involved.Hence it can be well summarized by mentioning that a combination of triangular pitch arrangement,one tube side pass and a value of fin height calculated from Eq.(30),when incorporated in the designing of a shell-and-tube heat exchanger with no baffles would certainly proclaim to give the best performance until and unless some restriction is being levied on in terms of pressure drop or number of tubes.10.ConclusionsIn this work the variation of heat transfer rate withfin height for afinned tube shell-and-tube heat exchanger was studied for two different pitch arrangements.It was found out that for par-J.Barman,A.K.Ghoshal/International Journal of Thermal Sciences46(2007)1311–13171317 Table8Values of various parameters involved in determining the optimumfin height and other important variables of Table4Number of tube side passes,n12468 Optimumfin height,H f×102,m0.45720.45720.45720.43180.38 Shell sideflow area,a s,m2 1.523 1.5619 1.6261 1.722 1.7725 Inside tube surface area,A i,m223.9422.0818.9914.512.238 Outside tube surface area,A o,m223.79321.9418.87514.4112.163 Shell side Reynolds number,Re s×10−4 3.483 3.778 4.391 6.0017.798 Fin heat transfer coefficient,h f,W m−2K−1106.34102.0495.184.3777.781 Fin efficiency,ηf0.96450.96590.96810.97460.9817 Heat transfer coefficient offins and outside406.34390.3364.4311.5263.33 tube surface with respect to inside tubesurface,h f i,W m−2K−1Inside tube surface heat transfer coefficient, 1.642 3.051 5.992 1.0285 1.4827 h i×10−3,W m−2K−1Overall heat transfer coefficient with respect325.75346.03343.51302.34258.74 to inside tube surface,U,W m−2K−1Table9Values of various parameters involved in determining the optimumfin height and other important variables of Table5Number of tube side passes,n12468 Optimumfin height,H f×102,m0.48260.45720.48260.40640.4064 Shell sideflow area,a s,m2 1.636 1.6644 1.692 1.795 1.8206 Inside tube surface area,A i,m218.3817.1515.7111.0379.788 Outside tube surface area,A o,m218.2717.0515.61310.979.728 Shell side Reynolds number,Re s×10−4 4.355 4.826 5.0978.249.291 Fin heat transfer coefficient,h f,W m−2K−194.791.0488.7275.9673.118 Fin efficiency,ηf0.9650.96940.9670.97960.9803 Heat transfer coefficient offins and outside376.81349.19353.61269.38259.44 tube surface with respect to inside tubesurface,h f i,W m−2K−1Inside tube surface heat transfer coefficient, 2.028 3.734 6.974 1.28 1.773 h i×10−3,W m−2K−1Overall heat transfer coefficient with respect317.78319.3336.54263.82255.69 to inside tube surface,U,W m−2K−1ticular shell and tube diameters an optimum value offin height exists,which gives the highest heat transfer rate.Moreover it was also found out that on increasing the number of tube side passes while keeping the shell diameter constant,though the number of tubes could be decreased but the performance on the basis of heat transfer rate kept on decreasing and tube side pressure drop values increased substantially.The optimumfin height also increased linearly with the increase of tube outer diameter.It is worth mentioning here that the Matlab coding designed for this problem and the results obtained on using it,might prove quite beneficial in choosing the most appropriatefin height,total number of tubes,tube dimensions,arrangements, number of tube side passes andfin dimensions for a known value of shell diameter as well as keeping the pressure drops in check.In this problem the physical properties of thefluids were assumed constant,tube andfin dimensions were assumed uniform,throughout the entire system.It can be further stated that no experimental verification could be possible due to lack of such experimental data.How-ever,it would be highly appreciated to carry experimental work in this regard.References[1]J.R.Backhurst,J.M.Coulson,J.H.Harkar,J.F.Richardson,Coulson&Richardson’s Chemical Engineering,Butterworth–Heinemann,Oxford, 2004.[2]D.Q.Kern,Process Heat Transfer,McGraw-Hill,New York,2000.[3]P.Harriott,W.L.McCabe,J.C.Smith,Unit Operations of Chemical Engi-neering,McGraw-Hill,New York,2001.[4]S.P.Dusan,R.K.Shah,Fundamentals of Heat Exchanger Design,John Wi-ley and Sons,New York,2003.[5]J.Barman,A.K.Ghoshal,in:Proceedings of Chemcon’05,58th AnnualChemical Engineering Congress,India,2005.。

外文文献：Design and Implementation of Heat Exchange Station Control SystemKeywords:Heat exchange station, Control system, PLC, Inverter, Configuration software.Abstract.This paper introduces a design and implementation of heat exchange station control systembased on PLC and industrial configuration software, which includes the contr ol scheme and principle,hardware selection and software design, etc. The circulating pumps and re plenishing pumps in thesystem can all be driven automatically by PLC and inverter. Main process parameters, such as steampressure and measurement temperature and so on,can all be shown on the industrial PC runningconfiguration software, and instructions could be sent by the engineer and operator on-the-spot via theHuman Machine Interface as well. The automatic pressures adjustment of stea m supply of the heaterby advanced PID algorithm has been realized finally. It is verified that the system is highly reliableand stable, and it greatly enhances the level of automation and pressure control accuracy of the heatexchange station and meets all the equipments running demands well. IntroductionWith the rapid development of economy and society, heat supply systems are the key power source inthe communities and plants in China. As a media between heat sources and heat loads in the systems,a heat exchange stations plays a very important role for the heat supplyquality. Traditionally, most ofthe pumps in the heat supply systems are operated by valves manually, s o it could bring about thepower energy consuming, high labor intensity and low operation automation. I n this paper a design ofcontrol system for heat exchange station based on PLC, inverter and indust rial configuration softwarewas proposed,accordingly the aim for power energy saving,high heat efficiency and operationautomation has been achieved.Process outline and Control demandsProcess outline.The process outline and control demands were put forward at first before the schemeand design of heat exchange station control system were proposed.Heat exchange station consists of a steam-driven heater,plus3ci rculating pumps,2replenishingpumps and electric control valve. By adjusting the steam flux into the mixture of water and steamaccording to the temperature sensors mounted indoors and outdoors, the pr ocess of heat exchangecould be completed. Among these equipments, the steam-driven heater, a heat exchanger containingmixture of steam-and-water, is the key appliance for heat supply system.Control demands.Major control demands for the control system were listed a s follows [1]:(a)Pumps driving.Pumps include3circulating pumps(2in operation,1for backup)and2replenishing pumps (1 in operation, 1 for backup). Among circulating ones one is driven by powerfrequency, the others are driven by variable frequency, with 75KW power ea ch; among replenishingones one is driven by power frequency, the other is driven by variable frequency, with 3KW powereach. The control signal should be originated from the pressure difference between the supply waterand return water.Pumps could be driven in stepless speed regulating when connecting variablepower;(b)Parameters Showing.The showing parameters contain temperatureshowing-temperature ofsupply water, return water, the indoor, the outdoor and steam - and pre ssure showing - pressure ofsupply water, return water and pre-valve and post-valve of the steam etc;(c) Butterfly valves driving.Two butterfly valves can be on or of f automatically when the wholesystem start or stop;(d) Motor-driven valves control. By continuously adjusting the opening of t he valves according to thesignal from the temperature sensors indoors and outdoors, the supply wate r temperature should bestabilized in the presetting values;(e) HMI (Human Machine Interface) Demands. The process flow chart of heatexchange station andmain process parameter can be shown in HMI, and instructions can be trans mitted via this interface;(f) Safeguard Function. The circulating pumps should be out of running when heat exchange system isin water needing, and steam should be kept out of the heater when the pumps are not revolving.Hardware Selection of the Control SystemFrom the control demands mentioned above, the controller of the control s ystem can process signalsboth relay and analog, having the ability of loop adjustment of analog q uantity. At the meantime thepumps could run in the working condition of variable frequency, so the hardware selection of thecontrol system for heat exchange station should be made deliberately.PLC Serving as Main Controller.As some experienced electrica l engineers known,PLC/PC(Program Controller) is a kind of popular industrial computer, and it can not only accomplish logiccontrol, but also complete many advanced functions, such as analog quanti ty loop adjustment, andmotion control, etc. According to the component amounts of input and outpu t and the needs of controlsystem, FX1N-60MR micro PLC of MITSUBISHI FX series is selected, which hav ing 36 inputs and24 outputs, and doing analog adjustment by using advanced instruction likePID instruction [2].Because of the sampling and driving of the analog signal necessarily, P LC should be extended toanalog input/output function module like FX2N-4AD (4AD) and FX2N-4DA (4DA) or somethinglike.On one hand,4AD adopted is an analog input module having4channels with12bit highresolution, which could receive 0~+10V voltage signal, 0~20mA or 4~20mA cu rrent signal. On theother hand, 4DA chosen could send standard voltage signal and/or current signal, having 4 channelswith 12 bit high resolution also. It is something to be mentioned here , the wiring form of currentinput/output (4~20mA) must be adopted in order to avoiding the strong elec tromagnetism disturbancein the working field [3].Inverter completing Stepless Speed Regulating.At present, inverter, as an im port power electronicconverter, can convert constantly power frequency into continually variable frequency. Thus, energysaving, cost consuming and noise reduction can be easily reached by this equipment.In this control system inverter of ACS510 series of ABB Corporation were elaborately chosen, whichhas many advantages, such as Direct Torque Control (DTC) and advanced appl ying macro and so on.Its main good points and characteristics are illustrated as follows: it can acquire maximum startingtorque (200% normal torque) by using direct excitation; it can be applied to multiple driving systemsby using master-slave function;input and output programmable function;high precision of speedregulating, perfect safeguard and alarming steps. Owing to these highlights of this inverter, pumpsdriving of stepless speed regulating can be easily obtained.There are many applying macro inACS510 series, but we should only choose manual/automatic macro here as we need.IPC Acting as Monitor&Control Interface.IPC(Industrial Personal Computer)has strongcompatibility,extensibility and reliability,which can connect PLC by RS-232serial portconveniently. In the hardware configuration we select IPC H610 series of A DVANTECH as HMI.MCGS(Monitor Control Generated System),fashionable home-mad e industrial configurationsoftware, is running on the ADVANTECH IPC. Using this HMI, the visualizati on process of Monitorand Control is realized easily, intuitively and vividly.With the sensor/transducer,analog input/output modules,PLC a nd actuators,.inverter andmotor-driven valve, the loop adjustment of steam pressure can be precisely attained, and temperatureof all measure points could be measured also[4].The overall hardware configuration of heatexchange station control system see Fig. 1.Fig. 1 The overall hardware configuration of heat exchange station control systemSoftware Design of the Control SystemLAD Diagram Programming.Out of the thoughts of modular programming, the whole programstructure can be divided into such several modules as Initialization Function,upper IPCCommunication Function, Relay Control Function, Analog Sampling, Fuzzy PID Adjust Functionand Safeguard Function, etc. The flow chart of LAD diagram programming of PLC is shown in Fig. 2.Among these modular functions, it is something worthy to mention of Fuzzy PID Adjust Function.Under some circumstances the using of PID instruction of PLC was not so good at what we expected;therefore, the self-made program of Fuzzy PID adjustment of steam pressure was done from deviationand deviation acceleration of temperature between the indoor and the outdoo r in accordance with theFuzzy Control Theory and its application [5].HMI Configuration.For the sake of the appearance beauty and personalizati on between machineand human, the MCGS- Monitor Control Generated Software of Beijing MCGS Tech Co. Ltd wasadopted. This industrial configuration software has very quick, easy devel opment of configurationprocess, which can build bi-directional and high speed communication betwee n PLC and upper IPCthru RS422/232 serial port.In the development environment of MCGS, all needed windows and pictures we re created, includingMain Window of Process Flow, Process Parameters Showing Window, and Key P arameters SettingWindow, etc. Vivid and readily interaction between human and machine can be completed by suchbeautiful pictures and animations when IPC running MCGS.ConclusionsThis design of heat exchange station control system based onFX series PLC,MCGS,and ABBinverter has been realized the pressure automatic adjustment of steam-driven heater as originallyexpected.More over,design demands of power energy savi ng,high heat efficiency and lowequipments noise can all be well met. Finally, the practical operation ver ifies that the system is highlyreliable and stable, and it greatly enhances the level of automation and pressure control accuracy ofheat exchange station and meets equipments requirements of energy saving an d green driving.BEGINInitializationFuzzy PIDAdjust FunctionCommunicationFunctionAnalog OutputNoRelay ControlFunctionAnalog FilteringFunctionCall AnalogSample FunctionSample OverYesAnalog InputLinear TransferLinear TransferAnalog OutputDrivingSafeguard FunctionFailure OccurNoYesFailure HandlingRelated MemoryResetENDFig. 2 The flow chat of LAD diagram programming of PLCAcknowledgementComposition of this paper was with the help and under the direction ofSenior Engineer Nian-huiZhang of Qingdao Wellborn Automation Corporation.References[1]Information on H. Zhang, . Li:The Principle of PLC with itsApplications to Process Control(China PowerPress, Beijing 2008).[3]H. Zhang:The Design and Development of MITSUBISHI FX Series PLC( China Machine Press,Beijing 2009).[4]H. Zhang: Process Automation Instrumentation, Vol. 31(4) (2010), p. 34-36, in Chinese.[5]. Zadeh:Fuzzy Sets and their Applications(Academic Press, New Yor k 1975).Progress in Civil Engineeringand Implementation of Heat Exchange Station Control System外文翻译：换热站控制系统的设计和实现关键词:换热站、控制系统、PLC、变频器、配置软件。

