(译文)换热器英文参考文献

换热器英语文献

College of Environmental and Energy Engineering, Beijing University of Technology, Beijing 100022, PR China Received 23 October 2004; received in revised form 23 May 2005; accepted 10 June 2005 Available online 27 July 2005
Abቤተ መጻሕፍቲ ባይዱtract
In this part of the paper, the performance of the simultaneous charging/discharging operation modes of the heat pipe heat exchanger with latent heat storage is experimentally studied. The experimental results show that the device may operate under either the ﬂuid to ﬂuid heat transfer with charging heat to the phase change material (PCM) or the ﬂuid to ﬂuid heat transfer with discharging heat from the PCM modes according to the initial temperature of the PCM. The melting/solidiﬁcation curves, the performances of the heat pipes and the device, the inﬂuences of the inlet temperature and the mass ﬂow rate of the cold water on the operation performance are investigated by extensive experiments. The experimental results also disclose that under the simultaneous charging/discharging operation mode, although the heat transfer from the hot water directly to the cold water may vary, it always takes up a major part of the total heat recovered by the cold water due to the very small thermal resistance compared with the thermal resistance of the PCM side. The melting/solidiﬁcation processes taking place in the simultaneous charging/discharging operation are compared with those in the charging only and discharging only processes. By applying a simpliﬁed thermal resistance analysis, a criterion for predicting the exact operation modes was derived and used to explain the observed experimental phenomena. Ó 2005 Elsevier Ltd. All rights reserved.

