İklimlendirme cihazı ısıl simülasyon programı için kanatlı borulu ısı değiştiricilerinin modellenmesi - Tuztaş, Metin (Fen Bilimleri Enstitüsü, 1999)
Yapılan bu çalışma ile ısı değiştiricilerinin performans deneylerinin yapılabileceği bir deney düzeneğinin kurulması ve farklı geometrilere sahip ısı değiştiricilerinin deneylerinin yapılarak bu konuda bir bilgi bankası oluşturulması amaçlanmıştır. Kurulan deney düzeneği değişik geometrilere sahip kanatlı borulu ısı değiştiricilerinin ıslak ve kuru yüzey hallerindeki ısı taşınım katsayısı ve sürtünme faktörü değerlerinin bulunabilmesi için hem soğutma hem de ısıtma yapabilecek şekilde tasarlanmıştır. Bu çalışmanın amacı, ARÇELİK A.Ş. klimalarında kullanılan kanatlı borulu ısı değiştiricilerinin performans büyüklüklerinin ölçülmesi ve bu bilgilerin kullanılarak bir matematiksel modelin oluşturulmasıdır. Böylece oluşturulan matematiksel model kullanılarak geometrik özellikleri belirlenmiş bir ısı değiştiricinin performans büyüklüklerinin tespit edilmesi mümkün olabilmektedir. Bu çalışmada üç adet farklı geometrik özelliklere sahip kanatlı borulu ısı değiştirici için, kuru çalışma koşullan altında, hava tarafı ısı taşınım katsayısı ve sürtünme faktörü deneysel olarak belirlenmiştir. Yapılan deneylerde hava hızı, 0.8 ile 4.4 m/s arasında seçilmiştir. Hava tarafı giriş koşullan, 27 °C ile 28 °C arasındadır. Isı değiştiricinin içinde dolaşan suyun sıcaklığı, 47 °C ± 0.1 °C olarak ayarlanmıştır. Su debisi, soğutucu akışkanın buharlaştıncı borulan içinde sabit sıcaklıkta akması durumunu simüle edebilmek amacıyla, su giriş / çıkış sıcaklıklan arasındaki fark 2 °C - 3 °C olacak şekilde ayarlanmıştır. Kanatlı borulu ısı değiştirgeçlerinin ısı taşınım katsayısı, ısı değiştirgeci geometrisine ve Reynolds sayısına bağlıdır. Reynolds sayısı arttıkça ısı taşınım katsayısı azalmaktadır. Yapılan deneylerde değişken olarak seçilen kanat aralığı azaldıkça, ısı taşınım katsayısı azalmaktadır. Kanatlı borulu ısı değiştirgeçlerinin sürtünme faktörü ise ısı taşınım katsayısında olduğu gibi ısı değiştirgeci geometrisine ve Reynolds sayısına bağılıdır. Yine aynı şekilde, Reynolds sayısı arttıkça sürtünme faktörü azalmaktadır. Isı değiştirgeçlerinin kanat aralığı arttığında ise sürtünme faktörü değeri azalmaktadır.

The purpose of this study is to build an experimental setup on which the performance tests of heat exchangers are performed and to have a database about the heat exchangers having different geometry. The experimental setup had been designed to find out heat transfer coefficients and friction factors of finned-tube heat exchangers for both cases of dry and wet surfaces. In this study, doing performance measurements of finned-tube heat exchangers that are used in air conditioners of ARÇELİK A.S. and defining a mathematical model by using the results of the performance measurements was aimed. Having such a model gives the opportunity to determine the performance of a heat exchanger whose geometrical properties are given. Under dry conditions, the air-side heat transfer coefficients and the friction factors of three finned-tube heat exchangers having different geometry were found out. In the heat exchangers used in ARÇELİK A.Ş. air-conditioners, the heat is transferred to gaseous fluids. Due to geometrical limitations and low heat transfer coefficients of gases, fins are used in the heat exchangers of air - conditioners. These types of heat exchangers are named "Finned - Tube Heat Exchangers". The heat transfer in finned - tube heat exchangers depends on the parameters such as, fin type, fin height, space between the fins, fin thickness, diameter of tubes, tube arrangement, space between fins in horizontal and vertical direction and flow type. The following assumptions are made in order to evaluate the data obtained from the experiments : - The air passes through the tubes and fins has uniform velocity and temperature. Condensation does not occur on the fins of evaporator. - The evaporation of the refrigerant is occurred in two - phases. - In the axial direction on the tubes, the heat transfer by conduction does not occurred. Hence, the heat obtained from air is transferred to the fluid without any loss. - The heat transfer properties of the fluids used in the mathematical model is calculated regarding to the average temperatures of fluids. XIV During the experiments, the air