One of the most effective methods of increasing the rate of heat transfer in heat exchangersis using tubes with lengthwise corrugations (Fig. i). Among the different methodsknown here and abroadfor making such tubes (longitudinal and rotary rolling, welding, drawing,forging), cold drawing occupies an important position. This is because of the highproductivity of this method, the accuracy of the tube dimensions, the good surface finish,and the fact that the tool is relatively simple to fabricate. A technology has been developedand introduced at several nonferrous metallurgical plants for drawing copper-alloy tubeswith lengthwise corrugations.The Pervouralsk New Tube Plant is developing a technology for drawing such tubes madeof carbon steel. Trial lots of tubes with a corrugated outer surface have been made andstudies are being conducted to determine the optimum geometry of the die.There are certain distinctive features of drawing stainless steel tubes that owe to theproperties of the material. The cold working of alloy steels -- thus, stainless steels -- ischaracterized by a high susceptibility to work hardening, low thermal conductivity, and thepresence of a hard and strong film on the surface which is passive to lubricants. The presenceof the film leads to seizing of the tube in the die. Existing lubricants and prelubricantcoatings do not provide a plasticized layer that will prevent the metal from adheringto the die and ensure a uniform strain distribution over the tube wall thickness.The Ural Polytechnic Institute and the Institute of Electrochemistry of the Ural ScienceCenter under the Academy of Science of the USSR have developed a technology for applying acopper coating to the surface of tubes made of corrosion-resistant steels. The coating isapplied in the form of a melt containing copper salts at 400-500~ and allowed to stand fori0 min. The layer of copper 10-40 ~m thick formed on the surface by this operation is stronglybound to the base metal. The copper coating makes it possible to draw tubes of stainlesssteel on a mandrel.The sector metallurgical-equipment laboratory at the UralPolytechnic Institute studiedthe process of drawing stainless steel tubes using the copper coating on short(stationary)and long (movable) mandrels. The study showed that the metal does not adhere to the die, thecoating is strongly bound to the base metal, and large reductions can be made in one pass.These results suggested that stainless-steel tubes with lengthwise corrugations could beproduced by cold drawing. Thus, the laboratory prepared trial lots of corrugated steel12KhI8NIOT tubes.其中一个最有效的办法来增加率换热器的传热利用管corrugations与纵向(Fig.我)。

管壳式换热器强化传热研究摘要：从管程强化和壳程强化两方面论述了管壳式换热器强化传热技术的机理，指出了管壳式换热器今后发展中的主要方向;同时对换热器的防腐措施以及改进动向作了介绍。

关键词：强化传热;管壳式换热器;防腐Abstract: shell and tube heat exchanger was discussed from two aspects of the strengthening of the tube side and the strengthening of the shell to strengthen the mechanism of heat transfer technology, pointing out that the main direction of future development of the shell and tube heat exchanger; heat exchanger anti-corrosion measures well as improved trends were introduced. Keywords: heat transfer enhancement; shell and tube heat exchanger; anti-corrosion引言管壳式换热器是当今应用最广泛的换热设备，它具有高的可靠性和简单易用性。

特别是在较高参数的工况条件下，管壳式更显示了其独有的长处“目前在提高该类换热器性能所开展的研究主要是强化传热，适应高参数和各类有腐蚀介质的耐腐材料以及为大型化的发展所作的结构改进。

一、换热器的强化传热研究换热器的强化传热就是采用一定的措施增大换热设备的传热速率，力图用较少的传热面积或体积的设备来完成传热任务。

各种强化型换热器在石油、化工、制冷、航空、车辆、动力机械等工业部门己得到广泛应用。

强化传热已被学术界称为第二代传热技术。

应用计算数值的方法来研究流体的粘度变化对板式换热器性能的影响M.A. Mehrabian and M. KhoramabadiDepartment of Mechanical Engineering, Shahid Bahonar University of Kerman,Kerman, Iran摘要目的--本文的目的是在逆流和稳态条件下，通过数值计算，研究流体粘度的变化对板式换热器热特性的影响。

设计/工艺/方法--实现这篇文章目的的方法，源于由4部分组成的热量交换板中间通道中冷热流体的一维能量平衡方程。

有限差分法已经用于计算温度分布及换热器的热性能。

在侧边通道中，水作为将被冷却的热流体，然而在中央通道中，大量随温度变化同时粘度随之变化剧烈的流体作为将要被加热的冷流体。

发现—这个程序的运行实现了工作流体的结合，例如水与水，水与异辛烷，水与苯，水与甘油和水与汽油等。

对于以上所有工作流体的结合，两种流体的温度分布已经沿流动通道划分。

总传热系数可以通过冷流体和热流体的温度来绘制。

研究发现，若总传热系数呈线性变化，在温度变化范围内既不是冷流体和热流体的温度。

当粘度已受温度影响或者冷流体的性质改变时，换热器的影响效果并不是很显著。

创意/价值--对于由2块板为边界的温度控制体来说，本文包含一个可以得到能量平衡方程数值解的新方法。

通过对数值计算结果与实验结果进行比较，验证了这种数值计算方法。

关键词:热交换器、热传递、数值分析、有限差分法研究类型:研究性论文。

术 语2:m A 板传热面积，m b 板间距，:等式常数:CC ︒W/:C 热容，C kg J C p ︒⋅/:定压比热容，m D e 当量直径，:Cm W h ︒⋅2/:对流传热系数， 指定轴截面:jC m W k ︒⋅/:板传导率，m L 板长度，:粘度修正系数:ms kg m /:质量流量，•之间的斜率与e r R NuP n 31:-NTU: 传热单元数Nu: 努塞尔数Pr: 普朗特数Q: 传热速率, WRe: 雷诺数r: 方程指数 (8)t: 时间, sT: 温度, ℃u: 流速, m/sC m W U ︒⋅2/：总传热系数，C m W U ︒-⋅2/:平均传热系数，3:m V 通道体积，w: 流动宽度, mx: 横向坐标y: 轴向坐标sm kg m ⋅/:流体动粘度系数， 3/:m kg r 流体密度，l: 换热器有效性d: 板厚度, mf: 板投影面积的比值下标c : 冷流体Cv: 控制体h : 热流体m : 平均值min:最小值w : 板壁介绍板式换热器在不同产业发展进程中的贡献日益增加。