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Chemical Engineering Science62(2007)1948–1957/locate/cesStructure and rate of growth of whey protein deposit from in situ electrical conductivity during fouling in a plate heat exchanger Romuald Guérin,Gilles Ronse,Laurent Bouvier,Pascal Debreyne,Guillaume Delaplace∗UR638,Génie des Procédés et Technologie Alimentaires,INRA,F-59651,Villenueve d’Ascq,FranceReceived7August2006;received in revised form13December2006;accepted15December2006Available online30December2006This paper is dedicated to the memory of Dr Jean-Claude LeulietAbstractThe inﬂuences of calcium concentrations(70.88mg/l),Reynolds number(2000–5000)and temperature(60.96◦C)upon the deposit structure and the rate of growth deposition have been investigated in a plate heat exchanger.This was done from in situ measurements of the deposit electrical conductivity via implementation of stainless steel electrodes in channels combined with assessments of deposit thickness.Calcium ions affect structures of deposits and increase the rate of deposit growth upon heated surfaces.This was attributed to the formation of weaker size aggregates at higher calcium concentrations and a higher number of calcium bindings,which reinforce adhesion forces between protein aggregates.Structures and appearances of deposits also were affected byﬂow rates whatever the calcium concentrations.Deposit growth rate was enhanced by increasingﬂow rate below a critical Reynolds number comprised between3200and5000.On the contrary,above the critical Reynolds number,a limitation of the deposit and/or an escape of the deposit from the fouled layer into the coreﬂuid occurred,caused by the predominance of particle breakage on the deposit formation.Fouling tended to be reduced at higherﬂow rate.It was noteworthy that rates of growth decrease during fouling experiments which may be explained by an increase in local shear stresses leading to particle breakage.᭧2007Elsevier Ltd.All rights reserved.Keywords:Fouling;Whey protein;Calcium ions;Reynolds number;Shear stress;Deposit structure;Plate heat exchanger;Electrical conductivity1.IntroductionPlate heat exchangers(PHEs)are widely used in food indus-tries.Several advantages in using PHEs have been discussed elsewhere(Corrieu,1980;Bond,1981).The main problems en-countered by users of heat exchangers are linked to fouling,cor-rosion or mechanical resistance.Bott(1992)shows that fouling of heat exchangers,classically observed with dairy products, is in the front row of the industrial preoccupations.Fouling of heated surfaces directly contributes toward increased costs in production and energy losses,cleaning,and hinders a constant product quality and overall process efﬁciency(Yoon and Lund, 1989;Delplace et al.,1994;Jeurnink and de Kruif,1995;Visser and Jeurnink,1997).∗Corresponding author.E-mail address:delapla@lille.inra.fr(G.Delaplace).0009-2509/$-see front matter᭧2007Elsevier Ltd.All rights reserved. doi:10.1016/j.ces.2006.12.038In dairy industries,deposits consist of a layer of protein aggregates and minerals(Tissier and Lalande,1986).Among all milk proteins, -lactoglobulin has been identiﬁed as one of the major contributors to fouling as it undergoes thermal denaturation(Lalande et al.,1985;Lalande and Rene,1988; Gotham et al.,1989).Consequently,whey protein concentrate (WPC)solutions often have applied as a modelﬂuid to mimic fouling reactions during pasteurisation of milk both in the bulk and in the deposit at heat surfaces.There has been a considerable amount of work showing that fouling affects hydrodynamic and thermodynamic perfor-mances of heat exchangers.These studies,carried out with different dairy product compositions and process conditions, put forward the main parameters which interfere upon the foul-ing deposit mass(De Jong et al.,1992;Belmar-Beiny et al., 1993;Delplace et al.,1994;Delplace and Leuliet,1995;Fryer et al.,1996;Changani et al.,1997;Visser and Jeurnink,1997;R.Guérin et al./Chemical Engineering Science62(2007)1948–19571949Christian et al.,2002;Prakash et al.,2006).The main param-eters were for instance wall temperatures andﬂow rate as pro-cess parameters and ionic force,type of ions,pH and protein concentration as chemical composition parameters.All these works represent an important step forward in the generation of predictive models both on -lactoglobulin denatu-ration or global thermal performance degradation for the whole heat exchangers(Fryer and Slater,1985;De Jong et al.,1992; Delplace et al.,1994;Fryer et al.,1996;Visser and Jeurnink, 1997).Unfortunately,these models are not powerful enough to explain the distinct cleaning behaviours experimentally ob-served.For instance,Christian et al.(2002)showed that overall cleaning times and cleaning rates,under standard conditions, were dependant on the deposit composition.There is a lack of knowledge concerning the inﬂuence of deposit structure and kinetic of deposit mass upon the cleaning efﬁciency to get rid off the total mass deposit.To overcome these difﬁculties,stud-ies that report the effect of process parameters and composi-tion upon the structure of the fouled layer are required.The aim of this work was partly to contribute to thisﬁeld.In par-ticular,various controlled conditions ofﬂow rate and calcium concentration of a WPC solution in a PHE were carried out to determine the inﬂuence of these parameters upon the structure and the kinetic of fouled layers.The structure and the growth of the fouled layer were estimated in-line from in situ measurements of the electrical conductivity of the fouled deposit.This was done by the im-plementation of two opposite stainless steel electrodes in PHE channels.In the last section,the electrode system was imple-mented in various channels to determine the inﬂuence of the temperature upon the deposit.2.Materials and methods2.1.ModelﬂuidThe modelﬂuid used in this study was reconstituted from WPC75supplied by Armor Proteines(France).The compo-sition of the powder as given by the manufacturer is shown in Table1.Proteins are the main components of the WPC pow-der(76%w/w)in which -lactoglobulin and -lactalbumin represent63%(w/w)and11%(w/w),respectively.Minerals represented less than4%(w/w)of the total dry weight of the powder.To produce solutions with higher mineral concentra-tion,the powder was dispersed in controlled quality water. Water consisted of a mixture of tap water of Lille(France)and soft water using a water softener(HI-FLO1,Culligan, Purolite C100E resin,France).The calcium and sodium con-tents of tap water,determined by atomic absorption spectropho-tometry(Philips,Pye Unicam),varied in the range170–200 and44–64mg/l,respectively.The range of calcium and sodium contents of the soft water were1.0–3.0and304–341mg/l, respectively.The desired content of calcium of the fouling ﬂuid was obtained by mixing raw water,soft water and aﬁxed amount of powder(1%w/w).Water electrical conductivity var-ied from0.113to0.116S/m at20◦C for calcium concentration varying from35to55mg/l.The product electrical conductivity Table1Composition of WPC powder(Armor Protéines,France)and1.0%WPC solutionComponent WPC75powder(%w/w)1.0%WPC75solution(%w/w) Water 5.599.05Lactose100.1Lipids 3.70.037Protein760.76Casein––-lactoglobulin480.48-lactalbumin8.40.084Other19.80.198Minerals40.04Calcium0.450.007–0.00875 Sodium0.700.0277–0.0472 Potassium0.33N.D.Chloride0.40N.D. Phosphorus0.30N.D. Magnesium0.045N.D.Iron0.008N.D.Other 1.77N.D.Other0.8N.D.:Not determined.Fig.1.Schematic of the experimental setup.varied from0.142to0.146S/m at20◦C for a range of calcium of70–90mg/l.The addition of protein powder to the mixing of water modiﬁed the electrical conductivity value of20%. The pH of the modelﬂuid remained between7.3and7.7.2.2.Fouling experimentThe experimental set-up of pilot plant scale is shown in Fig.1.Although there are two heat exchangers(model V7 plates,Alfa-Laval Vicarb,France)in the setup,the fouling1950R.Guérin et al./Chemical Engineering Science 62(2007)1948–1957ELECTRICAL CONDUCTIVITY SENSOR TEMPERATURE PROBE Pi PLATE NUMBERCi RODUCT CHANNEL NUMBERHOT WATER HOT WATER Fig.2.Heating plate heat exchanger ﬂow arrangement and implementation of stainless steel electrodes inside channels.observations were focused on the second.The ﬁrst one was used only to pre-heat the model ﬂuid up to 60◦C where fouling was negligible.Water was used as the heating medium.The model ﬂuid was heated from 60to 96◦C in a countercurrent mode.The choice of temperatures was made taking into account the value of the denaturation temperature of the -lactoglobulin protein.Temperature value for the denaturation of -lactoglobulin is 74.76◦C (Matsudomi et al.,1991;Xiong,1992;Gotham et al.,1989;Liu et al.,1994).PHE setup consisted of 13plates form-ing six passes of one channel for the two sides (Fig.2).The equivalent space between two consecutive plates was 3.93mm.In order to keep the feed composition constant,the ﬂuid was not re-circulated once it was heated through PHEs.During ex-periments,the inlet temperature of hot water was adjusted to ensure a constant outlet model ﬂuid temperature close to 96◦C and a constant proﬁle of product temperature along the PHE as a function of time (i.e.,constant heat ﬂux).The ﬂuid foul-ing layer interface temperature in each channel was assumed constant during fouling runs.In the beginning,the PHE was brought to thermal equilibrium and desired process temperature using reverse osmosis (RO)water.The feed was switched from RO water to model ﬂuid and the experimental run was contin-ued for 330min.After the fouling experiment,model ﬂuid was replaced by cold RO water to bring the temperature of PHE and deposits to ambient temperature.Experiments were performed for various calcium concentrations and Reynolds numbers as shown in Table 2.Reynolds numbers were computed based on physical properties of water,assuming that the presence of 1%WPC in water does not modify them signiﬁcantly.Average Reynolds number for the clean heat exchanger was determined from the distribution of Re along the PHE (Re =2 Q/ w ).Inlet and outlet model ﬂuids and hot water temperatures were measured with platinum resistance probes (type pt100)with a precision of 0.1◦C.Bulk and wall temperatures in chan-nels were measured from J-type thermocouples with a preci-sion of 0.5◦C.Flow rates were measured using electromag-netic ﬂowmeters (Krohne IFM,Germany).All parameters were collected via a data acquisition system (Agilent Technologies 34970A,USA)with an acquisition period of 30s.2.3.Measurements of fouled layer thicknessDeposit thickness on the different plates was obtained by two ways:•Using a pneumatic lifting device of a uniaxial compres-sion machine (DY30Model,Adamel Lhomargy,TMI,USA)which allows to determine the distance between the support of the device and the upper of the fouled or cleaned plate as shown in Fig.3.The precision of the measurement was 0.01mm.The assessments were performed at nine positions on the plate surface.The average value of the deposit thick-ness was computed from the nine positions.•By weighing plates before and after fouling runs using a Mettler apparatus (PM3000,Switzerland)with a preci-sion of 0.1g.From a wet deposit density value equal to 1000kg /m 3(Lalande et al.,1985),the average deposit thick-ness upon each plate was obtained.Of course,this method assumes that the deposit occurs uniformly upon the plate surface.2.4.Electrical conductivity of the depositTwo AISI 304L stainless steel electrodes 0.015×0.01m were implemented in channels C3,C5and C6(Fig.2).Elec-trodes were connected to a commercial conditioning system (STRATOS 2402Cond,Knick,Germany).Electrodes were electrically insulated from metal plates using an insulating stick (Araldite A V138M-HV998,USA).The cell constant of the de-vice was determined with salt solutions whose electrical con-ductivity value was known with precision.The stainless steel electrodes provide an indication of the equivalent electrical resistance R eq through the channel (Fig.4).For ﬁxed operating conditions,the Kirchhoff’s rule allows decoupling the equivalent electrical resistance in terms of fouling ﬂuid electrical resistance (R p )and deposit electrical resistance (R d )as follows:R eq =R p +2R d .(1)R.Guérin et al./Chemical Engineering Science 62(2007)1948–19571951T a b l e 2S u m m a r y o f m e a s u r e d a n d c a l c u l a t e d p a r a m e t e r s d u r i n g h e a t t r a n s f e r t o s t u d y f o u l i n g b e h a v i o u r o f 1%W P C s o l u t i o nR u n M e a n R e (–)C a 2+(m g /l )N a +(m g /l )i ,p (◦C ) o ,p (◦C ) i ,h w (◦C ) o ,h w (◦C )M a s s o f d e p o s i tt =0t ft =0t f t =0t f t =0t fi n c h a n n e l 5(g )A 200072.9344.062.360.096.897.2102.5104.572.873.471.9B 200378.9303.260.059.796.596.6102.6107.771.976.0118.6C 204082.2280.061.561.397.196.3102.7107.973.479.1147.1D 204085.6277.460.461.595.896.9102.0109.271.980.4180.6E 339470.0323.663.863.995.595.7102.0107.475.080.8100.3F 322076.3472.061.361.395.795.5101.7115.973.489.1170.0G 321478.0364.962.662.295.095.0100.0112.374.083.9201.2H 323286.5331.262.763.694.694.6101.4121.773.293.2240.6I 493874.6329.060.861.495.495.2103.4110.974.383.490.2J 492077.4303.061.360.896.296.0103.1113.574.084.994.2K 494277.8340.063.263.995.495.1103.0111.875.286.8116.6L 492687.4306.061.261.495.995.7103.5121.374.393.3190.5R u n¯e d (M D -5)(m m )¯e d (U C M )(m m )w *(◦C ) b (◦C ) e q *(S /m ) p a t b (S /m ) p a t 100◦C (S /m ) d *a t w (S /m )d *a t 100◦C (S /m )k ×104(S /m m i n )A 0.480.4097.991.10.3030.3600.3880.2080.2101.90B 0.800.61101.990.60.3030.3730.4030.2630.2612.33C 0.980.92102.491.60.3070.3710.3980.2850.2832.82D 1.201.16105.490.10.2970.3720.4040.2750.2703.57E 0.670.69103.689.70.2480.3610.3940.1700.1673.96F 1.201.28112.289.70.2840.4720.5040.2510.2408.19G 1.341.42108.889.30.2320.3750.4090.2240.2166.50H 1.601.55118.789.80.3320.4770.5090.3370.3206.88I 0.600.59105.289.80.2460.3910.4230.1470.1425.67J 0.630.63108.590.10.2250.3600.3920.1410.1335.60K 0.780.75106.889.50.2130.3690.4020.1510.1454.75L 1.271.37121.189.80.2730.3790.4110.2280.2096.70p , d a n d e q :f o u l i n g p r o d u c t ,d e p o s i t a n d e q u i v a l e n t e l e c t r i c a l c o n d u c t i v i t y ,r e s p e c t i v e l y ; i ,p a n d o ,p :i n l e t a n d o u t l e t t e m p e r a t u r e o f t h e p r o d u c t ; i ,h w a n d o ,h w :i n l e t a n d o u t l e t t e m p e r a t u r e o f t h e h o t w a t e r ; w a n d b :w a l l a n d b u l k t e m p e r a t u r e i n c h a n n e l 5;¯ed (U C M ):a ve r a g e d e p o s i t t h i c k n e s sf r o m t h e u n i a x i a l c o m p r e s s i o n m a c h i n e ;e d (M D -5):d e p o s i t t h i c k n e s s f r o m m a s s d e p o s i t i n c h a n n e l 5;k :r a t e o f c h a ng e o f th e e q ui v a l e n t e l e c t r i c a l c o n d u c t i v i t y .∗A t330m i n .1952R.Guérin et al./Chemical Engineering Science 62(2007)1948–1957Based on the general relationship linking the electrical resis-tance to the electrical conductivity for a pair of electrodes [ =e E /(AR)with e E the length between the electrodes,A the cross-section and R the electrical resistance]and assuming that (i)the cross-section A is a constant value and (ii)the space of the ﬂuid ﬂow (e ﬂ)is deﬁned as the difference between thespace0.000 N0.005 N e 10.000 N0.005 N e 2Fouling layerStainless steel plateacbdFig.3.The thickness measurement technique using a pneumatic lifting device of a uniaxial compression machine (DY30Model,Adamel Lhomargy,TMI,USA).P 8P 9Isolating materia l Stainless steel electrodes Fluid flow Fouling layer R eqR dR dRp ABFlow directioneEFig.4.Schematic of the fouling layer and equivalent electric resistance diagram.separating the two electrodes (e E )minus the total deposit thick-ness (2e d )(Eq.(2)),the deposit electrical conductivity (DEC, d )can be expressed as a function of model ﬂuid ( p )and equivalent ( eq )electrical conductivities as shown in Eq.(3).e ﬂ=e E −2e d ,(2)d (t =t f )=e E2e d (t =t f ) 1eq (t =tf )−1p+1p−1.(3)At the beginning of the fouling experiment (i.e.,clean PHE),the value of the equivalent electrical conductivity,measured by the device,corresponded to the electrical conductivity of the model ﬂuid ( p )at the product temperature.The electrical con-ductivity of the model ﬂuid was invariant during fouling runs since the inlet temperature of hot water was adjusted to ensure a constant product temperature inside channels as a function of time.At the end of fouling runs (i.e.,fouled plates,t =t f )the deposit thickness was measured and the measurement of the equivalent electrical conductivity allowed to obtain the electri-cal conductivity of the deposit.In order to compare electrical conductivity values of each run,all conductivities were deter-mined at 100◦C as follows (Ayadi,2005): d(100◦C )= d( w )+0.0009×(100− w ), p(100◦C )= p( b )+0.0032×(100− b ),(4)where d(100◦C )represents the DEC value at 100◦C, d( w )is the DEC determined from Eq.(3)at the end wall tempera-ture w , p(100◦C )is the electrical conductivity of the fouling product at 100◦C, p( b )represents the value of the electrical conductivity of the product at the bulk temperature b .Considering a deposit temperature nearly constant and an invariant viscosity value for the product,the only parametersR.Guérin et al./Chemical Engineering Science 62(2007)1948–195719530.20.220.240.260.280.30.320.340.360.380.4Time, (min )E q u i v a l e n t e l e c t r i c a l c o n d u c t i v i t y , (S m -1)Fig.5.Equivalent electrical conductivity during fouling run using 1.0%WPC solution with calcium concentration 78.0mg /l at Re =3200.which affect DEC values are mobility and concentration of ions (Benoıˆt and Deransart,1976).However,considering a poor mobility and diffusion of ions from the bulk ﬂuid through the fouled layers due to protein networks,the DEC values are affected in the majority by the concentration of ions embedded inside the protein structure.Thus,these values constitute a good indicator of the deposit structure.3.Results and discussion3.1.Effect of calcium content on foulingTypical equivalent electrical conductivity change as a func-tion of time,measured in-line from the electrodes in the ﬁfth channel,is illustrated in Fig.5.After the switch from RO water to fouling ﬂuid,the equivalent electrical conductivity reaches a maximum value at t =15min.This value corresponds to the electrical conductivity of the product at the bulk temperature.Data reported in Table 2show that product electrical conduc-tivity values at 100◦C are little affected by the modiﬁcation of the ionic concentration (i.e.,calcium and sodium in the tested range of concentration)of the solution.At the beginning of fouling stages,very slow decreases in equivalent electrical conductivity are recorded with an initial rate k ∗(Fig.5).This region may be attributed to a homogeneous thin layer of irreversibly adsorbed individual protein molecules on clean metal surfaces (Arnebrant et al.,1985;Visser and Jeurnink,1997).Tissier and Lalande (1986)showed that this sublayer had a thickness of 0.02 m after only few minutes of contact;0.4and 1 m after 10and 30min of fouling run.This weak thickness may explain the slight decrease of the slope between t =15and 30min.The slightly decreasing slope (k ∗)indicates that the fouling mechanism starts immediately when fouling product is present in the heating zone,for a temperature higher than unfolding temperature of -lactoglobulin.After this period,the equivalent electrical conductivity decreases24681060657075808590Calcium content, (mg/l)k x 104, (S .m -1.m i n -1)parison of rates of deposit growth (k )as a function of calcium concentration in WPC solution for Re =2000,3200and 5000.(The trend lines represent the curve ﬁt of data .)linearly with time.The rate of electrical conductivity changes k is relatively high (Fig.5).The second decrease in eq may be attributed to the growth and structure changes of fouled layers.Indeed,whatever the Reynolds number,it is observed that the rate of electrical conductivity changes k rises with in-creasing calcium concentrations (Fig.6).This observation is in agreement with Li et al.(1994)observing that calcium induces conformational changes of the -lactoglobulin,facilitating the protein denaturation,but also increases the kinetic of the aggregate formation.A small change in the calcium concen-tration has an important impact upon the kinetic parameter k ,i.e.,the formation of the fouled layer.Fig.7a illustrates the electrical conductivity values of the fouled layer (DEC)obtained at a wall temperature of 100◦C in the ﬁfth channel for varying calcium concentrations at three Reynolds numbers.Whatever the Reynolds number,the DEC increases with the calcium concentration.Considering a low mobility of ions inside the deposit due to protein networks and a constant temperature in C5,this indicates that the DEC1954R.Guérin et al./Chemical Engineering Science 62(2007)1948–19570.10.20.30.4Calcium concentration,(mg.l -1)E l e c t r i c a l c o n d u c t i v i t y o f t h e d e p o s i t , (S .m -1)00.20.40.60.811.21.41.61.8Calcium concentration, (mg.l -1)F o u l e d l a y e r t h i c k n e s s e d , (m m )05001000150020002500300035004000Calcium content, (mg/l)A m o u n t o f d e p o s i t i n c h a n n e l 5, (g /m 2)parison of (a)deposit electrical conductivity at 100◦C,(b)fouled layer deposit and (c)amount of deposit in channel 5after 5.5h of heat transfer in PHE as a function of calcium concentration in WPC solution for Re =2000,3200and 5000.(The trend lines represent the curve ﬁt of data.)is affected by the deposit thickness and its structure which depends on the calcium concentration (Fig.7b).A small change in the calcium concentration has an important impact upon the fouling behaviour.Figs.6and 7indicate that calcium ions (i)are essential in the growth of fouled layers as suggested by Xiong (1992)since amounts of deposit increase with calcium concentration (Fig.7c),(ii)modify the rate of protein aggregation and (iii)lead to a greater cohesion between protein aggregates modify-ing the deposit structure.Indeed,visual analysis of the deposit after fouling runs at Re 3200using 1.0%WPC solutions revealed that fouled layers formed with low calcium content(78mg/l)have a spongy and soft texture whereas deposits formed at higher calcium content (86.5mg/l)are denser and elastic.This observation is in agreement with Pappas and Rothwell (1991)who showed that -lactoglobulin completely aggregated to form compact structures when heated with cal-cium.Simmons et al.(2007)also showed that increasing the levels of calcium had a dramatic effect on the size of the aggre-gates produced,which decreased with increasing mineral con-centration.An explanation for the difference in structure and kinetic is that calcium ions,essentially present in the deposit solid (Tissier and Lalande,1986),lead to lower size aggregates in the range of calcium concentration (70–88mg/l)and favour the growth of fouled layers by formation of bridges between adsorbed proteins and the protein aggregates formed in the bulk (Fig.8).Bridges may be formed via carboxyl groups of amino acids of -lactoglobulin as suggested by Xiong (1992).In-creasing the level of calcium would lead in a higher number of bridges resulting in a bigger stabilisation of protein aggregates as interpreted by Dauﬁn et al.(1987)and Xiong (1992),forming a narrow network which embed other ions present in the solution (i.e.,sodium,magnesium,phosphate,calcium,…),and reinforce the adhesion forces between proteins.3.2.Effect of hydrodynamics conditions on foulingFig.5shows that rates of equivalent electrical conductivity changes are not constant as function of time since the slope of equivalent electrical conductivity decreased after t =180min.This slope modiﬁcation in eq may be due to a decrease of the aggregate deposit and/or an escape of the deposit from the fouled layer into the core ﬂuid caused by particle breakage.This can be a consequence of an additional local shear stress as deposit thickness evolved (Fig.9).Indeed,shear stress in a channel is a function of channel section which is reduced with the growing fouled layer [ = .¯u. /(2(e E −2e d ))].This conﬁrms the assumption of Kern and Seaton (1959)who were the ﬁrst to underline that the formation of a fouled layer is a consequence of the rate of aggregate entry and the rate at which they escape.Fig.10illustrates the evolution of the kinetic parameter k during fouling with a 1.0%WPC solution at calcium concen-tration of 78mg /l as a function of Reynolds number.The in-crease of k between Re 2000and 3200can be explained by a weaker size of aggregates at higher shear rate for a ﬁxed tem-perature (Simmons et al.,2007)favouring the growth of the deposit and resulting in a different deposit structures (Fig.8).Deposit masses in channel 5conﬁrm this trend namely for a ﬁxed calcium concentration,the amount of deposit in C5in-creases between Re 2000and 3200(Fig.7c).Nevertheless,the k parameter decreases between Re 3200and 5000at a ﬁxed calcium concentration.Thus,the decrease of the rate of deposit is due to the increase of Reynolds number which may limit the deposit,compact the structure upon the heated surface and increase the rate of particle breakage.Visual analysis of the appearance of the deposit as a function of Reynolds number conﬁrm the trends.Deposit formed after fouling with a calcium concentration close to 78mg /l at Re 2000has a granular aspectR.Guérin et al./Chemical Engineering Science 62(2007)1948–19571955Adsorbed protein AggregatesStainless steel electrodes Stainless steel plate With lower calcium concentraionEmbedded ionsCalcium bindingsWith higher calcium concentraionFig.8.Schematic illustration of the proposed formation of the deposit with lower and higher calcium concentrations of the WPC solution.0.20.40.60.811.21.47076.37886.5Calcium concentration, (mg/l)L o c a l s h e a r s t r e s s , (P a )20406080100120140160180200Shear stress for fouled channel (C5)Shear stress for cleaned channel (C5)Increase of shear stressI n c r e a s e o f s h e a r s t r e s s , (%)Fig.9.Increase in shear stresses due to fouled layer growth as a function of calcium concentrations in 1.0%WPC solutions at Re =3200.probably due to higher size aggregates whereas deposit formed at Re 3200appears more denser (i.e.,lower aggregates size).Finally,the deposit obtained after fouling at Re 5000appears more smooth and compact which may be the consequence of the increase of the local shear stress (Fig.9).Another way to underline the inﬂuence of shear stress upon the structure of the deposit is the measurement of the electri-cal conductivity of deposits according to Reynolds numbers for a ﬁxed calcium content and temperature (Fig.7a–c).Fig.7a shows that DEC values are similar at Re 2000and 3200while amounts of deposit in C5,and so the thickness of the deposit (Figs.7b and c),are completely different with a scatter close to 35%.In the same order,amounts of deposit in C5are similar for Re 2000and 5000while the correspondingvalues of DEC1234567200032005000Reynolds number, (-)k x 104, (S m -1 m i n -1)Fig.10.Rates of deposit growth (k )as a function of Reynolds number during fouling runs with a 1.0%WPC solution with calcium concentration comprised between 76.0and 78.0mg /l.differ each other.Moreover,for a ﬁxed calcium concentration,DEC values increase with a variation of Reynolds number from 2000to 3200while a decrease in DEC values can be observed above a critical Reynolds number,which could be comprised between 3200and 5000.Since calcium concentration and tem-perature are invariant and considering a poor mobility of ions due to protein networks,differences in the structure and/or the composition of the deposit can be explained by DEC variations.This indicates that shear stress has a dramatic effect upon the structure and the appearance of the deposit whatever the cal-cium concentrations.The differences in the structure of these。

