Effects of Chevron Angle on Thermal Performance of Corrugated Plate Heat Exchanger

doi: 10.14355/ijepr.014.0301.0 Effects of Chevron Angle on Thermal Performance of Corrugated Plate Heat Exchanger Masoud Asadi, Ramin Haghighi Khoshkhoo Department of Mechanical Engineering, Azad Islamic University Science and Research branch, Tehran, Iran Department of Mechanical Engineering, Power and Water University of Technology, Tehran, Iran Masoud471@gmail.com, khsohkhoo@pwut.ac.ir Abstract There are two types of plate heat exchangers;one of which is the plate and frame heat exchanger developed in the 1930s for food industries because of easy cleaning and quality control; while the other is the brazed plate heat exchanger that usually allows higher pressures and temperatures. This paper presents a study on chevron angle effects. Here, thermal-hydraulic performances are evaluated in terms of the heat transfer coefficient, friction factor and pressure drop. Results denote that friction factor will rise with the growing chevron angle. On the other hands, when the chevron angle has a optimal value, the results demonstrate that friction factor has inverse relationship with mass flow rate for both Laminar and Turbulent regime. Eventually, it is indicated that the optimal chevron angle is 60. Keywords Chevron Angle; Corrugated Plate Heat Exchanger; Friction Factor; Heat Transfer Coefficient Introduction The Plate Heat Exchanger(PHE) consists of a series of parallel plates that are corrugated both to increase turbulence and to give mechanical rigidity. They normally have flow ports in all four corners and are clamped together in a frame that carries bushes and nuzzles lined up with the plate ports and connected to the external pipework that carries the two liquid streams. TABLE 1 STANDARD PERFORMANCE LIMITS Maximum Operating Pressure 5 bar Maximum Operating Temperature 160 o C Maximum Operating Flow rate 3600 m 3 / hr Heat Transfer Coefficient 3500-7000 W / m. C Heat Transfer Area 0.1-00 m NTU 0.3-4.0 Heat Recovery As high as 90% The plates themselves are fitted with gaskets which are shaped and located both to prevent external leakage and to direct the two liquids normally counter currently through the relatively narrow passage between alternate pairs of heat transfer plates. Most applications for plate heat exchangers are for heat recovery duties. Table (1) provides standard performance limits for PHE. Here, plates are available in a variety of corrugated or embossed patterns. The basic objective of providing corrugation to the plates is to impart high turbulence to the fluids, which results in a very high heat transfer coefficient compared to those obtainable in a shell and tube heat exchanger for similar duties. These embossing patterns also result in increased effective surface areas and provide additional strength to the plates by means of many contact points over the plates to withstand differential pressure that exists between the adjacent plates. These days, over 60 different plate patterns have been developed worldwide. The pattern and geometry are propriety. The most widely used corrugation types are the Intermating troughs and chevron patterns. Nomenclature A :corrugated area t a :amplitude D :hydraulic diameter h f :friction factor H :height p h :heat transfer coefficient K :thermal conductivity f PHE :plate heat exchanger Re :Reynolds number W :width P Greek symbols µ :dynamic viscosity β :chevron angle φ :surface enlargement factor N :number of channels per each fluid γ : surface waviness c N :number of channels per pass p : wave length N :number of plates t δ : thickness N :number of wavelength per plate In chevron patterns, the corrugation are pressed to the same depth as the plate spacing. The chevron angle is reversed on adjacent plates so that when the plates are clamped together, the corrugation cross one another to 8

International Journal of Engineering Practical Research (IJEPR) Volume 3 Issue 1, February 014 www.seipub.org/ijepr provides numerous contact points. (M.K. Bassiouny 1984) According to this fact that chevron pattern type is the most common type in use today, our objective in this paper studies effects of chevron angle on friction factor and heat transfer coefficient. Thermal Performance FIG. 1 CHEVRON-TYPE PLATE The wavelength of the chevron pattern is the corrugation pitch as shown in Figure (1). The amplitude of the corrugation is denoted as a, where a is the amplitude of the sinusoidal corrugation with the plate thickness δ. The number of wavelength per plate N is calculated by dividing the width W P by the wavelength. W P N = (1) If N t be introduced as total number of plates, then the amplitude a can be expressed in terms of the PHE height, H P : 1 H ( P a = δ ) () Nt + 1 The surface waviness can be essentially represented by two dimensionless parameters, the corrugation aspect ration γ and the surface enlargement factor φ. The corrugation aspect ratio γ is: 4a γ = (3) It is noted that when γ = 0, a flat-parallel plate is obtained. So, γ 0 in indicative of surface area enlargement as well. Rising γ enlarges the surface area, but high γ may induce vortexes at the top and bottom of the channel which can trap the fluids locally and reduce the heat transfer. So, the plate heat exchanger design is commonly limited to γ 1, depending on the Reynolds number. The corrugated area is obtained using the path distance of the sinusoidal corrugation. It is noticeable that the chevron angle ( β ) has not any effects on the corrugated area. Therefore, the enlarged length per wavelength is : πa πx L = 1+ cos dx (4) 0 So, the corrugated area for each fluid can be calculated by At = LNLN P C (5) Where, N c is the number of channels per each fluid. Also, the number of channels per pass is N p, where: Nt + 1 Nc = (6) NP Here, the surface enlargement factor, φ, is expressed as L. N φ = (7) Wp 4a Dh = (8) φ One of the important correlation for the friction factor is provided by Martin et, al. (1996), 0.5 cos β 1 cos β f = + 0.5 (9) ( 0.045 tan β + 0.09sin β + f0 / cos β) 3.8 f1 Where, 16 forre f Re 0 = (10).0 ( 1.56 ln Re 3.0) for Re 149.5 + 0.965 forre Re f1 = (11) 9.75 forre 0.89 Re In this formula, heat transfer coefficient is: 1/6 K f. L. N 0.333 0.374 h 0.05..Pr.( f.re sin ). µ = β (1) 4. awp µ s Where 10 β 80 and K f is considered as thermal conductivity of the fluid. µ s is also dynamic viscosity at the wall temperature. Results To analyze the influence of the chevron angle on the thermal performance of corrugated plate heat 9