毕业设计（论文）外文文献翻译文献、资料中文题目：U形管换热器文献、资料英文题目：文献、资料来源：文献、资料发表（出版）日期：院（部）：专业：过程装备与控制工程专业班级：姓名：学号：指导教师：翻译日期： 2017.02.14毕业设计（论文）外文翻译毕业设计（论文）题目： U形管式换热器设计外文题目： U-tube heat exchangers译文题目：指导教师评阅意见U-tube heat exchangersM. Spiga and G. Spiga, Bologna1 Summary:Some analytical solutions are provided to predict the steady temperature distributions of both fluids in U-tube heat exchangers. The energy equations are solved assuming that the fluids remain unmixed and single-phased. The analytical predictions are compared with the design data and the numerical results concerning the heat exchanger of a spent nuclear fuel pool plant, assuming distinctly full mixing and no mixing conditions for the secondary fluid (shell side). The investigation is carried out by studying the influence of all the usual dimensionless parameters (flow capacitance ratio, heat transfer resistance ratio and number of transfer units), to get an immediate and significant insight into the thermal behaviour of the heat Exchanger.More detailed and accurate studies about the knowledge of the fluid temperature distribution inside heat exchangers are greatly required nowadays. This is needed to provide correct evaluation of thermal and structural performances, mainly in the industrial fields (such as nuclear engineering) where larger, more efficient and reliable units are sought, and where a good thermal design can not leave integrity and safety requirements out of consideration [1--3]. In this view, the huge amount of scientific and technical informations available in several texts [4, 5], mainly concerning charts and maps useful for exit temperatures and effectiveness considerations, are not quite satisfactory for a more rigorous and local analysis. In fact the investigation of the thermomechanieal behaviour (thermal stresses, plasticity, creep, fracture mechanics) of tubes, plates, fins and structural components in the heat exchanger insists on the temperature distribution. So it should be very useful to equip the stress analysis codes for heat exchangers withsimple analytical expressions for the temperature map (without resorting to time consuming numerical solutions for the thermal problem), allowing a sensible saving in computer costs. Analytical predictions provide the thermal map of a heat exchanger, aiding in the designoptimization.Moreover they greatly reduce the need of scale model testing (generally prohibitively expensive in nuclear engineering), and furnish an accurate benchmark for the validation of more refined numerical solutions obtained by computer codes. The purpose of this paper is to present the local bulk-wall and fluid temperature distributions forU-tube heat exchangers, solving analytically the energy balance equations.122 General assumptionsLet m, c, h, and A denote mass flow rate (kg/s), specific heat (J/kg -1 K-l), heat transfer coefficient(Wm -2 K-l), and heat transfer surface (m2) for each leg, respectively. The theoretical analysis is based on classical assumptions [6] :-- steady state working conditions,-- equal flow distribution (same mass flow rate for every tube of the bundle),-- single phase fluid flow,-- constant physical properties of exchanger core and fluids,-- adiabatic exchanger shell or shroud,-- no heat conduction in the axial direction,-- constant thermal conductances hA comprehending wall resistance and fouling.According to this last assumption, the wall temperature is the same for the primary and secondary flow. However the heat transfer balance between the fluids is quite respected, since the fluid-wall conductances are appropriately reduced to account for the wall thermal resistance and thefouling factor [6]. The dimensionless parameters typical of the heat transfer phenomena between the fluids arethe flow capacitance and the heat transfer resistance ratiosand the number of transfer units, commonly labaled NTU in the literature,where (mc)min stands for the smaller of the two values (mc)sand (mc)p.In (1) the subscripts s and p refer to secondary and primary fluid, respectively. Only three of the previous five numbers are independent, in fact :The boundary conditions are the inlet temperatures of both fluids3 Parallel and counter flow solutionsThe well known monodimensional solutions for single-pass parallel and counterflow heat exchanger,which will be useful later for the analysis of U-tube heat exchangers, are presented below. If t, T,νare wall, primary fluid, and secondary fluid bulk temperatures (K), and ξ and L represent the longitudinal space coordinate and the heat exchanger length (m), the energy balance equations in dimensionless coordinate x = ξ/L, for parallel and counterflow respectivelyread asM. Spiga and G. Spiga: Temperature profiles in U-tube heat exchangersAfter some algebra, a second order differential equation is deduced for the temperature of the primary (or secondary) fluid, leading to the solutionwhere the integration constants follow from the boundary conditions T(0)=T i , ν(0)≒νifor parallel T(1) = Ti ,ν(0) = νifor counter flow. They are given-- for parallel flow by - for counterflow byWishing to give prominence to the number of transfer units, it can be noticed thatFor counterflow heat exchangers, when E = 1, the solutions (5), (6) degenerate and the fluidtemperatures are given byIt can be realized that (5) -(9) actually depend only on the two parametersE, NTU. However a formalism involving the numbers E, Ns. R has been chosen here in order to avoid the double formalism (E ≤1 and E > 1) connected to NTU.4 U-tube heat exchangerIn the primary side of the U-tube heat exchanger, whose schematic drawing is shown in Fig. 1, the hot fluid enters the inlet plenum flowing inside the tubes, and exits from the outlet plenum. In the secondary side the fluid flows in the tube bundle (shell side). This arrangement suggests that the heat exchanger can be considered as formed by the coupling of a parallel and a counter-flow heat exchanger, each with a heigth equal to the half length of the mean U-tube. However it is necessary to take into account the interactions in the secondary fluid between the hot and the cold leg, considering that the two flows are not physically separated. Two extreme opposite conditions can be investigated: no mixing and full mixing in the two streams of the secondary fluid. The actual heat transfer phenomena are certainly characterized by only a partial mixing ofthe shell side fluid between the legs, hence the analysis of these two extreme theoretical conditions will provide an upper and a lower limit for the actual temperature distribution.4.1 No mixing conditionsIn this hypothesis the U-tube heat exchanger can be modelled by two independent heat exchangers, a cocurrent heat exchanger for the hot leg and a eountercurrent heat exchanger for the cold leg. The only coupling condition is that, for the primary fluid, the inlet temperature in the cold side must be the exit temperature of the hot side. The numbers R, E, N, NTU can have different values for the two legs, because of thedifferent values of the heat transfer coefficients and physical properties. The energy balance equations are the same given in (2)--(4), where now the numbers E and Ns must be changed in E/2 and 2Ns in both legs, if we want to use in their definition the total secondary mass flow rate, since it is reduced in every leg to half the inlet mass flow rate ms. Of course it is understood that the area A to be used here is half of the total exchange area of the unit, as it occurs for the length L too. Recalling (5)--(9) and resorting to the subscripts c and h to label the cold and hot leg, respectively, the temperature profile is given bywhere the integration constants are:M. Spiga and G. Spiga: Temperature profiles in U-tube heat exchangersIf E, = 2 the solutions (13), (14) for the cold leg degenerate into4.2 Full mixing conditionsA different approach can be proposed to predict the temperature distributions in the core wall and fluids of the U-tube heat exchanger. The assumption of full mixing implies that the temperaturesof the secondary fluid in the two legs, at the same longitudinal section, are exactly coinciding. In this situation the steady state energy balance equations constitute the following differential set :The bulk wall temperature in both sides is thenand (18)--(22) are simplified to a set of three equations, whose summation gives a differential equation for the secondary fluid temperature, withgeneral solutionwhere # is an integration constant to be specified. Consequently a second order differential equation is deduced for the primary fluid temperature in the hot leg :where the numbers B, C and D are defined asThe solution to (24) allows to determine the temperaturesand the number G is defined asThe boundary conditions for the fluids i.e. provide the integration constantsAgain the fluid temperatures depend only on the numbers E and NTU.5 ResultsThe analytical solutions allow to deduce useful informations about temperature profiles and effectiveness. Concerning the U-tube heat exchanger, the solutions (10)--(15) and (25)--(27) have been used as a benchmark for the numerical predictions of a computer code [7], already validated, obtaining a very satisfactory agreement.M. Spiga and G. Spiga: Temperature profiles in U-tube heat exchangers 163 Moreover a testing has been performed considering a Shutte & Koerting Co. U-tube heat exchanger, designed for the cooling system of a spent nuclear fuel storage pool. The demineralized water of the fuel pit flows inside the tubes, the raw water in the shell side. The correct determination of the thermal resistances is very important to get a reliable prediction ; for every leg the heat transfer coefficients have been evaluated by the Bittus-Boelter correlation in the tube side [8], by the Weisman correlation in the shell side [9] ; the wall material isstainless steel AISI 304.and the circles indicate the experimental data supplied by the manufacturer. The numbers E, NTU, R for the hot and the cold leg are respectively 1.010, 0.389, 0.502 and 1.011, 0.38~, 0.520. The difference between the experimental datum and the analytical prediction of the exit temperature is 0.7％ for the primary fluid, 0.9% for the secondary fluid. The average exit temperature of the secondary fluid in the no mixing model differs from the full mixing result only by 0.6%. It is worth pointing out the relatively small differences between the profiles obtained through the two different hypotheses (full and no mixing conditions), mainly for the primary fluid; the actual temperature distribution is certainly bounded between these upper and lower limits,hence it is very well specified. Figures 3-5 report the longitudinal temperaturedistribution in the core wall, τw = (t -- νi)/(Ti -- νi), emphasizing theeffects of the parameters E, NTU, R.As above discussed this profile can be very useful for detailed stress analysis, for instance as anM. Spiga and G. Spiga: Temperature profiles in U-tube heat exchangersinput for related computer codes. In particular the thermal conditions at the U-bend transitions are responsible of a relative movement between the hot and the cold leg, producing hoop stresses with possible occurrence of tube cracking . It is evident that the cold leg is more constrained than the hot leg; the axial thermal gradient is higher in the inlet region and increases with increasing values of E, NTU, R. The heat exchanger effectiveness e, defined as the ratio of the actual heat transfer rate(mc)p (Ti-- Tout), Tout=Tc(O), to the maximum hypothetical rateunder the same conditions (mc)min (Ti- νi), is shown in Figs. 6, 7respectively versus the number of transfer units and the flow capacitance ratio. As known, the balanced heat exchangers E = 1) present the worst behaviour ; the effectiveness does not depend on R and is the same for reciprocal values of the flow capacitance ratio.U形管换热器m . Spiga和g . Spiga,博洛尼亚摘要:分析解决方案提供一些两相流体在u形管换热器中的分布情况。