换热器的外文翻译

换热器的外文翻译(总15页) -CAL-FENGHAI.-(YICAI)-Company One1-CAL-本页仅作为文档封面，使用请直接删除毕业设计论文资料翻译专业班级： XXXXXXXXXXXXXXXXXXXXX姓名 XX学号： XXXXXXX外文出处： International Journal of Heat andMass Transfer 49 (2006) 601–610管壳式换热器中壳程流体流动的压降模型,Satish Chand摘要本文提出了一种管壳式换热器壳程压降的理论模型该模型结合了换热管进出口的压降和折流板间隔处的压降的影响模型得出的结果的雷诺数在103到105之间，相比于其他用于研究不同形态的换热器的分析模型，该模型的雷诺数更接近可参考的资料中的实验结果关键词横流速度；压降；折流板；管壳式换热器；壳程流速；圆缺区流速1 引言单相流管壳式换热器被广泛的应用在化学，石油，发电和过程工业中在这种换热器中，一种流体在换热管中流动，同时另一种流体穿过管束在壳程流动一台换热器的设计需要取得热交换和压降之间的平衡因为压降导致泵和风机等流体输送设备操作费用的增加这说明设计换热器的换热能力和测定穿过换热器的压降是同等重要的计算管内流体流动的压力损失相对容易，但是壳程流体流动的压力损失的计算很复杂图1 壳程流路分布图为了计算壳程压力损失需要了解不同流动路径的本质以及他们各自产生的效果。