velocity were selected between 0.8 m/s and 4.4 m/s. The air-side inlet temperature was 27 °C - 28 °C and the water temperature in the heat exchanger was (47 ±0.1) °C. In order to simulate the flow of refrigerant at constant temperature through the tubes of the evaporator, water flow rate was arranged so that the temperature difference between water inlet and outlet was to be 2 °C - 3 °C. Total heat transfer due to the temperature difference between fluids can be expressed as Q = Ud x Ad x ATlmtÇ = UdxAdxFx ATlnii2 (1) The evaporator samples were selected as one - row. In this type heat exchangers, the refrigerant passes through the copper tubes intersects with the fluid, flows over the copper tubes and fins, more than once. The correction factor for cross - flow, F, can be obtained from the tables prepared for cross - flow types. The table prepared for the case of one row - cross - flow, in which the fluid is mixed, is shown at Fig. 2.7[2]. The correction factor for cross - flow, F, is obtained as 0.987, when the temperatures of the refrigerant fluid and the air, are used. Overall conductance, U, can be defined as Ud = Ad +t»xAd + hx x 4 kw x Apm hd x [Atd +Tjfx Af) (2) The logarithmic temperature difference, AT^, in equation (1) can be defined by A7L2 = (Tag-Thç)-{Tao-Thg) In T -T T -T 1aç Âhg (3) The total heat transfer through air - side had been calculated by the equation, Q = mhxcphx(Thg-Thç) (4). The total heat transfer through water - side had been calculated by the equation, Q = msuxcpsux(Tsg-Tsç) (5) The specific heat of water in equation (5), CpSU, was calculated regarding to the average water temperature. XV In ASHRAE 33-78 [26], the heat capacity of evaporator is calculated having the arithmetical average of heat transfer calculated regarding to the air - side and water - side. Vhava + "su (6) The efficiency of the plate-finned tubes can be calculated from the following equation. _ tanh(mreO) mr$ (7) The parameter m in this equation is given by, m [2xhd (8) and re, O r" = 0,535 xStxU- 0,3 (9) O 2r -1 Vdo A l + 0,351n 2r (10) Jo J Reynolds number used in the calculations done for airside is defined as Re = ^^- 01) In equation (11), dh is the hydraulic diameter and it is selected equal to the external diameter of the tube. While the air flows through the external surfaces of finned-tubes, a pressure drop occurs due to the friction and turbulence. This pressure drop [3] can be calculated by C2 An =-^-x ^ 2xp, Mi + />? y.. P~ mm rm (12) During the experiments, water was used as a refrigerant. The heat transfer coefficient for water-side can be calculated by Dittus-Boeltar equation. XVI Nu = -^-J- = 0.023 x ReT x Pr*3 (13) Dimensionless Colburn factor, j, which defines the heat transfer, is expressed as Nu Rex Pr' J = 1T-^I3 O4) The data obtained by the experiments performed under the above conditions are evaluated in a macro of Microsoft Excel for the case of the dry conditions. The most appropriate equations are obtained by using the finite element method. The data obtained from the experiments were evaluated by using the equations given above. Hence, the equations can be expressed as follows: ; = axReb (15) and for the friction factor, / = cxRe" (16) When the fin factor is used, j =axRebx8CI (17) and for the friction factor, f =axRebxsCI (18) Hence, the equations for heat transfer coefficient and friction factor dependent upon Re and Pr numbers was defined. The aim of the experimental study is to obtained a heat transfer coefficient and a friction factor which are valid for all of the samples those having different fin spacing and to use this only equation for different geometry. The 29 experimental data obtained from the samples numbered 1, 2 and 3 of finned - tube heat exchangers were evaluated and the heat transfer coefficient could be expressed by the formula j = 0,81 072898 xRe^'^^xe-0'042014397 (19) and for the friction factor, f = 6,1 997998 x Re"0 -38458891 x e'0 29730721 (20) xvu The equations for air - side heat transfer coefficient and friction factor are valid in the range 500
Mathematical model of finned tube heat exchangers for thermal simulation soft ware of air conditioners
Tez (Yüksek Lisans) -- İstanbul Teknik Üniversitesi, Sosyal Bilimler Enstitüsü, 1999
Keywords: Isı değiştiriciler; Matematiksel modelleme; İklimlendirme; Heat exchangers; Mathematical modelling; Air conditioning