DESIGN OF HEAT EXCHANGER FOR HEAT RECOVERY IN CHP SYSTEMSABSTRACTThe objective of this research is to review issues related to the design of heat recovery unit in Combined Heat and Power (CHP) systems. To meet specific needs of CHP systems, configurations can be altered to affect different factors of the design. Before the design process can begin, product specifications, such as steam or water pressures and temperatures, and equipment, such as absorption chillers and heat exchangers, need to be identified and defined. The Energy Engineering Laboratory of the Mechanical Engineering Department of the University of Louisiana at Lafayette and the Louisiana Industrial Assessment Center has been donated an 800kW diesel turbine and a 100 ton absorption chiller from industries. This equipment needs to be integrated with a heat exchanger to work as a Combined Heat and Power system for the University which will supplement the chilled water supply and electricity. The design constraints of the heat recovery unit are the specifications of the turbine and the chiller which cannot be altered.INTRODUCTIONCombined Heat and Power (CHP), also known as cogeneration, is a way to generate power and heat simultaneously and use the heat generated in the process for various purposes. While the cogenerated power in mechanical or electrical energy can be either totally consumed in an industrial plant or exported to a utility grid, the recovered heat obtained from the thermal energy in exhaust streams of power generating equipment is used to operate equipment such as absorption chillers, desiccant dehumidifiers, or heat recovery equipment for producing steam or hot water or for space and/or process cooling, heating, or controlling humidity. Based on the equipment used, CHP is also known by other acronyms such as CHPB (Cooling Heating and Power for Buildings), CCHP (Combined Cooling Heating and Power), BCHP (Building Cooling Heating and Power) and IES (Integrated Energy Systems). CHP systems are much more efficient than producing electric and thermal power separately. According to the Commercial Buildings Energy Consumption Survey, 1995 [14], there were 4.6 million commercial buildings in the United States. These buildings consumed 5.3 quads of energy, about half of which was in the form of electricity. Analysis of survey data shows that CHP meets only 3.8% of the total energy needs of the commercial sector. Despite the growing energy needs, the average efficiency of power generation has remained 33% since 1960 and the average overall efficiency of generating heat and electricity using conventional methods is around 47%. And with the increase in prices in both electricity and natural gas, the need for setting up more CHP plants remains a pressing issue. CHP is known to reduce fuel costs by about 27% [15] CO released into the atmosphere. The objective of this research is to review issues related to the design of heat recovery unit in Combined Heat and Power (CHP) systems. To meet specific needs of CHP systems, configurations can be altered to affect differentfactors of the design. Before the design process can begin, product specifications, such as steam or water pressures and temperatures, and equipment, such as absorption chillers and heat exchangers, need to be identified and defined.The Mechanical Engineering Department and the Industrial Assessment Center at the University of Louisiana Lafayette has been donated an 800kW diesel turbine and a 100 ton absorption chiller from industries. This equipment needs to be integrated to work as a Combined Heat and Power system for the University which will supplement the chilled water supply and electricity. The design constraints of the heat recovery unit are the specifications of the turbine and the chiller which cannot be altered.Integrating equipment to form a CHP system generally does not always present the best solution. In our case study, the absorption chiller is not able to utilize all of the waste heat from the turbine exhaust. This is because the capacity of the chiller is too small as compared to the turbine capacity. However, the need for extra space conditioning in the buildings considered remains an issue which can be resolved through the use of this CHP system. BACKGROUND LITERATUREThe decision of setting up a CHP system involves a huge investment. Before plunging into one, any industry, commercial building or facility owner weighs it against the option of conventional generation. A dynamic stochastic model has been developed that compares the decision of an irreversible investment in a cogeneration system with that of investing in a conventional heat generation system such as steam boiler combined with the option of purchasing all the electricity from the grid [21]. This model is applied theoretically based on exempts. Keeping in mind factors such as rising emissions, and the availability and security of electricity supply, the benefits of a combined heat and power system are many.CHP systems demand that the performance of the system be well tested. The effects of various parameters such as the ambient temperature, inlet turbine temperature, compressor pressure ratio and gas turbine combustion efficiency are investigated on the performance of the CHP system and determines of each of these parameters [1]. Five major areas where CHP systems can be optimized in order to maximize profits have been identified as optimization of heat to power ratio, equipment selection, economic dispatch, intelligent performance monitoring and maintenance optimization [6].Many commercial buildings such as universities and hospitals have installed CHP systems for meeting their growing energy needs. Before the University of Dundee installed a 3 MW CHP system, first the objectives for setting up a cogeneration system in the university were laid and then accordingly the equipment was selected. Considerations for compatibility of the new CHP setup with the existing district heating plant were taken care by some alterations in pipe work so that neither system could impose any operational constraints on the other [5]. Louisiana State University installed a CHP system by contracting it to Sempra EnergyServices to meet the increase in chilled water and steam demands. The new cogeneration system was linked with the existing central power plant to supplement chilled water and steam supply. This project saves the university $ 4.7 million each year in energy costs alone and 2,200 emissions are equivalent to 530 annual vehicular emissions.Another example of a commercial CHP set-up is the Mississippi Baptist Medical Center. First the energy requirement of the hospital was assessed and the potential savings that a CHP system would generate [10]. CHP applications are not limited to the industrial and commercial sector alone. CHP systems on a micro-scale have been studied for use in residential applications. The cost of UK residential energy demand is calculated and a study is performed that compares the operating cost for the following three micro CHP technologies: Sterling engine, gas engine, and solid oxide fuel cell (SOFC) for use in homes [9].The search for different types of fuel cells in residential homes finds that a dominant cost effective design of fuel cell use in micro – CHP exists that is quickly emerging [3]. However fuel cells face competition from alternate energy products that are already in the market. Use of alternate energy such as biomass combined with natural gas has been tested for CHP applications where biomass is used as an external combustor by providing heat to partially reform the natural gas feed [16]. A similar study was preformed where solid municipal waste is integrated with natural gas fired combustion cycle for use in a waste-to-energy system which is coupled with a heat recovery steam generator that drives a steam turbine [4]. SYSTEM DESIGN CONSIDERATIONSIntegration of a CHP system is generally at two levels: the system level and the component level. Certain trade-offs between the component level metrics and system level metrics are required to achieve optimal integrated cooling, heating and power performance [18]. All CHP systems comprise mainly of three components, a power generating equipment or a turbine, a heat recovery unit and a cooling device such as an absorption chiller.There are various parameters that need to be considered at the design stage of a CHP project. For instance, the chiller efficiency together with the plant size and the electric consumption of cooling towers and condenser water pumps are analyzed to achieve the overall system design [20]. Absorption chillers work great with micro turbines. A good example is the Rolex Reality building in New York, where a 150 kW unit is hooked up with an absorption chiller that provides chilled water. An advantage of absorption chillers is that they don’t require any permits or emission treatment [2]Exhaust gas at 800°F comes out of the turbine at a flow rate of 48,880 lbs/h [7]. One important constraint during the design of the CHP system was to control the final temperature of this exhaust gas. This meant utilizing as much heat as required from the exhaust gas and subsequently bringing down the exit temperature. After running different iterations on temperature calculations, it was decided to divert 35% of the exhaust air to the heat exchanger whilethe remaining 65% is directed to go up the stack. This is achieved by using a diverter damper. In addition, diverting 35% of the gas relieves the problem of back pressure build-up at the end of the turbine.A diverter valve can also used at the inlet side of the heat exchanger which would direct the exhaust gas either to the heat exchanger or out of the bypass stack. This takes care of variable loads requirement. Inside the heat exchanger, exhaust gas enter the shell side and heats up water running in the tubes which then goes to the absorption chiller. These chillers run on either steam or hot water.The absorption chiller donated to the University runs on hot water and supplies chilled water. A continuous water circuit is made to run through the chiller to take away heat from the heat input source and also from the chilled water. The chilled water from the absorption chiller is then transferred to the existing University chilling system unit or for another use.Thermally Activated DevicesThermally activated technologies (TATs) are devices that transform heat energy for useful purposed such as heating, cooling, humidity control etc. The commonly used TATs in CHP systems are absorption chillers and desiccant dehumidifiers. Absorption chiller is a highly efficient technology that uses less energy than conventional chilling equipment, and also cools buildings without the use of ozone-depleting chlorofluorocarbons (CFCs). These chillers can be powered by natural gas, steam, or waste heat.Desiccant dehumidifiers are used in space conditioning by removing humidity. By dehumidifying the air, the chilling load on the AC equipment is reduced and the atmosphere becomes much more comfortable. Hot air coming from an air-to-air heat exchanger removes water from the desiccant wheel thereby regenerating it for further dehumidification. This makes them useful in CHP systems as they utilize the waste heat.An absorption chiller is mechanical equipment that provides cooling to buildings through chilled water. The main underlying principle behind the working of an absorption chiller is that it