如图1根据Tinker 和 Buffalo的流动分析可以看到，除了横流流路B在两块折流板间穿过管束外，还有旁流流路C绕过管束从管束和壳体之间通过，这对传热是没有作用的此外还可以看到，有一个泄漏流路D从折流板和壳体的间隙以及由换热管与折流板间的空隙泄漏，而后与主流体相互作用由图1还可以看出流体的流线是在圆缺区域穿过换热管束的这种流动方式在图4中做了进一步的描述由被折流板隔开的圆缺区内的主流体的流动方向相对于管内流体是不同的，在部分折流板之间形成了横流区这需要使用不同的方法计算圆缺区和穿过管束(横流区)的流体的压降类似的，也要把管束中的穿流和内部错流的不同考虑在内在现有参考资料中可使用的方法可以分为两类:(1)实验法，(2)理论模型的发展在参考文献[3，7，9，16，17，21]中给出了不同研究者的试验方法而进一步的关于模型的研究方法在参考文献[2，4，6，8，11，12，14，15，18，20，22，23]中给出纵观这些参考资料可以发现在一些研究方法中提出的理论模型对于计算是相当复杂的，而且，有的模型考虑了一些产生压降的因素却也忽略了另一些因素本文提出了一种主要涉及壳程流体流动的方法致力于建立一种简单的压力损失模型而且将该模型的结果和其他着作中得出的结果进行比较发现该模型的计算结果接近实际值，完全可以供设计者放心使用2 压降模型的发展本模型旨在确定从流体进入壳体到流出壳体的全部压降而这一过程的全部压降可分成如下几部分:（1）换热器横流区末端换热管进出口处的压降（2）横流区内部的压降这一区域的压降是由流过管束的流体和穿过圆缺区由一个内部横流区到下一个连续的内部横流区的流体共同决定的（3）由横流区入口穿过管束到下一块折流板然后穿过圆缺区的流体流动引起的压降相似的，横流区末端的压降也是由从前一个横流区穿过圆缺区而后穿过管束的流体来计算本文中的方法的主要作用是利用这种压降模型在内部横流区和圆缺区计算压降其他部分的压降在本文使用的参考文献中已做出表述穿过管束和圆缺区引起的压降的计算已经考虑了图1和图5表示出的实际流动形式接下来本文将逐一介绍以上提到的各种压降的组成部分换热管进出口的压降(n P ∆)为了计算一台换热器换热管进出口的压降，Gaddis 和Gnielinski 的研究中提到了关于这个问题的表述已在式(1)中给出了式中用∆P n 来定义换热管两端的总压降。

套管换热器英文文献翻译

能量转换和管理98(2015)69 - 2015内容列表可以在科学指引爱思唯尔（世界领先的科技及医学出版公司）能量转换和管理/locate/enconman homepage 日刊数值研究的一个创新设计翅片套管换热器与变量fin-tip 厚度a.高级研究中心纯粹和应用数学(CASPAM)Bahauddin 扎卡里亚正在搜索大学巴基斯坦68000年b.计算机科学、通讯卫星信息技术研究所、韦校园,Mailsi 路,木尔坦路,韦巴基斯坦c.数学系,政府爱默生学院60700年木尔坦,巴基斯坦d.基础科学和人文、工程科技大学Bahauddin 扎卡里亚正在搜索大学,木尔坦68000年,巴基斯坦分析充分发展的层流对流换热翅片的一个创新设计套管换热器(DPHE)和纵向鳍的变厚度提示受到施加力传热速率边界条件研究。

尖厚度控制的小费比底角作为参数的值从0到1对应于不同翅片形状不同的三角形和矩形截面。

到作者的知识,这个参数是首次被引入文学。

间断伽辽金有限元法(DG-FEM)从事目前的工作。

的整体性能提出DPHE 调查了考虑摩擦因素,努塞尔特数和j 因子。

努塞尔特数高达178%和89%的涨幅取得了j 因子相对rectangu-lar 横截面。

这样的收益相对于三角形截面分别为9.5%和19%。

结果表明,新引入的参数提示比底角证明发挥重要作用在套管换热器的设计在降低成本、重量和摩擦损失,提高换热器的传热速率和节能。

因此,它必须被视为一个重要的设计参数对换热器的设计。

通常,热交换器广泛用于这一目的。

有很多技术,用来提高换热器的传热速率但最有效的其中之一是使用安装鳍。

换热器的设计取决于许多特性即上浆、紧凑,传热性能估算,经济方面和压降分析。

目前的调查描述了数值研究的创新设计翅片套管换热器的变量fin-tip 厚度。

换热器外文文献

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.。

不锈钢制换热器的优化设计文献翻译

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)。

换热器效率

作者最近撰文介绍了换热器效率的概念【1—4】。换热器效率的定义为换热器的实际传热率与最佳的传热率之比。
（3）
换热器传热最佳传热效率是换热器的UA和其算术平均温差影响的结果，温差指的是冷热流体的平均温度差。任何具有相同的UA和AMTD的换热器的传热效率总是比传热速率的最佳值小（η≤1）。此外，最佳传热速率只能发生在平衡状态的逆流换热器中【1】。
图1是一个换热器效率作为翅片比数的函数。传热效率最大发生在Fa=0时，从表一可以看出，只发生一种平衡的逆流换热器，或一个平衡逆流换热器具有100％的效率。对于一个给定的Fa，可以从方程4或表一获得，且方程3决定传热。
对翅片的研究分析为换热器效率概念提供了新的见解。对于一个面积一定的翅片，效率可以由下式得出：
（5）
换热器单一翅片的传热效率可以写成
（6）
重新整理方程式3和4，可得逆流式换热器的传热效率方程为
（7）
虽然方程5表明，增加长度或翅片传热系数导致在一个翅片效率的降低，但是从方程6中可以看出增加这两个参数传热总额是实际增长的。在有条件限制中，无限长的翅片效率为零，即使它仍然转让有限数量的热量。同样的性能可以看出一个换热器。由于总传热系数或换热面积增加，翅片类比数Fa增加，导致在换热器效率下降。然而，可以从方程7中看出，传热效率实质上增加了。例如一个翅片，有一个无限大，零效率的换热器，尽管它所传输的热量是有限的。
这些方法发现，在换热器的设计适用范围有限，部分原因在于全局最优值导致了换热器具有无限大的面积【8】。把换热器换热效率和它的熵产率联系起来的努力也没有成功。最低不可逆转似乎没有同换热效率关联，正如Shah和Skiepko所指出的【9】。结果显示，在不可逆转的工作最低点，换热效率可以最高或最低，得出效率并不能判断换热器可逆性【9】。下文的分析是要表明，上面定义的效率是根据热力学第二定律。这将显示出最低的不可逆转性，和换热器最高的效率联系起来，明确如何把第二定律可以扩展应用到换热器上。考虑换热器有一个面积A和一个总传热系数U，冷热流体分别以温度t1和T1，热容分别为Cc和Ch进入。换热器效率可以从方程4得到。换热器的平均温度差是固定的，是从方程8确定的。如上所述，一个平衡的逆流换热器在冷热流体的热容是等于实际换热器的Cmin，具有相同的UA和AMTD将能传输最大的热量。平衡逆流式换热器冷热流体进口温度没有指定，因此，无穷多个换热器传递的热量相同。这篇文章的其余部分介绍了所有这些平衡逆流换热器，一个具有相同的温度比率（t1/ T1），换热器也可以产生最低熵量。

板式换热器外文文献(英文)

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.。

换热器中英文对照外文翻译文献

中英文对照外文翻译(文档含英文原文和中文翻译)换热器的优化选型【摘要】板式换热器的优化选型是根据换热器的用途和工艺过程中的参数和NTU＝KA／MC＝△t／△tm，即传热单元数NTU和温差比（对数平均温差—换热的动力）选择板片形状、板式换热器的类型和结构。

【关键词】平均温差 NTU 板式蒸发器冷凝器1 平均温差△tm从公式Q＝K△tmA，△tm＝1／A ∫A（t1－t2）dA中可知，平均温差△tm是传热的驱动力，对于各种流动形式，如能求出平均温差，即板面两侧流体间温差对面积的平均值，就能出换热器的传热量。

平均温差是一个较为直观的概念，也是评价板式换热器性能的一项重要指标。

1.1 对数平均温差的计算当换热器传热量为dQ ，温度上升为dt 时，则C ＝dQ ／dt ，将C 定义为热容量，它表示单位时间通过单位面积交换的热量，即dQ ＝K （t h －t c ）dA ＝K △tdA ，两种流体产生的温度变化分别为dt h ＝－dQ ／C h ，dt c ＝－dQ ／C c ，d △t ＝d （t h －t c ）＝dQ （1／C c －1／C h ）,则dA ＝[1／k （1／C c －1／C h ）]·（d △t ／△t ），当从A ＝0积分至A ＝A 0时，A 0＝[1／k （1／C c －1／C h ）]·㏑[（t ho －t ci ）／（t hi －t co ）]，由于两种流体间交换的热量相等，即Q ＝C h （t hi －t ho ）＝C c （t co －t ci ），经简化后可知，Q ＝KA 0｛[（t ho －t ci ）－（t hi －t co ）]／㏑[（t ho －t ci ）／（t hi －t co ）]｝，若△t 1＝t hi －t co ，△t 2＝t ho －t ci ，则Q ＝KA 0[（△t 1－△t 2）／㏑（△t 1／△t 2）]＝KA 0△tm ，式中的△tm ＝（△t 1－△t 2）／㏑（△t 1／△t 2）。

毕业设计换热器英文文献翻译中英对照

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）。

(译文)换热器英文参考文献

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。

毕业设计换热器英文文献翻译中英对照

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）。

case studies in thermal engineering参考文献缩写

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。

换热器(英文)