exchanger, a case study has been done. In the heat recovery application, a cold water will be heated by wastewater using metioned plate heat exchanger. The cold water with a flow rate of 130 kg/s enters the plate heat exchanger at o C, and it will be heated to 4 o C. The hot wastewater enters at a flow rate of 140 kg/s and 65 o C. The maximum permissible pressure drop for each fluid is 70 Kpa. The chevron plates are singlepass with the chevron angle of 30. The metal of the plates is stainless steel AISI 304.(Table of ()) TABLE DESIGNING INFORMATION Number of passes, 1 Chevron angle 30 Total number of plates 109 Plate thickness 0.6 mm Corrugation pitch 9 mm Port diameter 00 mm Thermal conductivity 14.9 W/m.K Here, we face a sizing problem and then a rating problem to determine T,q,ε and P. show the Geomery properties and outlet temperature and pressure drops for designed heat exchanger, respectively. With increasing N t, the surface area density increases while γ decreases. Consequently, the volume of the heat exchanger decreases with the increasing surface area density or number of plates. Based on Martin et, al. (1996) research, friction factor is a function of, Friction factor = f (Re, β ) (13) As expected, friction factor will increase with growing β (Figure ()). Here, the optimal chevron angle is 55 β 65. However, when the chevron angle is more than 70, this equation shows a significant error, about 10% to 40% dependent on the chevron angle. Therefore, this expression can be safely used when β is less than 70 to predict friction factor and then pressure drop. TABLE 3 THERMAL RESULTS OF DESIGNING PROCESS Hyrualic diameter 5.734 mm Effectiveness 0.465 7 Heat transfer rate 1.087 10 W 5 UA 4.581 10 W/K 4 Convective coefficient for hot stream 1.787 10 W/m.K 4 Convective coefficient for cold stream 1.4 10 W/m.K 4 Reynolds number for hot stream 1.501 10 3 Reynolds number for cold stream 7.64 10 Friction factor for hot stream 0.1 Friction factor for cold stream 0.10 TABLE 4 GEOMETRY RESULTS Number of passes 1 Number of plates 109 Corrugation pitch 9 mm Plate thickness 0.6 mm Chevron angle 30 Port diameter 00 mm Corrugation aspect ratio 0.874 Surface enlargement factor 1.37 3 Surface area density 697.61 m /m Heat transfer area for each fluid 80.81 m Amplitude 1973 mm TABLE 5 OUTLET TEMPERATURE AND PRESSURE DROPS Cold water outlet temperature Hot wastewater outlet temperature Pressure drop for hot wastewater Pressure drop for cold water 4 o C 46.45 o C 70 Kpa 60.857 Kpa After designing process, the thermal and hydraulic quantities are listed in Table (3). Table (4) and (5) also FIG. CHANGES OF FRICTION FACTOR VERSUS CHEVRON ANGLE However, typically, the chevron angle is 60. So, the friction factor is just a function of Reynolds number. Figure (3) shows relationship between friction factor and mass flow rate in Laminar regime. As it is clear that there is a dramatic decrease in friction factor when mass flow rate is between 0 and 40 kg / s. Note that increasing mass flow rate is accompanied with decrease in friction factor. However, when mass flow 10

International Journal of Engineering Practical Research (IJEPR) Volume 3 Issue 1, February 014 www.seipub.org/ijepr rate has a value between 140 to 00 kg / s friction factor is constant approximately. In Laminar regime the typical friction factor is 0.805. same as Laminar regime, there are several differences. The important note is that here friction factor decreases as constantly with growing mass flow rate (Unlike the Laminar regime which is constant when mas flow rate is between 140 to 00 kg / s ). The other difference is that here the typical friction factor is 0.54 (In Laminar flow was 0.805 ) which is reasonable because based on Anthon Cooper and J.Dennis Usher research, the friction factor is only a function of Reynolds number and naturally mass flow rate: 38 Laminar regime Re f = (14) 1. Turbulent regime 0.5 Re In this transition regime, the friction factor relationship can be determined by interpolating between the boundaries of the Turbulent and Laminar flow regimes. FIG. 3 FUNCTION OF FRICTION FACTOR VERSUS MASS FLOW RATE IN LAMINAR REGIME FIG. 4 FUNCTION OF FRICTION FACTOR VERSUS MASS FLOW RATE IN TURBULENT REGIME Also, Figure (4) demonstrates function of friction factor versus mass flow rate in Turbulent regime. Although the trend for the first part of diagram is the FIG. 5 FUNCTION OF HEAT TRANSFER COEFFICIENT VERSUS MASS FLOW RATE Diagram of heat transfer coefficient versus the chevron angle demonstrates a unexpected trend, as shown in Figure (5), where a moderate decrease has been seen after a dramatic increase. This graph again denotes that the chevron angle of 60 is a key angle. In other words, in transition regime, heat transfer coefficient decreases. So based on friction factor and heat transfer coefficient graphs, the optimal chevron angle is 60. 11

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