uses heat energy as input, instead of mechanical energy.Though the idea of using heat energy to obtain chilled water seems to be highly paradoxical, the absorption chiller is a highly efficient technology and cost effective in facilities which have significant heating loads. Moreover, unlike electrical chillers, absorption chillers cool buildings without using ozone-depleting chlorofluorocarbons (CFCs). These chillers can be powered by natural gas, steam or waste heat.Absorption chiller systems are classified in the following two ways:1. By the number of generators.i) Single effect chiller –this type of chiller, as the name suggests, uses one generator and the heat released during the absorption of the refrigerant back into the solution is rejected to the environment.ii) Double effect chiller –this chiller uses two generators paired with a single condenser, evaporator and absorber. Some of the heat released during the absorption process is used to generate more refrigerant vapor thereby increasing the chiller’s efficiency as more vapor is generated per unit heat or fuel input. A double effect chiller requires a higher temperature heat input to operate and therefore its use in CHP systems is limited by the type of electrical generation equipment it can be used with.iii) Triple effect chiller –this has three generators and even higher efficiency than a double effect chiller. As they require even higher heat input temperatures, the material choice and the absorbent/refrigerant combination is limited.2. By type of input:i) Indirect-fired absorption chillers –they use steam, hot water, or hot gases from a boiler, turbine, engine generator or fuel cell as a primary power input. Indirect-fired absorption chillers fit well into the CHP schemes where they increase the efficiency by utilizing the otherwise waste heat and producing chilled water from it.ii) Direct-fired absorption chillers –they contain burners which use fuel such as natural gas. Heat rejected from these chillers is used to provide hot water or dehumidify air by regenerating the desiccant wheel.An absorption cycle is a process which uses two fluids and some heat input to produce the refrigeration effect as compared to electrical input in a vapor compression cycle in the more familiar electrical chiller. Although both the absorption cycle and the vapor compression cycle accomplish heat removal by the evaporation of a refrigerant at a low pressure and the rejection of heat by the condensation of refrigerant at a higher pressure, the method of creating the pressure difference and circulating the refrigerant remains the primary difference between the two. The vapor compression cycle uses a mechanical compressor that creates the pressure difference necessary to circulate the refrigerant, while the same is achieved by using a secondary fluid or an absorbent in the absorption cycle [11].The primary working fluids ammonia and water in the vapor compression cycle with ammonia acting as the refrigerant and water as the absorbent are replaced by lithium bromide (LiBr) as the absorbent and water (H2O) as the refrigerant in the absorption cycle. The process occurs in two shells - the upper shell consisting of the generator and the condenser and the lower shell consisting of the evaporator and the absorber.Heat is supplied to the LiBr/H2O solution through the generator causing the refrigerant (water) to be boiled out of the solution, as in a distillation process. The resulting water vapor passes into the condenser where it is condensed back into the liquid state using a condensing medium. The water then enters the evaporator where actual cooling takes place as water is passes over tubes containing the fluid to be cooled.Heat ExchangerA very low pressure is maintained in the absorber-evaporator shell, causing the water to boil at a very low temperature. This results in water absorbing heat from the medium to be cooled and thereby lowering its temperature. The heated low pressure vapor then returns to the absorber where it mixes with the LiBr/H2O solution low in water content. Due to the solution’s low water content, vapor gets easily absorbed resulting in a weaker LiBr/H2O solution. This weak solution is pumped back to the generator where the process repeats itself.The heat recovery steam generator (HRSG) is primarily a boiler which generates steam from the waste heat of a turbine to drive a steam turbine. The heat recovery boiler design for cogeneration process applications covers many parameters. The boiler could be designed as a fire-tube, water tube or combination type. Further for each of these parameters, there is a variety of tube sizes and fin configurations. For a given boiler, a simplified method that determines the boiler performance has been developed [8].The shell and tube heat exchanger is the most common and widely used heat exchanger in different industrial applications [13]. It is compared to a classic instrument in a concert playing all the important nodes in different complex system set-ups and can be improved by using helical baffles. There are other ways to augment the heat transfer in a shell and tube exchanger such as through the use of wall-radiation [25].The design of a shell and tube heat exchanger fora combined heat and power system basically involves determining its size or geometry by predicting the overall heat transfer coefficient (U). The process of obtaining the heat transfer coefficient values is obtained from literature by correlating results from previous findings in the determination of heat exchanger designs.This involves listing assumptions at the beginning of the procedure, obtaining fluid properties, calculation of Reynolds number and the flow area to obtain the shell and tube sizes. Once U is calculated, the heat balances are calculated. This study also compares the theoretical U values with the actual experimental ones to prove the theoretical assumptions and to obtain the optimum design model [18].A mathematical simulation for the transient heat exchange of a shell and tube heat exchanger based on energy conservation and mass balance can be used to measure the performance. The design of the heat exchanger is optimized with the objective function being the total entropy generation rate considering the heat transfer and the flow resistance [20].Once a heat exchanger is designed, a total cost equation for the heat exchanger operation is deduced. Based on this, a program is developed for the optimal selection of shell-tube heat exchanger [24].The heat exchanger to be used in the CHP system in the end needs to be tested for its performance. A heat recovery module f orcogeneration is tested before use for CHP application through a microprocessor based control system to present the system design and performance data [19].The basis of a CHP system lies in efficiently capturing thermal energy and using it effectively. Generally in CHP systems, the exhaust gas from the prime mover is ducted to a heat exchanger to recover the thermal energy in the gas. The commonly used heat recovery systems are heat exchangers and Heat Recovery Steam Generators depending on whether hot water or steam is required.The heat exchanger is typically an air-to-water kind where the exhaust gas flows over some form of tube and fin heat exchange surface and the heat from the exhaust gas is transferred to make hot water. Sometimes, a diverter or a flapper damper is used to maintain a specific design temperature of the hot water or steam generation rate by regulating the exhaust flow through the heat exchanger.The HRSG is essentially a boiler that captures the heat from the exhaust of a prime mover such as a combustion turbine, gas or diesel engine to make steam. Water is pumped and circulated through the tubes which are heated by exhaust gases at temperatures ranging from 800°F to 1200°F. The water can then be held under high pressure to temperatures of 370°F or higher to produce high pressure steam [21].The Delaware method is a rating method regarded as the most suitable open-literature available for evaluating shell side performance and involves the calculation of the overall heat transfer coefficient and the pressure drops on both the shell and tube side for single-phase fluids [12]. This method can be used only when the flow rates, inlet and outlet temperatures, pressures and other physical properties of both the fluids and a minimum set of geometrical properties of the shell and tube are known. Emission ControlEmission control technologies are used in the CHP systems to remove SO2 (sulphur dioxide), SO3 (sulphur trioxide) NOx (nitrous oxide) and other particulate matter present in the exhaust of a prime mover. Some common emission control technologies are:1、Diesel Oxidation Catalyst (DOC) –They are know to reduce emissions of carbon monoxide by 70 percent, hydrocarbons by 60 percent, and particulate matter by 25 percent (Emissions Control : CHP Technologies Gulf Coast CHP 2007) when used with the ultra-low sulfur diesel (ULSD) fuel. Reductions are also significant with the use of regular diesel fuel.2、Diesel Particulate Filter (DPF) - DPF can reduce emissions of carbon monoxide, hydrocarbons, and particulate matter by approximately 90 to 95 percent (Emissions Control : CHP Technologies Gulf Coast CHP 2007). However, DPF are used only in conjunction with ultra-low sulfur diesel (ULSD) fuel.3、Exhaust Gas Recirculation (EGR) – They have a great potential for reducing NOx emissions.4、Selective Catalytic Reduction (SCR) –SCR cuts down high levels of NOx by reducing NOx to nitrogen (N2) and oxygen (O2).5、NOx absorbers –catalysts are used which adsorb NOx in the exhaust gas and dissociates it to nitrogen.CONCLUSIONSThe various components needed in a CHP system have been presented. Important parameters such as the mass flow rates of the exhaust gas and water can then be defined. The CHP system has been integrated by the use of a heat recovery unit, the design of which has been discussed. A shell and tube configuration is commonly selected based on literature survey. The pressure drops at both the shell and the tube side can be calculated after the exchanger has been sized.Integrating equipment to form a CHP system generally does not always present the best solution. In our case study, the absorption chiller is not able to utilize all of the waste heat from the turbine exhaust. Approximately 65% goes is left to go out the stack. This is because the capacity of the chiller is too small as compared to the turbine capacity. However, the need for extra space conditioning in the buildings considered remains an issue which can be resolved through the use of this CHP system.The heat exchanger designed can either be constructed following the TEMA standards or it can be built and purchased from an industrial facility. The design that is used is based on the methodology of the Bell-Delaware method and the approach is purely theoretical, so the sizing may be slightly different in industrial design. Also the manufacturing feasibility needs to be checked.After the heat exchanger is constructed, the CHP equipment can be hooked together. Again since the available equipment is integrated to work as a system, the efficiency of the CHP system needs to be calculated. Some kind of co ntrol module needs to be developed that can monitor the performance of the entire system. Finally, the cost of running the set-up needs to be determined along with the air-conditioning requirements.。