目录用于吸收式制冷剂组解吸塔上的板壳式换热器中的压降的研究 (1)摘要 (1)1引言 (1)2 循环吸收式没有压降的发电机的描述 (2)3 换热器中的沸点和压降 (3)3.1 溴化锂-水-发生器 (4)3.1.1 板式换热器的温度分布 (5)3.2 氨—水蒸发器 (6)3.3 实验注意事项 (7)4 板式换热器用作蒸汽发生器 (8)5 结论 (8)致谢 (9)参考文献 (9)用于吸收式制冷剂组解吸塔上的板壳式换热器中的压降的研究N.Gacía-Hernando a,*,J.A.Almendros-Ibáñez b,c,G.ruiz d,M.de Vega aa能源系统工程(ISE), Ingeniería Departamento de Térmica y Fluidos大学,马德里卡洛斯三世Avda•Leganés、第30条、第28911大学,马德里,西班牙b Industriales非政府Escuela de Albacete的c可再生能源研究所，第02071期，阿尔瓦塞特省，西班牙d 能原效率和可再生能源部门，Reunidas、S.A. 、C / Arapiles第13号、10ª,28015年马德里,西班牙文章历史条：2009年6月17日收到原文2010年2月2日收到修订表2010年10月6日审核通过2010年10月30日在网上发布关键词：吸收系统解吸板式换热器压力降板式换热器摘要我们对板式换热器中的压降对LiBr–H2O（苯丙氨酸）和NH3-H20溶液沸点的影响进行了研究。

对于NH3-H20溶液，压降和温度饱和之间的关系说明，在饱和温度变化很小的情况下，可以存在很大的压力降。

另外，在使用LiBr–H2O时，由于工作压力通常比较低，为了将板式换热器作蒸汽发生器之用，因此，必须将压降作为一个主要限制参数。

在这种情况下，压降会显著改变进入换热器的溶液的沸点，因此，也就需要一个更高的加热流体温度。

换热器外文翻译 (2)

Heat ExchangersKey Terms Baffles—evenly spaced partitions in a shell and tube heat exchanger that support the tubes, prevent vibration, control fluid velocity and direction, increase turbulent flow, and reduce hot spots. Channel head—a device mounted on the inlet side of a shell-and-tube heat exchanger that is used to channel tube-side flow in a multipass heat exchanger.Condenser—a shell-and-tube heat exchanger used to cool and condense hot vapors.Conduction—the means of heat transfer through a solid, nonporous material resulting from molecular vibration. Conduction can also occur between closely packed molecules.Convection—the means of heat transfer in fluids resulting from currents. Counterflow—refers to the movement of two flow streams in opposite directions; also called countercurrent flow.Crossflow—refers to the movement of two flow streams perpendicular to each other.Differential pressure—the difference between inlet and outlet pressures; represented as ΔP, or delta p.Differential temperature—the difference between inlet and outlet temperature; represented as ΔT, or delta t.Fixed head—a term applied to a shell-and-tube heat exchanger that has the tube sheet firmly attached to the shell.Floating head—a term applied to a tube sheet on a heat exchanger that is not firmly attached to the shell on the return head and is designed to expand (float) inside the shell as temperature rises. Fouling—buildup on the internal surfaces of devices such as cooling towers and heat exchangers, resulting in reduced heat transfer and plugging.Kettle reboiler—a shell-and-tube heat exchanger with a vapor disengaging cavity, used to supply heat for separation of lighter and heavier components in a distillation system and to maintain heat balance. Laminar flow—streamline flow that is more or less unbroken; layers of liquid flowing in a parallel path.Multipass heat exchanger—a type of shell-and-tube heat exchanger that channels the tubeside flow across the tube bundle (heating source) more than once.Parallel flow—refers to the movement of two flow streams in the same direction; for example, tube-side flow and shell-side flow in a heat exchanger; also called concurrent.Radiant heat transfer—conveyance of heat by electromagnetic waves from a source to receivers.Reboiler—a heat exchanger used to add heat to a liquid that was onceboiling until the liquid boils again.Sensible heat—heat that can be measured or sensed by a change in temperature.Shell-and-tube heat exchanger—a heat exchanger that has a cylindrical shell surrounding a tube bundle.Shell side—refers to flow around the outside of the tubes of ashell-and-tube heat exchanger. See also Tube side.Thermosyphon reboiler—a type of heat exchanger that generates natural circulation as a static liquid is heated to its boiling point.Tube sheet—a flat plate to which the ends of the tubes in a heat exchanger are fixed by rolling, welding, or both.Tube side—refers to flow through the tubes of a shell-and-tube heat exchanger; see Shell side.Turbulent flow—random movement or mixing in swirls and eddies of a fluid. Types of Heat Exchangers换热器的类型Heat transfer is an important function of many industrial processes. Heat exchangers are widely used to transfer heat from one process to another.A heat exchanger allows a hot fluid to transfer heat energy to a cooler fluid through conduction and convection. A heat exchanger provides heating or cooling to a process. A wide array of heat exchangers has been designed and manufactured for use in the chemical processing industry. In pipe coil exchangers, pipe coils are submerged in water or sprayed with water to transfer heat. This type of operation has a low heat transfer coefficient and requires a lot of space. It is best suited for condensing vapors with low heat loads.The double-pipe heat exchanger incorporates a tube-within-a-tube design. It can be found with plain or externally finned tubes. Double-pipe heat exchangers are typically used in series-flow operations in high-pressure applications up to 500 psig shell side and 5,000 psig tube side.A shell-and-tube heat exchanger has a cylindrical shell that surrounds a tube bundle. Fluid flow through the exchanger is referred to as tubeside flow or shell-side flow. A series of baffles support the tubes, direct fluid flow, increase velocity, decrease tube vibration, protect tubing, and create pressure drops.Shell-and-tube heat exchangers can be classified as fixed head, single pass; fixed head, multipass; floating head, multipass; or U-tube.On a fixed head heat exchanger (Figure 7.1), tube sheets are attached to the shell. Fixed head heat exchangers are designed to handle temperature differentials up to 200°F (93.33°C). Thermal expansion prevents a fixed head heat exchanger from exceeding this differential temperature. It is best suited for condenser or heater operations.Floating head heat exchangers are designed for high temperature differentia is above 200°F (93.33°C).During operation, one tube sheet is fixed and the other “floats” inside the shell.The floatingend is not attached to the shell and is free toexpand.Figure 7.1 Fixed Head Heat ExchangerReboilers are heat exchangers that are used to add heat to a liquid that was once boiling until the liquid boils again. Types commonly used in industry are kettle reboilers and thermosyphon reboilers.Plate-and-frame heat exchangers are composed of thin, alternating metal plates that are designed for hot and cold service. Each plate has an outer gasket that seals each compartment. Plate-and-frame heat exchangers have a cold and hot fluid inlet and outlet. Cold and hot fluid headers are formed inside the plate pack, allowing access from every other plate on the hot and cold sides. This device is best suited for viscous or corrosive fluid slurries. It provides excellent high heat transfer. Plate-and-frame heat exchangers are compact and easy to clean. Operating limits of 350 to 500°F (176.66°C to 260°C) are designed to protect the internal gasket. Because of the design specification, plate-and-frame heat exchangers are not suited for boiling and condensing. Most industrial processes use this design in liquid-liquid service.Air-cooled heat exchangers do not require the use of a shell in operation. Process tubes are connected to an inlet and a return header box. The tubes can be finned or plain. A fan is used to push or pull outside air over the exposed tubes. Air-cooled heat exchangers are primarily used in condensing operations where a high level of heat transfer is required.Spiral heat exchangers are characterized by a compact concentric design that generates high fluid turbulence in the process medium. As do otherexchangers, the spiral heat exchanger has cold-medium inlet and outlet and a hot-medium inlet and outlet. Internal surface area provides the conductive transfer element. Spiral heat exchangers have two internal chambers.The Tubular Exchanger Manufacturers Association (TEMA) classifies heat exchangers by a variety of design specifications including American Society of Mechanical Engineers (ASME) construction code, tolerances, and mechanical design:●Class B, Designed for general-purpose operation (economy and compactdesign)●Class C. Designed for moderate service and general-purpose operation(economy and compact design)●Class R. Designed for severe conditions (safety and durability) Heat Transfer and Fluid FlowThe methods of heat transfer are conduction, convection, and radiant heat transfer (Figure 7.2). In the petrochemical, refinery, and laboratory environments, these methods need to be understood well. A combination of conduction and convection heat transfer processes can be found in all heat exchangers. The best conditions for heat transfer are large temperature differences between the products being heated and cooled (the higher the temperature difference, the greater the heat transfer), high heating or coolant flow rates, and a large cross-sectional area of the exchanger.ConductionHeat energy is transferred through solid objects such as tubes, heads,baffles, plates, fins, and shell, by conduction. This process occurs when the molecules that make up the solid matrix begin to absorb heat energy from a hotter source. Since the molecules are in a fixed matrix and cannot move, they begin to vibrate and, in so doing, transfer the energy from the hot side to the cooler side.ConvectionConvection occurs in fluids when warmer molecules move toward cooler molecules. The movement of the molecules sets up currents in the fluid that redistribute heat energy. This process will continue until the energy is distributed equally. In a heat exchanger, this process occurs in the moving fluid media as they pass by each other in the exchanger. Baffle arrangements and flow direction will determine how this convective process will occur in the various sections of the exchanger.Radiant Heat TransferThe best example of radiant heat is the sun’s warming of the earth. The sun’s heat is conveyed by electromagnetic waves. Radiant heat transfer is a line-of-sight process, so the position of the source and that of the receiver are important. Radiant heat transfer is not used in a heat exchanger.Laminar and Turbulent FlowTwo major classifications of fluid flow are laminar and turbulent (Figure 7.3). Laminar—or streamline—flow moves through a system in thin cylindrical layers of liquid flowing in parallel fashion. This type of flow will have little if any turbulence (swirling or eddying) in it. Laminar flow usually exists atlow flow rates. As flow rates increase, the laminar flow pattern changes into a turbulent flow pattern. Turbulent flow is the random movement or mixing of fluids. Once the turbulent flow is initiated, molecular activity speeds up until the fluid is uniformly turbulent.Turbulent flow allows molecules of fluid to mix and absorb heat more readily than does laminar flow. Laminar flow promotes the development of static film, which acts as an insulator. Turbulent flow decreases the thickness of static film, increasing the rate of heat transfer. Parallel and Series FlowHeat exchangers can be connected in a variety of ways. The two most common are series and parallel (Figure 7.4). In series flow (Figure 7.5), the tube-side flow in a multipass heat exchanger is discharged into the tubeside flow of the second exchanger. This discharge route could be switched to shell side or tube side depending on how the exchanger is in service. The guiding principle is that the flow passes through one exchanger before it goes to another. In parallel flow, the process flow goes through multiple exchangers at the same time.Figure 7.5 Series Flow Heat ExchangersHeat Exchanger EffectivenessThe design of an exchanger usually dictates how effectively it can transfer heat energy. Fouling is one problem that stops an exchanger’s ability to transfer heat. During continual service, heat exchangers do not remain clean. Dirt, scale, and process deposits combine with heat to form restrictions inside an exchanger. These deposits on the walls of the exchanger resist the flow that tends to remove heat and stop heat conduction by i nsulating the inner walls. An exchanger’s fouling resistance depends on the type of fluid being handled, the amount and type of suspended solids in the system, the exchanger’s susceptibility to thermal decomposition, and the velocity and temperature of the fluid stream. Fouling can be reduced by increasing fluid velocity and lowering the temperature. Fouling is often tracked and identified usingcheck-lists that collect tube inlet and outlet pressures, and shell inlet and outlet pressures. This data can be used to calculate the pressure differential or Δp. Differential pressure is the difference between inlet and outlet pressures; represented as ΔP, or delta p. Corrosion and erosion are other problems found in exchangers. Chemical products, heat, fluid flow, and time tend to wear down the inner components of an exchanger. Chemical inhibitors are added to avoid corrosion and fouling. These inhibitors are designed to minimize corrosion, algae growth, and mineral deposits.Double-Pipe Heat ExchangerA simple design for heat transfer is found in a double-pipe heat exchanger.A double-pipe exchanger has a pipe inside a pipe (Figure 7.6). The outside pipe provides the shell, and the inner pipe provides the tube. The warm and cool fluids can run in the same direction (parallel flow) or in opposite directions (counterflow or countercurrent).Flow direction is usually countercurrent because it is more efficient. This efficiency comes from the turbulent, against-the-grain, stripping effect of the opposing currents. Even though the two liquid streams never come into physical contact with each other, the two heat energy streams (cold and hot) do encounter each other. Energy-laced, convective currents mix within each pipe, distributing the heat.In a parallel flow exchanger, the exit temperature of one fluid can only approach the exit temperature of the other fluid. In a countercurrent flowexchanger, the exit temperature of one fluid can approach the inlet temperature of the other fluid. Less heat will be transferred in a parallel flow exchanger because of this reduction in temperature difference. Static films produced against the piping limit heat transfer by acting like insulating barriers.The liquid close to the pipe is hot, and the liquid farthest away from the pipe is cooler. Any type of turbulent effect would tend to break up the static film and transfer heat energy by swirling it around the chamber. Parallel flow is not conducive to the creation of turbulent eddies. One of the system limitations of double-pipe heat exchangers is the flow rate they can handle. Typically, flow rates are very low in a double-pipe heat exchanger, and low flow rates are conducive to laminar flow. Hairpin Heat ExchangersThe chemical processing industry commonly uses hairpin heat exchangers (Figure 7.7). Hairpin exchangers use two basic modes: double-pipe and multipipe design. Hairpins are typically rated at 500 psig shell side and 5,000 psig tube side. The exchanger takes its name from its unusual hairpin shape. The double-pipe design consists of a pipe within a pipe. Fins can be added to the internal tube’s external wall to increase heat transfer. The multipipe hairpin resembles a typical shell-and-tube heat exchanger, stretched and bent into a hairpin.The hairpin design has several advantages and disadvantages. Among its advantages are its excellent capacity for thermal expansion because of its U-tube type shape; its finned design, which works well with fluids that have a low heat transfer coefficient; and its high pressure on the tube side. In addition, it is easy to install and clean; its modular design makes it easy to add new sections; and replacement parts are inexpensive and always in supply. Among its disadvantages are the facts that it is not as cost effective as most shell-and-tube exchangers and it requires special gaskets.Shell-and-Tube Heat ExchangersThe shell-and-tube heat exchanger is the most common style found inindustry. Shell-and-tube heat exchangers are designed to handle high flow rates in continuous operations. Tube arrangement can vary, depending on the process and the amount of heat transfer required. As the tube-side flow enters the exchanger—or “head”—flow is directed into tubes that run parallel to each other. These tubes run through a shell that has a fluid passing through it. Heat energy is transferred through the tube wall into the cooler fluid. Heat transfer occurs primarily through conduction (first) and convection (second). Figure 7.8 shows a fixed head,single-pass heat exchanger.Fluid flow into and out of the heat exchanger is designed for specific liquid–vapor services. Liquids move from the bottom of the device to the top to remove or reduce trapped vapor in the system. Gases move from top to bottom to remove trapped or accumulated liquids. This standard applies to both tube-side and shell-side flow.Plate-and-Frame Heat ExchangersPlate-and-frame heat exchangers are high heat transfer and high pressure drop devices. They consist of a series of gasketed plates, sandwiched together by two end plates and compression bolts (Figures 7.20 and 7.21). The channels between the plates are designed to create pressure drop and turbulent flow so high heat transfer coefficients can be achieved.The openings on the plate exchanger are located typically on one of the fixed-end covers.As hot fluid enters the hot inlet port on the fixed-end cover, it is directed into alternating plate sections by a common discharge header. The header runs the entire length of the upper plates. As cold fluid enters the countercurrent cold inlet port on the fixed-end cover, it is directed into alternating plate sections. Cold fluid moves up the plates while hot fluid drops down across the plates. The thin plates separate the hot and cold liquids, preventing leakage. Fluid flow passes across the plates one time before entering the collection header. The plates are designed with an alternating series of chambers. Heat energy is transferred through the walls of the plates by conduction and into the liquid by convection. The hot and cold inlet lines run the entire length of the plate heater and function like a distribution header. The hot and cold collection headers run parallel and on the opposite side of the plates from each other. The hot fluid header that passes through the gasketed plate heat exchanger is located in the top. This arrangement accounts for the pressure drop and turbulent flow as fluid drops over the plates and into the collection header. Cold fluid enters the bottom of the gasketed plate heat exchanger and travels countercurrent to the hot fluid. The cold fluid collection header is located in the upper section of the exchanger.Plate-and-frame heat exchangers have several advantages and disadvantages. They are easy to disassemble and clean and distribute heat evenly so there are no hot spots. Plates can easily be added or removed. Other advantages of plate-and-frame heat exchangers are their low fluid resistance time, low fouling, and high heat transfer coefficient. In addition, if gaskets leak, they leak to the outside, and gaskets are easy to replace.The plates prevent cross-contamination of products. Plate-and-frame heat exchangers provide high turbulence and a large pressure drop and are small compared with shell-and-tube heat exchangers.Disadvantages of plate-and-frame heat exchangers are that they have high-pressure and high-temperature limitations. Gaskets are easily damaged and may not be compatible with process fluids.Spiral Heat ExchangersSpiral heat exchangers are characterized by a compact concentric design that generates high fluid turbulence in the process medium (Figure 7.22). This type of heat exchanger comes in two basic types: (1) spiral flow on both sides and (2) spiral flow–crossflow. Type 1 spiral exchangers are used in liquid-liquid, condenser, and gas cooler service. Fluid flow into the exchanger is designed for full counterflow operation. The horizontal axial installation provides excellent self-cleaning of suspended solids.Type 2 spiral heat exchangers are designed for use as condensers, gas coolers, heaters, and reboilers. The vertical installation makes it an excellent choice for combining high liquid velocity and low pressure drop on the vapor-mixture side. Type 2 spirals can be used in liquid-liquid systems where high flow rates on one side are offset by low flow rates on the other.Air-Cooled Heat ExchangersA different approach to heat transfer occurs in the fin fan or air-cooled heat exchanger. Air-cooled heat exchangers provide a structured matrix of plain or finned tubes connected to an inlet and return header (Figure 7.23). Air is used as the outside medium to transfer heat away from the tubes. Fans are used in a variety of arrangements to apply forced convection for heattransfer coefficients. Fans can be mounted above or below the tubes in forced-draft or induced-draft arrangements. Tubes can be installed vertically or horizontally.The headers on an air-cooled heat exchanger can be classified as cast box, welded box, cover plate, or manifold. Cast box and welded box types have plugs on the end plate for each tube. This design provides access for cleaning individual tubes, plugging them if a leak is found, and rerolling to tighten tube joints. Cover plate designs provide easy access to all of the tubes. A gasket is used between the cover plate and head. The manifold type is designed for high-pressure applications.Mechanical fans use a variety of drivers. Common drivers found in service with air-cooled heat exchangers include electric motor and reduction gears, steam turbine or gas engine, belt drives, and hydraulic motors. The fan blades are composed of aluminum or plastic. Aluminum blades are d esigned to operate in temperatures up to 300°F (148.88°C), whereas plastic blades are limited to air temperatures between 160°F and 180°F(71.11°C, 82.22°C).Air-cooled heat exchangers can be found in service on air compressors, in recirculation systems, and in condensing operations. This type of heat transfer device provides a 40°F (4.44°C) temperature differential between the ambient air and the exiting process fluid.Air-cooled heat exchangers have none of the problems associated with water such as fouling or corrosion. They are simple to construct and cheaper to maintain than water-cooled exchangers. They have low operating costs and superior high temperature removal (above 200°F or 93.33°C). Their disadvantages are that they are limited to liquid or condensing service and have a high outlet fluid temperature and high initial cost of equipment. In addition, they are susceptible to fire or explosion in cases of loss of containment.。