Heat exchanger is the equipment that transfer the heat from the heat fluid to the cold fluid. It is widely used in the chemical field, power field, food and other industrial field, so it plays an important role in the industry. In the chemical field, heat exchanger can be taken as the heater, cooler, condenser, evaporator and reboiler and so on. According to the working principle, it can be divided into three types: recuperative heat exchanger, mixing heat exchanger and regenerative heat exchanger. Among these three typies heat exchangers, recuperative heat exchanger is most widely used.换热器，是将热流体的部分热量传递给冷流体的设备，又称热交换器。

换热器是化工、石油、动力、食品及其它许多工业部门的通用设备，在生产中占有重要地位。

在化工生产中换热器可作为加热器、冷却器、冷凝器、蒸发器和再沸器等，应用更加广泛。

换热器种类很多，但根据冷、热流体热量交换的原理和方式基本上可分三大类即：间壁式、混合式和蓄热式。

在三类换热器中，间壁式换热器应用最多。

弓形折流板换热器螺旋折流板换热器板翅式换热器板式换热器混合式换热器（冷却塔）。

最新精品文档，知识共享！化学工程与工艺102(2016)1–8Contents lists available at ScienceDirect化学工程与工艺:增强过程期刊主页: /locate/cepT.Srinivas,A.VenuVinod*化学工程技术研究所,瓦朗加尔506004,印度文章信息文章历史:收到 2015年10月10日收到修订版 2016年1月8日 接收 2016年1月11日 可在线2016年1月14日 关键字: Dean 数 增强 传热率 螺旋形线圈 纳米流体ã2016ElsevierB.V.Allrightsreserved.1.引言* 作者通讯地址.E-mail address: ****************(A. VenuVinod)./10.1016/j.cep.2016.01.005 0255-2701/ã2016Elsevier B.V. All rights reserved. 采用水性纳米流体在壳侧和螺旋管换热器的传热强化摘要纳米流体已被报道为能够加强热的交换。

外壳和螺旋盘管换热器的性能已经使用三个水性纳米流体实验验证。

（氧化铝，氧化铜和二氧化钛）。

这些研究是在不同浓度的纳米流体，以及纳米流体的温度，搅拌速度和线圈侧的流体溢流率进行的。

三种纳米流体的浓度为0.3，0.6，1，按重量计 1.5至2％的制备。

使用十六烷基三甲基溴（CTAB ）用作稳定剂。

纳米流体作为加热介质（外壳侧）和水作为线圈侧的流体。

结果发现，在纳米流体浓度的增加以及热传递速率增加，纳米流体浓度，搅拌速度和壳侧的值越高，热交换器有越高的效率。

当与水进行对比时发现Al2O3，CuO 和纳米TiO2 /纳米水的浓度在30.37％，32.7％和26.8％时有最大增加率。

热交换器的传热可用主动，被动和复合热转移技术实现。

该活跃的技术需要外部力量，例如，电动场，表面振动等的无源技术需要流体的添加剂（例如，纳米颗粒），或特殊的表面几何形状（例如，螺旋线圈）。

板式换热器介绍范文Plate Heat Exchanger IntroductionPlate heat exchangers are often used as a way to efficiently recover energy from industrial processes as the plates can be designed to optimize heat transfer in the specific application. Plate heat exchangers are also able to minimize the total energy consumption of a process by using the ideal number of plates to maximize the heat transfer surface area while minimizing the pressure drop. This minimizes the amount of energy required to transition heat between different fluids, and can providegreater energy efficiency.Plate heat exchangers can also be used for cooling applications. For example, in HVAC systems, a plate heat exchanger can be used to reduce temperature of inlet air by exchanging heat between the inlet air and cooler air or water. This can help to reduce the amount of energy required to coolthe air by allowing the heat to be exchanged with a cooler fluid, rather than using direct cooling.Plate heat exchangers can also be used in order to better control the temperature of fluids in a process or system. By exchanging heat between two fluids, it is possible to regulate temperature of a fluid to a desired temperature by controllingthe flow of the hot and cold fluids. This can help to ensure aconsistent temperature of the fluid, allowing for better control of the process.In addition to cooling and temperature control applications, plate heat exchangers can be used in a variety of other applications as well. For example, plate heat exchangers can be used for condensing, boiling, heating, and other applications where efficient heat transfer is needed.The use of plate heat exchangers allows for more efficient heat transfer than traditional shell and tube heat exchangers and can be designed to optimize heat transfer in a specific application. Furthermore, plate heat exchangers require less maintenance than shell and tube heat exchangers and require less space for installation. This makes them ideal for use in many different applications.Overall, plate heat exchangers are one of the most efficient and reliable types of heat transfer equipment available. They are often used for cooling, temperature control, condensing, boiling, and other applications where efficient heat transfer is needed. Additionally, plate heat exchangers require minimal maintenance and installation space, allowing for greater flexibility in their application.。

赛科 PROPRIETARY INFORMATIONTO BE MAINTAINED IN CONFIDENCELEARNING & DEVELOPMENT SERVICES培训和开发服务PRESENTS课程介绍MODULE: Exchangers-1模块：热交换器-1DESIGNED FORENHANCING OPERATIONS KNOWLEDGE & SKILLS适用于提高操作知识和技能STUDENT PACKAGE学生部分SECCO赛科LEARNING AND DEVELOPMENT SERVICES培训和开发服务MODULE: Exchangers-1模块：热交换器1HEAT TRANSFER and SHELL & TUBE HEAT EXCHANGERS热传递和管壳式热交换器赛科培训和开发服务能量守恒–热交换器赛科培训和开发服务能量守恒–热交换器赛科培训和开发服务能量守恒–热交换器TABLE OF CONTENTS目录If viewing this TOC on a computer, you can move directly to a subject area by pointing with your cursor and single clicking.如果你在计算机上查看本目录，你可以通过移动光标到想要主题的标题上，单击之后直接进入到想要的主题区域INTRODUCTION (9)LEARNING ACTIVITIES (10)DEFINITION (11)G ENERAL C ATEGORIES OF H EAT E XCHANGER E QUIPMENT (12)HOW HEAT EXCHANGERS WORK (13)H EAT T RANSFER (13)M ETHODS OF H EAT T RANSFER (13)Convection (13)Conduction (17)Radiation (19)FACTORS THAT AFFECT HEAT TRANSFER (23)T YPE AND A MOUNT OF F LUIDS (25)T EMPERATURE D IFFERENCE (25)M ATERIALS U SED (25)C ONTAMINATION污染 (26)HEAT EXCHANGERS (27)赛科培训和开发服务能量守恒–热交换器O VERVIEW (27)T YPES OF H EAT E XCHANGERS (27)BASIC DESIGN OF HEAT EXCHANGERS (30)S HELL AND T UBE E XCHANGERS (30)C OMPONENTS OF S HELL AND T UBE E XCHANGERS (33)Shell Side Components (33)Tube Side Components (35)U-T UBE AND F IXED T UBE E XCHANGERS (37)U-Tube (37)Fixed Tube Sheets (39)DETERMINING THE FLOW OF PRODUCT STREAMS (41)T UBE S IDE P RODUCT S TREAMS (41)S HELL S IDE P RODUCT S TREAMS (43)CONDENSERS (45)H OT V APORS I N热蒸汽进口 (45)L IQUID O UT液体出口 (45)W ARM W ATER热水 (45)W ATER C OOLED C ONDENSERS (45)HEAT EXCHANGER OPERATING PARAMETERS AND PROBLEMS (47)O PERATING P ARAMETERS (47)O PERATING P ROBLEMS (49)外部泄漏：热交换器外部法兰、阀或垫圈的故障导致产品流向外界泄漏。

case studies in thermal engineering参考文献缩写案例研究在热工程中的应用引言：热工程是一个涉及热能转化和传递的学科领域，涵盖了从热机、热泵到能源系统等广泛的应用。

在热工程中，案例研究是一种常见的方法，以实际的案例为基础，通过分析解决方案和实践经验，提供对特定问题的研究、评估和改进。

本文将介绍两个案例研究，在热工程领域中展示了该方法的应用并取得了令人满意的结果。

案例一：热交换器的优化设计引用文献：Wankhede A, Marathe A, Puntambekar P N. Thermal optimization of double pipe heat exchanger[J]. Case Studies in Thermal Engineering, 2019, 14.热交换器是热工程领域中广泛应用的一种设备，用于热能传递和能量效率的提高。

在这个案例研究中，研究人员对一个双管热交换器的热力学性能进行了优化设计。

他们首先确定了设计参数，包括管道尺寸、材料和换热流体的性质，并建立了相应的数学模型。

通过对模型的数值仿真和实验数据的有效验证，研究人员发现通过调整管道的截面积和长度可以显著改善热交换器的换热效率。

他们还发现在一定程度上增加流体的流速可以提高传热性能。

这些结果为进一步优化设计提供了有价值的参考。

通过案例研究，研究人员得出了一些结论和建议。

首先，设计者应该考虑流体的性质和实际应用中的换热要求来选择合适的材料和尺寸。

其次，改变流体的流速和温度差异可以使热交换器实现更高的换热效率。

最后，优化设计需要与热工程实践相结合，建立完善的数学模型和实验验证方法。

在这个案例研究中，研究人员通过案例分析和实验验证，证明了优化设计对于提高热交换器性能的重要性。

这个案例也为工程师和设计者提供了指导，使他们能够更好地设计和选择热交换器。

案例二：热泵空调系统的性能改进引用文献：Li J, Chen C, Zou X. Performance improvement of a heat pump air-conditioning system: A case study[J]. Case Studies in Thermal Engineering, 2020, 16.热泵空调系统在提供舒适的室内温度的同时，还能有效地提高能源利用率。

Air-cooled heat exchangers are used extensively throughout the entire oil and gas industry from upstream production to refineries and petrochemical plants, under difficult conditions including high pressures and temperatures, as well as corrosive fluids and environments. This article presents a reliability and integrity process analysis of an air-cooled heat exchanger for hydrocarbon service. This process as part of an AIMP (Asset Integrity Management Program) including, Risk Based Inspection Strategy (RBIS), Acoustic Pulse Reflectometry (APR) inspection and degradation analysis to perform the remaining life assessment (RLA), helping to engineers, maintenance manager and plant manager to make the right decision even in an uncertainty environment. Some defects and discontinuities can be introduced during the heat exchanger manufacturing process and may not necessarily be detected by non-destructive testing processes. Other damage mechanisms, like erosion-corrosion, sulfide stress corrosion cracking, thermal fatigue, and pitting corrosion due to CO2 or because of chloride content in the process stream could come up while heat exchangers are in-service. Due to the equipment complexity, an asset integrity management programme should be developed and applied on-site. An asset integrity management programme seeks to ensure that all equipment and particularly those physical assets that are subject to internal pressure, are designed, constructed, inspected and maintained to the appropriated standards and best practices. This is in order to pursue the maintenance efforts, optimization and cost-effective maintenance decisions, guaranteeing a sustainable and safe operation.。

A Survey on a Heat Exchangers Network to Decrease EnergyConsumption by Using Pinch TechnologyB.Raei and A.H.TarighaleslamiChemical Engineering Faculty,Mahshahr Branch,lslamic Azad University,Mahshahr 63519,lranReceived：April27,2011／Accepted：July7，2011／Published：December20,2011 Abstract：There are several ways to increase the efficiency of energy consumption and to decrease energy consumption．In this paper.The application of pinch technology in analysis of the heat exchangers network(HEN)in order to reduce the energy consumption in a thermal system is studied.Therefore，in this grass root design，the optimum value ofΔTmin is obtained about10℃and area efficiency(α)is0.95.The author also depicted the grid diagram and driving force plot for additional analysis.In order to increase the amount of energy saving，heat transfer from above to below the pinch point in the diagnosis stage is verified for all options including re-sequencing，re-piping,add heat exchanger and splitting of the flows.Results show that this network has a low potential of retrofit to decrease the energy consumption，which pinch principles are planned to optimize energy consumption of the unit.Regarding the results of pinch analysis，it is suggested that in order to reduce the energy consumption.No alternative changes in the heat exchangers network of the unit is required.The acquired results show that the constancy of network is completely confirmed by the high area efficiency infirmity of the heat exchanger to pass the pinch point and from of deriving force plot.Key words：Pinch technology，heat exchangers network，energy consumption，composite curve，grand composite curve1．At the end of1970s，Umeda and his co-workers in Chiyoda established new technology for optimization of process.During1978to1982，this team by presenting of the concept of processes analysis and composite curve showed how the utility consumption can be evaluated and heat recovery and reduction can be done with using this method.At the same time，Linnhoff and his co-workers considered the analysis of heat exchangers network(HEN)for energy consumption reduction and introduced the concepts such as composite curve as an important tool for heat energy recovery.But contrary to Chiyoda team，they emphasized on a pinch point as a key point for heat recovery and by this reason they chose the name of pinch technology for this method.When the time passed，pinch technology has been developed.As the same as HEN,it is used for optimization of energy consumption in distillation towers，furnaces,evaporators，turbines and reactors.Pinch technology is a systematic method based on first and second laws of thermodynamic，which is used for analysis of chemical processes and utilities.Pinch analysis of an industrial process is used for definition of energy and capital costs of HEN before design and also definition of pinch point.In this method,before design,minimum consumption of utility，minimum demanded network area and minimum number of demanded heat unit at pinch pointare targeted for given process．At next stage，design of HEN will be done to satisfy performed target.Finally，minimum annual cost is obtained with comparison between energy cost and capital cost and trade of them.Therefore，the main goal of pinch analysis is the optimization of process heat integration，increase the process-process heat recovery,and decrease the amount of utility consumption．For analysis,at first,shifted temperature is obtained then temperature and enthalpy plot draw(half of amount of minimum temperature are deducted from hot stream and added to cold stream).Fig.1shows the composite curve and grand composite curve as tools for pinch Analysis.The composite curves(CCs)present the relationship between cumulative enthalpy flow rate and temperature for the HEN hot and cold streams．In practice，CCs are generated by a cumulative process over a temperature range，and the resulting hot and cold CCs are labelled CCh and CCc，respectively.2．Methods and Data2.1Presentation of a Heat Exchanger NetworkIn a heat exchanger network，arrangement of exchangers in the network is important.For representing such arrangement，the concept of“stage"is used.In every stage,the input and output heat of the stage is equal for the entire exchangers that settled on special stream，whereas the number of stages is not too many in an optimal network.In this part，stages of heat exchanger networks analysis for reduction of energy consumption using pinch technology were explained.Since targeting and design is based on extracted data any mistake and careless in data assembling can lead to completely unreal results.In pinch analysis，design data such as supply and target temperature of streams，flow and heat capacity of stream was used and on the other hand，heat exchangers design was related to heat transfer coefficient directly.In Table1,the necessary extracted information and a sample network is represented．In this research，Aspen pinch software has been used．Fig.1Tools for pinch analysis：composite curve(CC)and grand composite curve(Gee)．Table1Extracted data．2.2Economical DataCorrect economic data including operation time ，interest rate and equipment life have an important role on successful execution ofretrofit and preparation ．The values are shown in Table 2．The condition of utilities which includes steam and cooling water is shown in Table 3．Capital cost and energy cost of network can be calculated with respect to the shells number and the cost of any exchanger calculates with using Eq ．(1)：c Area b a t CapitalCos )(+=(1)In this equation ，a ，b and C are constant ．So that ，“a”is function of pressure intensity ．“b”is function of exchanger material and “c”is function of type of exchanger that is different for various exchangers ；SO 0<C <1．Types of exchanger are defined by designer based on nature of chemical materials ，pressure of flows ，pressure condition and ability of corrosion ．For carbon-still exchanger ，cost equation is as follow ：81.0)(75030800Area t CapitalCos +=。

外文文献：Design and Implementation of Heat Exchange Station Control System Keywords: Heat exchange station, Control system, PLC, Inverter, Configuration software.Abstract. This paper introduces a design and implementation of heat exchange station control system based on PLC and industrial configuration software, which includes the control scheme and principle, hardware selection and software design, etc. The circulating pumps and replenishing pumps in the system can all be driven automatically by PLC and inverter. Main process parameters, such as steam pressure and measurement temperature and so on,can all be shown on the industrial PC running configuration software, and instructions could be sent by the engineer and operator on-the-spot via the Human Machine Interface as well. The automatic pressures adjustment of steam supply of the heater by advanced PID algorithm has been realized finally. It is verified that the system is highly reliable and stable, and it greatly enhances the level of automation and pressure control accuracy of the heat exchange station and meets all the equipments running demands well.IntroductionWith the rapid development of economy and society, heat supply systems are the key power source in the communities and plants in China. As a media between heat sources and heat loads in the systems, a heat exchange stations plays a very important role for the heat supply quality. Traditionally, most of the pumps in the heat supply systems are operated by valves manually, so it could bring about the power energy consuming, high labor intensity and low operation automation. In this paper a design of control system for heat exchange station based on PLC, inverter and industrial configuration software was proposed,accordingly the aim for power energy saving,high heat efficiency and operation automation has been achieved.Process outline and Control demandsProcess outline. The process outline and control demands were put forward at first before the scheme and design of heat exchange station control system were proposed.Heat exchange station consists of a steam-driven heater,plus3circulating pumps,2replenishing pumps and electric control valve. By adjusting the steam flux into the mixture of water and steam according to the temperature sensors mounted indoors and outdoors, the process of heat exchange could be completed. Among these equipments, the steam-driven heater, a heat exchanger containing mixture of steam-and-water, is the key appliance for heat supply system.Control demands. Major control demands for the control system were listed as follows [1]:(a)Pumps driving.Pumps include3circulating pumps(2in operation,1for backup)and 2 replenishing pumps (1 in operation, 1 for backup). Among circulating ones one is driven by power frequency, the others are driven by variable frequency, with 75KW power each; among replenishing ones one is driven by power frequency, the other is driven by variable frequency, with 3KW power each. The control signal should be originated from the pressure difference between the supply water and return water.Pumps could be driven in stepless speed regulating when connecting variable power;(b)Parameters Showing.The showing parameters contain temperature showing-temperature of supply water, return water, the indoor, the outdoor and steam - and pressure showing - pressure of supply water, return water and pre-valve and post-valve of the steam etc;(c) Butterfly valves driving.Two butterfly valves can be on or off automatically when the whole system start or stop;(d) Motor-driven valves control. By continuously adjusting the opening of the valves according to the signal from the temperature sensors indoors and outdoors, the supply water temperature should be stabilized in the presetting values;(e) HMI (Human Machine Interface) Demands. The process flow chart of heat exchange station and main process parameter can be shown in HMI, and instructions can be transmitted via this interface; (f) Safeguard Function. The circulating pumps should be out of running when heat exchange system is in water needing, and steam should be kept out of the heater when the pumps are not revolving. Hardware Selection of the Control SystemFrom the control demands mentioned above, the controller of the control system can process signals both relay and analog, having the ability of loop adjustment of analog quantity. At the meantime the pumps could run in the working condition of variable frequency, so the hardware selection of the control system for heat exchange station should be made deliberately.PLC Serving as Main Controller.As some experienced electrical engineers known,PLC/PC (Program Controller) is a kind of popular industrial computer, and it can not only accomplish logic control, but also complete many advanced functions, such as analog quantity loop adjustment, and motion control, etc. According to the component amounts of input and output and the needs of control system, FX1N-60MR micro PLC of MITSUBISHI FX series is selected, which having 36 inputs and 24 outputs, and doing analog adjustment by using advanced instruction like PID instruction [2]. Because of the sampling and driving of the analog signal necessarily, PLC should be extended to analog input/output function module like FX2N-4AD (4AD) and FX2N-4DA (4DA) or something like.On one hand,4AD adopted is an analog input module having4channels with12bit high resolution, which could receive 0~+10V voltage signal, 0~20mA or 4~20mA current signal. On the other hand, 4DA chosen could send standard voltage signal and/or current signal, having 4 channels with 12 bit high resolution also. It is something to be mentioned here, the wiring form of current input/output (4~20mA) must be adopted in order to avoiding the strong electromagnetism disturbance in the working field [3].Inverter completing Stepless Speed Regulating. At present, inverter, as an import power electronic converter, can convert constantly power frequency into continually variable frequency. Thus, energy saving, cost consuming and noise reduction can be easily reached by this equipment.In this control system inverter of ACS510 series of ABB Corporation were elaborately chosen, which has many advantages, such as Direct Torque Control (DTC) and advanced applying macro and so on. Its main good points and characteristics are illustrated as follows: it can acquire maximum starting torque (200% normal torque) by using direct excitation; it can be applied to multiple driving systems by using master-slave function;input and output programmable function;high precision of speed regulating, perfect safeguard and alarming steps. Owing to these highlights of this inverter, pumps driving of stepless speed regulating can be easily obtained.There are many applying macro in ACS510 series, but we should only choose manual/automatic macro here as we need.IPC Acting as Monitor&Control Interface.IPC(Industrial Personal Computer)has strong compatibility,extensibility and reliability,which can connect PLC by RS-232serial port conveniently. In the hardware configuration we select IPC H610 series of ADV ANTECH as HMI. MCGS(Monitor Control Generated System),fashionable home-made industrial configuration software, is running on the ADV ANTECH IPC. Using this HMI, the visualization process of Monitor and Control is realized easily, intuitively and vividly.With the sensor/transducer,analog input/output modules,PLC and actuators,e.g.inverter and motor-driven valve, the loop adjustment of steam pressure can be precisely attained, and temperatureof all measure points could be measured also[4].The overall hardware configuration of heat exchange station control system see Fig. 1.Fig. 1 The overall hardware configuration of heat exchange station control systemSoftware Design of the Control SystemLAD Diagram Programming.Out of the thoughts of modular programming, the whole program structure can be divided into such several modules as Initialization Function,upper IPC Communication Function, Relay Control Function, Analog Sampling, Fuzzy PID Adjust Function and Safeguard Function, etc. The flow chart of LAD diagram programming of PLC is shown in Fig. 2. Among these modular functions, it is something worthy to mention of Fuzzy PID Adjust Function. Under some circumstances the using of PID instruction of PLC was not so good at what we expected; therefore, the self-made program of Fuzzy PID adjustment of steam pressure was done from deviation and deviation acceleration of temperature between the indoor and the outdoor in accordance with the Fuzzy Control Theory and its application [5].HMI Configuration. For the sake of the appearance beauty and personalization between machine and human, the MCGS- Monitor Control Generated Software of Beijing MCGS Tech Co. Ltd was adopted. This industrial configuration software has very quick, easy development of configuration process, which can build bi-directional and high speed communication between PLC and upper IPC thru RS422/232 serial port.In the development environment of MCGS, all needed windows and pictures were created, including Main Window of Process Flow, Process Parameters Showing Window, and Key Parameters Setting Window, etc. Vivid and readily interaction between human and machine can be completed by such beautiful pictures and animations when IPC running MCGS.ConclusionsThis design of heat exchange station control system based on FX series PLC,MCGS,and ABB inverter has been realized the pressure automatic adjustment of steam-driven heater as originally expected.More over,design demands of power energy saving,high heat efficiency and lowequipments noise can all be well met. Finally, the practical operation verifies that the system is highly reliable and stable, and it greatly enhances the level of automation and pressure control accuracy of heat exchange station and meets equipments requirements of energy saving and green driving.BEGINInitializationFuzzy PIDAdjust FunctionCommunicationFunctionAnalog OutputNoRelay ControlFunctionAnalog FilteringFunctionCall AnalogSample FunctionSample Over?YesAnalog InputLinear TransferLinear TransferAnalog OutputDrivingSafeguard FunctionFailure Occur?NoYesFailure HandlingRelated MemoryResetENDFig. 2 The flow chat of LAD diagram programming of PLCAcknowledgementComposition of this paper was with the help and under the direction of Senior Engineer Nian-hui Zhang of Qingdao Wellborn Automation Corporation.References[1]Information on /[2]H. Zhang, S.D. Li: The Principle of PLC with its Applications to Process Control (China PowerPress, Beijing 2008).[3]H. Zhang: The Design and Development of MITSUBISHI FX Series PLC (China Machine Press,Beijing 2009).[4]H. Zhang: Process Automation Instrumentation, V ol. 31(4) (2010), p. 34-36, in Chinese.[5]L.A. Zadeh: Fuzzy Sets and their Applications (Academic Press, New York 1975).Progress in Civil Engineering10.4028//AMM.170-173Design and Implementation of Heat Exchange Station Control System 10.4028//AMM.170-173.2666外文翻译：换热站控制系统的设计和实现关键词:换热站、控制系统、PLC、变频器、配置软件。