换热器文献翻译之英文部分

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 +=。

换热器的书籍

换热器的书籍换热器（Heat Exchanger）是化工、制冷、暖通空调等行业中用于热量交换的关键设备。

关于换热器设计、操作和维护的专业书籍有很多，以下是一些推荐的书籍：1. 《换热器技术手册》（Handbook of Heat Transfer Equipment）- Michael Jensen, Tusher Ghosh & Sreedhar Kodituwakku- 这本综合性手册详细介绍了各种类型的换热器设计和工程应用。

2. 《换热器设计手册》（Perry's Chemical Engineers' Handbook）- Robert H. Perry- 这是一本经典的化学工程手册，其中包括了换热器设计的章节，提供了大量的设计数据和计算方法。

3. 《换热器设计基础》（Fundamentals of Heat Exchanger Design）- John Zimmerman- 这本书适合初学者，系统地介绍了换热器设计的基本概念和步骤。

4. 《传热》（Heat Transfer: A Practical Approach）- Yun Wang- 虽然这本书不专门针对换热器设计，但它提供了传热学的基础知识，这是进行换热器设计的必备知识。

5. 《制冷与空调装置中的换热器》（Heat Exchangers for Refrigeration and Air Conditioning）- Alberto Lamberti- 这本书专注于制冷和空调领域的换热器应用，讨论了相关的设计和优化问题。

6. 《换热器分析与设计》（Analysis and Design of Heat Exchangers）- N. P. Choudhuri- 本书提供了换热器分析和设计的全面指导，包括理论和实践两方面的内容。

7. 《换热器维修与故障排除手册》（Heat Exchanger Maintenance and Troubleshooting Guide）- David F. Van Gorp- 这本手册专注于换热器的维护和故障诊断，对于保持设备有效运行非常有帮助。

换热器外文翻译 (2)