1.2. Basic Heat Exchanger Equations1.2.1. The Overall Heat Transfer CoefficientConsider the situation in Fig. (1.18). Heat is being transferred from the fluid inside (at a local bulk or average temperature of T i ), through a dirt or fouling film, through the tube wall, through another fouling film to the outside fluid at a local bulk temperature of T o . A i and A o are respectively inside and outside surface areas for heat transfer for a given length of tube. For a plain or bare cylindrical tube,i o i o i o r r L r L r A A ==ππ22 (1.13)The heat transfer rate between the fluid inside the tubeand the surface of the inside fouling film is given by anequation of the form Q/A = h(T f - T s ) where the area isA i and similarly for the outside convective processwhere the area is A o . The values of h i and h o have to becalculated from appropriate correlations.On most real heat exchanger surfaces in actual service, afilm or deposit of sediment, scale, organic growth, etc.,will sooner or later develop. A few fluids such as air orliquefied natural gas are usually clean enough that thefouling is absent or small enough to be neglected. Heattransfer across these films is predominantly by conduc-tion, but the designer seldom knows enough about eitherthe thickness or the thermal conductivity of the film to treat the heat transfer resistance as a conductionproblem. Rather, the designer estimates from a table of standard values or from experience a fouling factor R f .R f is defined in terms of the heat flux Q/A and thetemperature difference across the fouling ΔT f by theequation:A Q T R ff /Δ= (1.14)From Eq.1.14, it is clear that R f is equivalent to a reciprocal heat transfer coefficient for the fouling, h f :f f f T A Q R h Δ==1 (1.15)and in many books, the fouling is accounted for by a "fouling heat transfer coefficient," which is still an estimated quantity. The effect of including this additional resistance is to provide an exchanger somewhat larger than required when it is clean, so that the exchanger will still provide the desired service after it has been on stream for some time and some fouling has accumulated.The rate of heat flow per unit length of tube must be the same across the inside fluid film, the inside dirt film, the wall, the outside dirt film, and the outside fluid film. If we require that the temperature differences across each of these resistances to heat transfer add up to the overall temperature difference, (T i - T o ), we obtain for the case shown in Fig.1.18 the equation()o o o fo w i o i fi i i o i A h A R Lk r r A R A h T T Q 12/ln 1++++−=π (1.16)In writing Eq. (1.16), the fouling is assumed to have negligible thickness, so that the values of r i , r o , A i and A o are those of the clean tube and are independent of the buildup of fouling. Not only is this convenient – we don't know enough about the fouling to do anything else.Now we define an overall heat transfer coefficient U * based on any convenient reference area A *:(o i T T A U Q −≡∗∗) (1.17)Comparing the last two equations gives:()o o o fo w i o i fi i i A h A A A R Lk r r A A A R A h A U ******2/ln 1++++=π (1.18)Frequently, but not always , A * is chosen to be equal to A o , in which case U * = U o , and Eq. (1.18) becomes:()o fo w i o o i o fi i i o o h R Lk r r A A A R A h A U 12/ln 1++++=π (1.19)If the reference area A * is chosen to be A i , the corresponding overall heat transfer coefficient U i is given by:()o o i o i fo w i o i fi i i A h A A A R Lk r r A R h U ++++=π2/ln 11 (1.20)The equation as written applies only at the particular point where (T i - T o ) is the driving force. The question of applying the equation to an exchanger in which T i and T o vary from point to point is considered in the next section.The wall resistance is ordinarily relatively small, and to a sufficient degree of precision for bare tubes, we may usually write()()()()w i o i w i o i w i o o w i o o k r r X r Lk r r n A k r r X r Lk r r n A +Δ≅+Δ≅212/;212/ππl l (1.21)Inspection of the magnitudes of the terms in the denominator of Eqs. 1. 19 or 1.20 for any particular design case quickly reveals which term or terms (and therefore which heat transfer resistance) predominates. This term (or terms) controls the size of the heat exchanger and is the one upon which the designer should concentrate his attention. Perhaps the overall heat transfer coefficient can be significantly improved by a change in the design or operating conditions of the heat exchanger. In any case, the designer must give particular attention to calculating or estimating the value of the largest resistance, because any error or uncertainty in the data, the correlation, or the calculation of this term has a disproportionately large effect upon the size of the exchanger and/or the confidence that can be placed in its ability to do the job.1.2.2. The Design IntegralIn the previous section, we obtained an equationthat related the rate of heat transfer to the localtemperature difference (T-t) and the heat transferarea A, through the use of an overall heat transfercoefficient U. In most exchanger applications,however, one or both of the stream temperatureschange from point to point through the flow pathsof the respective streams. The change intemperature of each stream is calculated from theheat (enthalpy) balance on that stream and is aproblem in thermodynamics.Our next concern is to develop a method applyingthe equations already obtained to the case in whichthe temperature difference between the two streamsis not constant. We first write Eq. (1. 17) indifferential form()t T U dQ dA −=** (1.22)and then formally integrate this equation over the entire heat duty of the exchanger, Q t :∫−=t Q o t T U dQ A ** (1.23)This is the basic heat exchanger design equation, or the design integral.U * and A * may be on any convenient consistent basis, but generally we will use U o and A o . U * may be, and in practice sometimes is, a function of the amount of heat exchanged. If 1/U *(T-t) may be calculated as a function of Q , then the area required may be calculated either numerically or graphically, as shown in Fig. (1.19).The above procedure involving the evaluation of Eq. (1.23) is, within the stated assumptions, exact, and may always be used. It is also very tedious and time consuming. We may ask whether there is not a shorter and still acceptably accurate procedure that we could use. As it happens, if we make certain assumptions, Eq. (1.23) can be analytically integrated to the form of Eq. (1.24)()MTD U Q A t**= (1.24)where U * is the value (assumed constant) of the overall heat transfer coefficient and MTD is the "Mean Temperature Difference," which is discussed in detail in the following section.。