Heat ExchangersKey Terms Baffles—evenly spaced partitions in a shell and tube heat exchanger that support the tubes, prevent vibration, control fluid velocity and direction, increase turbulent flow, and reduce hot spots. Channel head—a device mounted on the inlet side of a shell-and-tube heat exchanger that is used to channel tube-side flow in a multipass heat exchanger.Condenser—a shell-and-tube heat exchanger used to cool and condense hot vapors.Conduction—the means of heat transfer through a solid, nonporous material resulting from molecular vibration. Conduction can also occur between closely packed molecules.Convection—the means of heat transfer in fluids resulting from currents. Counterflow—refers to the movement of two flow streams in opposite directions; also called countercurrent flow.Crossflow—refers to the movement of two flow streams perpendicular to each other.Differential pressure—the difference between inlet and outlet pressures; represented as ΔP, or delta p.Differential temperature—the difference between inlet and outlet temperature; represented as ΔT, or delta t.Fixed head—a term applied to a shell-and-tube heat exchanger that has the tube sheet firmly attached to the shell.Floating head—a term applied to a tube sheet on a heat exchanger that is not firmly attached to the shell on the return head and is designed to expand (float) inside the shell as temperature rises. Fouling—buildup on the internal surfaces of devices such as cooling towers and heat exchangers, resulting in reduced heat transfer and plugging.Kettle reboiler—a shell-and-tube heat exchanger with a vapor disengaging cavity, used to supply heat for separation of lighter and heavier components in a distillation system and to maintain heat balance. Laminar flow—streamline flow that is more or less unbroken; layers of liquid flowing in a parallel path.Multipass heat exchanger—a type of shell-and-tube heat exchanger that channels the tubeside flow across the tube bundle (heating source) more than once.Parallel flow—refers to the movement of two flow streams in the same direction; for example, tube-side flow and shell-side flow in a heat exchanger; also called concurrent.Radiant heat transfer—conveyance of heat by electromagnetic waves from a source to receivers.Reboiler—a heat exchanger used to add heat to a liquid that was onceboiling until the liquid boils again.Sensible heat—heat that can be measured or sensed by a change in temperature.Shell-and-tube heat exchanger—a heat exchanger that has a cylindrical shell surrounding a tube bundle.Shell side—refers to flow around the outside of the tubes of ashell-and-tube heat exchanger. See also Tube side.Thermosyphon reboiler—a type of heat exchanger that generates natural circulation as a static liquid is heated to its boiling point.Tube sheet—a flat plate to which the ends of the tubes in a heat exchanger are fixed by rolling, welding, or both.Tube side—refers to flow through the tubes of a shell-and-tube heat exchanger; see Shell side.Turbulent flow—random movement or mixing in swirls and eddies of a fluid. Types of Heat Exchangers换热器的类型Heat transfer is an important function of many industrial processes. Heat exchangers are widely used to transfer heat from one process to another.A heat exchanger allows a hot fluid to transfer heat energy to a cooler fluid through conduction and convection. A heat exchanger provides heating or cooling to a process. A wide array of heat exchangers has been designed and manufactured for use in the chemical processing industry. In pipe coil exchangers, pipe coils are submerged in water or sprayed with water to transfer heat. This type of operation has a low heat transfer coefficient and requires a lot of space. It is best suited for condensing vapors with low heat loads.The double-pipe heat exchanger incorporates a tube-within-a-tube design. It can be found with plain or externally finned tubes. Double-pipe heat exchangers are typically used in series-flow operations in high-pressure applications up to 500 psig shell side and 5,000 psig tube side.A shell-and-tube heat exchanger has a cylindrical shell that surrounds a tube bundle. Fluid flow through the exchanger is referred to as tubeside flow or shell-side flow. A series of baffles support the tubes, direct fluid flow, increase velocity, decrease tube vibration, protect tubing, and create pressure drops.Shell-and-tube heat exchangers can be classified as fixed head, single pass; fixed head, multipass; floating head, multipass; or U-tube.On a fixed head heat exchanger (Figure 7.1), tube sheets are attached to the shell. Fixed head heat exchangers are designed to handle temperature differentials up to 200°F (93.33°C). Thermal expansion prevents a fixed head heat exchanger from exceeding this differential temperature. It is best suited for condenser or heater operations.Floating head heat exchangers are designed for high temperature differentia is above 200°F (93.33°C).During operation, one tube sheet is fixed and the other “floats” inside the shell.The floatingend is not attached to the shell and is free toexpand.Figure 7.1 Fixed Head Heat ExchangerReboilers are heat exchangers that are used to add heat to a liquid that was once boiling until the liquid boils again. Types commonly used in industry are kettle reboilers and thermosyphon reboilers.Plate-and-frame heat exchangers are composed of thin, alternating metal plates that are designed for hot and cold service. Each plate has an outer gasket that seals each compartment. Plate-and-frame heat exchangers have a cold and hot fluid inlet and outlet. Cold and hot fluid headers are formed inside the plate pack, allowing access from every other plate on the hot and cold sides. This device is best suited for viscous or corrosive fluid slurries. It provides excellent high heat transfer. Plate-and-frame heat exchangers are compact and easy to clean. Operating limits of 350 to 500°F (176.66°C to 260°C) are designed to protect the internal gasket. Because of the design specification, plate-and-frame heat exchangers are not suited for boiling and condensing. Most industrial processes use this design in liquid-liquid service.Air-cooled heat exchangers do not require the use of a shell in operation. Process tubes are connected to an inlet and a return header box. The tubes can be finned or plain. A fan is used to push or pull outside air over the exposed tubes. Air-cooled heat exchangers are primarily used in condensing operations where a high level of heat transfer is required.Spiral heat exchangers are characterized by a compact concentric design that generates high fluid turbulence in the process medium. As do otherexchangers, the spiral heat exchanger has cold-medium inlet and outlet and a hot-medium inlet and outlet. Internal surface area provides the conductive transfer element. Spiral heat exchangers have two internal chambers.The Tubular Exchanger Manufacturers Association (TEMA) classifies heat exchangers by a variety of design specifications including American Society of Mechanical Engineers (ASME) construction code, tolerances, and mechanical design:●Class B, Designed for general-purpose operation (economy and compactdesign)●Class C. Designed for moderate service and general-purpose operation(economy and compact design)●Class R. Designed for severe conditions (safety and durability) Heat Transfer and Fluid FlowThe methods of heat transfer are conduction, convection, and radiant heat transfer (Figure 7.2). In the petrochemical, refinery, and laboratory environments, these methods need to be understood well. A combination of conduction and convection heat transfer processes can be found in all heat exchangers. The best conditions for heat transfer are large temperature differences between the products being heated and cooled (the higher the temperature difference, the greater the heat transfer), high heating or coolant flow rates, and a large cross-sectional area of the exchanger.ConductionHeat energy is transferred through solid objects such as tubes, heads,baffles, plates, fins, and shell, by conduction. This process occurs when the molecules that make up the solid matrix begin to absorb heat energy from a hotter source. Since the molecules are in a fixed matrix and cannot move, they begin to vibrate and, in so doing, transfer the energy from the hot side to the cooler side.ConvectionConvection occurs in fluids when warmer molecules move toward cooler molecules. The movement of the molecules sets up currents in the fluid that redistribute heat energy. This process will continue until the energy is distributed equally. In a heat exchanger, this process occurs in the moving fluid media as they pass by each other in the exchanger. Baffle arrangements and flow direction will determine how this convective process will occur in the various sections of the exchanger.Radiant Heat TransferThe best example of radiant heat is the sun’s warming of the earth. The sun’s heat is conveyed by electromagnetic waves. Radiant heat transfer is a line-of-sight process, so the position of the source and that of the receiver are important. Radiant heat transfer is not used in a heat exchanger.Laminar and Turbulent FlowTwo major classifications of fluid flow are laminar and turbulent (Figure 7.3). Laminar—or streamline—flow moves through a system in thin cylindrical layers of liquid flowing in parallel fashion. This type of flow will have little if any turbulence (swirling or eddying) in it. Laminar flow usually exists atlow flow rates. As flow rates increase, the laminar flow pattern changes into a turbulent flow pattern. Turbulent flow is the random movement or mixing of fluids. Once the turbulent flow is initiated, molecular activity speeds up until the fluid is uniformly turbulent.Turbulent flow allows molecules of fluid to mix and absorb heat more readily than does laminar flow. Laminar flow promotes the development of static film, which acts as an insulator. Turbulent flow decreases the thickness of static film, increasing the rate of heat transfer. Parallel and Series FlowHeat exchangers can be connected in a variety of ways. The two most common are series and parallel (Figure 7.4). In series flow (Figure 7.5), the tube-side flow in a multipass heat exchanger is discharged into the tubeside flow of the second exchanger. This discharge route could be switched to shell side or tube side depending on how the exchanger is in service. The guiding principle is that the flow passes through one exchanger before it goes to another. In parallel flow, the process flow goes through multiple exchangers at the same time.Figure 7.5 Series Flow Heat ExchangersHeat Exchanger EffectivenessThe design of an exchanger usually dictates how effectively it can transfer heat energy. Fouling is one problem that stops an exchanger’s ability to transfer heat. During continual service, heat exchangers do not remain clean. Dirt, scale, and process deposits combine with heat to form restrictions inside an exchanger. These deposits on the walls of the exchanger resist the flow that tends to remove heat and stop heat conduction by i nsulating the inner walls. An exchanger’s fouling resistance depends on the type of fluid being handled, the amount and type of suspended solids in the system, the exchanger’s susceptibility to thermal decomposition, and the velocity and temperature of the fluid stream. Fouling can be reduced by increasing fluid velocity and lowering the temperature. Fouling is often tracked and identified usingcheck-lists that collect tube inlet and outlet pressures, and shell inlet and outlet pressures. This data can be used to calculate the pressure differential or Δp. Differential pressure is the difference between inlet and outlet pressures; represented as ΔP, or delta p. Corrosion and erosion are other problems found in exchangers. Chemical products, heat, fluid flow, and time tend to wear down the inner components of an exchanger. Chemical inhibitors are added to avoid corrosion and fouling. These inhibitors are designed to minimize corrosion, algae growth, and mineral deposits.Double-Pipe Heat ExchangerA simple design for heat transfer is found in a double-pipe heat exchanger.A double-pipe exchanger has a pipe inside a pipe (Figure 7.6). The outside pipe provides the shell, and the inner pipe provides the tube. The warm and cool fluids can run in the same direction (parallel flow) or in opposite directions (counterflow or countercurrent).Flow direction is usually countercurrent because it is more efficient. This efficiency comes from the turbulent, against-the-grain, stripping effect of the opposing currents. Even though the two liquid streams never come into physical contact with each other, the two heat energy streams (cold and hot) do encounter each other. Energy-laced, convective currents mix within each pipe, distributing the heat.In a parallel flow exchanger, the exit temperature of one fluid can only approach the exit temperature of the other fluid. In a countercurrent flowexchanger, the exit temperature of one fluid can approach the inlet temperature of the other fluid. Less heat will be transferred in a parallel flow exchanger because of this reduction in temperature difference. Static films produced against the piping limit heat transfer by acting like insulating barriers.The liquid close to the pipe is hot, and the liquid farthest away from the pipe is cooler. Any type of turbulent effect would tend to break up the static film and transfer heat energy by swirling it around the chamber. Parallel flow is not conducive to the creation of turbulent eddies. One of the system limitations of double-pipe heat exchangers is the flow rate they can handle. Typically, flow rates are very low in a double-pipe heat exchanger, and low flow rates are conducive to laminar flow. Hairpin Heat ExchangersThe chemical processing industry commonly uses hairpin heat exchangers (Figure 7.7). Hairpin exchangers use two basic modes: double-pipe and multipipe design. Hairpins are typically rated at 500 psig shell side and 5,000 psig tube side. The exchanger takes its name from its unusual hairpin shape. The double-pipe design consists of a pipe within a pipe. Fins can be added to the internal tube’s external wall to increase heat transfer. The multipipe hairpin resembles a typical shell-and-tube heat exchanger, stretched and bent into a hairpin.The hairpin design has several advantages and disadvantages. Among its advantages are its excellent capacity for thermal expansion because of its U-tube type shape; its finned design, which works well with fluids that have a low heat transfer coefficient; and its high pressure on the tube side. In addition, it is easy to install and clean; its modular design makes it easy to add new sections; and replacement parts are inexpensive and always in supply. Among its disadvantages are the facts that it is not as cost effective as most shell-and-tube exchangers and it requires special gaskets.Shell-and-Tube Heat ExchangersThe shell-and-tube heat exchanger is the most common style found inindustry. Shell-and-tube heat exchangers are designed to handle high flow rates in continuous operations. Tube arrangement can vary, depending on the process and the amount of heat transfer required. As the tube-side flow enters the exchanger—or “head”—flow is directed into tubes that run parallel to each other. These tubes run through a shell that has a fluid passing through it. Heat energy is transferred through the tube wall into the cooler fluid. Heat transfer occurs primarily through conduction (first) and convection (second). Figure 7.8 shows a fixed head,single-pass heat exchanger.Fluid flow into and out of the heat exchanger is designed for specific liquid–vapor services. Liquids move from the bottom of the device to the top to remove or reduce trapped vapor in the system. Gases move from top to bottom to remove trapped or accumulated liquids. This standard applies to both tube-side and shell-side flow.Plate-and-Frame Heat ExchangersPlate-and-frame heat exchangers are high heat transfer and high pressure drop devices. They consist of a series of gasketed plates, sandwiched together by two end plates and compression bolts (Figures 7.20 and 7.21). The channels between the plates are designed to create pressure drop and turbulent flow so high heat transfer coefficients can be achieved.The openings on the plate exchanger are located typically on one of the fixed-end covers.As hot fluid enters the hot inlet port on the fixed-end cover, it is directed into alternating plate sections by a common discharge header. The header runs the entire length of the upper plates. As cold fluid enters the countercurrent cold inlet port on the fixed-end cover, it is directed into alternating plate sections. Cold fluid moves up the plates while hot fluid drops down across the plates. The thin plates separate the hot and cold liquids, preventing leakage. Fluid flow passes across the plates one time before entering the collection header. The plates are designed with an alternating series of chambers. Heat energy is transferred through the walls of the plates by conduction and into the liquid by convection. The hot and cold inlet lines run the entire length of the plate heater and function like a distribution header. The hot and cold collection headers run parallel and on the opposite side of the plates from each other. The hot fluid header that passes through the gasketed plate heat exchanger is located in the top. This arrangement accounts for the pressure drop and turbulent flow as fluid drops over the plates and into the collection header. Cold fluid enters the bottom of the gasketed plate heat exchanger and travels countercurrent to the hot fluid. The cold fluid collection header is located in the upper section of the exchanger.Plate-and-frame heat exchangers have several advantages and disadvantages. They are easy to disassemble and clean and distribute heat evenly so there are no hot spots. Plates can easily be added or removed. Other advantages of plate-and-frame heat exchangers are their low fluid resistance time, low fouling, and high heat transfer coefficient. In addition, if gaskets leak, they leak to the outside, and gaskets are easy to replace.The plates prevent cross-contamination of products. Plate-and-frame heat exchangers provide high turbulence and a large pressure drop and are small compared with shell-and-tube heat exchangers.Disadvantages of plate-and-frame heat exchangers are that they have high-pressure and high-temperature limitations. Gaskets are easily damaged and may not be compatible with process fluids.Spiral Heat ExchangersSpiral heat exchangers are characterized by a compact concentric design that generates high fluid turbulence in the process medium (Figure 7.22). This type of heat exchanger comes in two basic types: (1) spiral flow on both sides and (2) spiral flow–crossflow. Type 1 spiral exchangers are used in liquid-liquid, condenser, and gas cooler service. Fluid flow into the exchanger is designed for full counterflow operation. The horizontal axial installation provides excellent self-cleaning of suspended solids.Type 2 spiral heat exchangers are designed for use as condensers, gas coolers, heaters, and reboilers. The vertical installation makes it an excellent choice for combining high liquid velocity and low pressure drop on the vapor-mixture side. Type 2 spirals can be used in liquid-liquid systems where high flow rates on one side are offset by low flow rates on the other.Air-Cooled Heat ExchangersA different approach to heat transfer occurs in the fin fan or air-cooled heat exchanger. Air-cooled heat exchangers provide a structured matrix of plain or finned tubes connected to an inlet and return header (Figure 7.23). Air is used as the outside medium to transfer heat away from the tubes. Fans are used in a variety of arrangements to apply forced convection for heattransfer coefficients. Fans can be mounted above or below the tubes in forced-draft or induced-draft arrangements. Tubes can be installed vertically or horizontally.The headers on an air-cooled heat exchanger can be classified as cast box, welded box, cover plate, or manifold. Cast box and welded box types have plugs on the end plate for each tube. This design provides access for cleaning individual tubes, plugging them if a leak is found, and rerolling to tighten tube joints. Cover plate designs provide easy access to all of the tubes. A gasket is used between the cover plate and head. The manifold type is designed for high-pressure applications.Mechanical fans use a variety of drivers. Common drivers found in service with air-cooled heat exchangers include electric motor and reduction gears, steam turbine or gas engine, belt drives, and hydraulic motors. The fan blades are composed of aluminum or plastic. Aluminum blades are d esigned to operate in temperatures up to 300°F (148.88°C), whereas plastic blades are limited to air temperatures between 160°F and 180°F(71.11°C, 82.22°C).Air-cooled heat exchangers can be found in service on air compressors, in recirculation systems, and in condensing operations. This type of heat transfer device provides a 40°F (4.44°C) temperature differential between the ambient air and the exiting process fluid.Air-cooled heat exchangers have none of the problems associated with water such as fouling or corrosion. They are simple to construct and cheaper to maintain than water-cooled exchangers. They have low operating costs and superior high temperature removal (above 200°F or 93.33°C). Their disadvantages are that they are limited to liquid or condensing service and have a high outlet fluid temperature and high initial cost of equipment. In addition, they are susceptible to fire or explosion in cases of loss of containment.。

换热器英文文献

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.。