Levha kanatlı boru demetlerinde ısı kütle ve momentum geçişi
Heat mass and momentum transfer in plate fin heat exchangers
- Tez No: 14300
- Danışmanlar: PROF.DR. AHMET RASİM BÜYÜKTÜR
- Tez Türü: Doktora
- Konular: Makine Mühendisliği, Mechanical Engineering
- Anahtar Kelimeler: Belirtilmemiş.
- Yıl: 1990
- Dil: Türkçe
- Üniversite: İstanbul Teknik Üniversitesi
- Enstitü: Fen Bilimleri Enstitüsü
- Ana Bilim Dalı: Belirtilmemiş.
- Bilim Dalı: Belirtilmemiş.
- Sayfa Sayısı: 199
Özet
Ö2ET Bu çalışmada levha kanatlı boru denetlerinde ısı» kütle ve momentum geçişi deneysel olarak incelenmiş ve üzerinde yoğusma olan yüzeylerdeki ısı geçişi için hesap sekli verilmiştir. Kuru ve ıslak kanat durumlarındaki ısı kütle ve momentum geçişinin deneysel olarak belirlenebilmesi için bir deney tesisatı kurulmuştur. Levha kanatlı boru demetleri havanın ısıtılmadı, soğutulması ve nem azaltılmasında kullanılmaktadır. Bu ısı değiştiricilerinin boyutlandırılabilmesi için ıslak ve kuru yüzey durumlarına göre ısı tasınım katsayıları ve sürtünme faktörleri bağıntılarının bilinmesi gerekir. Kuru yüzey halini içeren bağıntılar yeteri kadar olmasına rağmen ıslak yüzey durumundaki bağıntılar sınırlı sayıdadır. Deneyler havanın ısıtılması ve soğutulması durumları için yapılarak farklı Reynolds sayılarında ısı tasınım katsayıları, Colburn ve sürtünme faktörleri ve yüzey verimleri hesaplanmıştır. Bu çalışmada değişik geometriye ait 7 deney parçasında deneyler yapılarak ıslak ve kuru yüzey haline göre ısı taşınım için. J = A ReB sürtünme faktörü için. f - C ReD seklinde boyutsuz bağıntılar elde edilmiştir. -X
Özet (Çeviri)
SUMMARY HEAT MASS AND MOMENTUM TRANSFER IN PLATE PIN HEAT EXCHANGERS In this study heat» mass and momentum transfer in banks of tubes with plate fins on which condensation occur was studied both theoretically and experimentally. A computational procedure was given for the sizing of heat exchangers on which condensation occurs. Convective heat transfer coefficients and friction factors obtained experimentally were compared with the values predicted by the computational procedure. The experimental setup constructed for this study was designed for both heating and cooling of the moist air flowing over the finned tubes so that the heat transfer coefficients for the wet and dry operation could be measured. The setup consists of the wind tunnel, water heater» chiller, air heater, humidifier, water tank and flow straigtheners. In this study, heat transfer coefficient, friction factor and the fin efficiency were experimentally determined for seven different plate heat exhangers having different geometries for both wet and dry operating conditions. Experiments were done on air speeds between 0.9 and 4 m/s. For the case of the cooling, the entry conditions of moist air were between 27 and 30 C dry bulb temparature and between SO and 60 percent relative humidity. The cooling water temperature was kept between 6 and 7 C. For the case of heating, hot water temperature was kept between 47 and 48 C. When the temperature of the cold surface used for cooling of the moist air is below the dew point temperature, the water vapour present in the moist air will condense. Therefore, the sensible heat transfer is accompanied by the latent heat transfer. The condensed vapour forms a water film on the metal surface. The temperature, absolute humidity and velocity change in a boundary layer over the film. The saturated water vapour in contact with the film can be considered at the same temperature as the water film. The sensible heat transfer due to temperature difference can be expressed by the following equation : dQ = h t T - T > dA -Xi-The latent heat transfer can be axpressed by the following equation : dQ = h f W - W ) C H - H ) dA g d c b au Total heat transfer is the sum of the sensible and latent heat transfers. If the Lewis number is taken as 1.0» the total heat transfer can be expressed as : h dA dQ = - -* C H - ^ ) C In the banks of tubes with plate fins» the total heat transfer for the wet operation can practically be written in terms of a total heat transfer coefficient K, y Q^ ? K A. AH y t m where K is given as below : y The logarithmic mean entalpy difference AH is calculated in the following form : H - H ağ aç AH » n R ln[ R + In ( 1 - RP > R - R H - H hg he sc sg p _ p _ _ -. H - H * H ' - H aç ag hg &g XilIf it Is assusaed that the film condensation occurs and conduction in the water film is only in the direction perpendicular to the fins the efficiency of the plate finned tubes can be calculated from the following equation. tanh ) V m rw * The parameter m in this equation is given by h 2 ** M = k k The equivalent convective heat transfer coefficent for wet surfaces is defined by the following equation : h = * C y, pro f + b h k d y a The parameter tp is calculated from : r r - ( - ~ - D 1 1 + 0.33 InC- p- ) bd bd Reynolds number on the air side is defined as Re !L where d is the hydraulic diamater calculated from n 4 A. a, mm 1Air flowing outside the tubes loses pressure due to internal friction and interaction with the heat exchanger surfaces. The pressure loss of the air was measured by the slant alcohol manometer and the friction factor was calculated by : AP = The heating and cooling fluid used in the experiments was water. The convective heat transfer coefficient for the flow inside of the tubes was calculated from the Dittus-Boelter equation : Nu = 0.023 ReOBPrn The nondimensional paremeter expressing the heat convection coefficient was the Colburn factor. The Col burn factor is defined as : h u C 2/3 J = y, * P"> v Q C k pm h The following equation was used to relate the convective heat transfer coefficients for a wet surface to that for a dry surface : h K - 1. 067 h, V°-m yh k The experimental data was analyzed and reduced to the equations for heat transfer coefficient in the form : J ? A Re and for the friction factor. f = C Re° by using the least square fitting criteria. -Xiv-Experimental results are shown graphically In section 6. Results of this study can be summarized as follows -. Convective heat transfer coefficients when condensation occurs are greater than convective heat transfer coefficients for the dry operations. The difference between the wet and dry surface convection heat transfer coefficients depends on the heat exhanger geometry and the Reynolds number of the flow. Results of the experiments show that when condensation occurs the convection heat transfer coefficients become IS H greater than the dry case for the fluted plate fins and S3 % greater for smooth plate fins. The difference between the wet and dry convection heat transfer coefficients increases as the Reynolds number increases. Flutes on fins increases the convective heat transfer coefficients for both the wet and the dry cases. The reason for this increase is the additional turbulence and the separation of the boundary layer. The convective heat transfer coefficient increases as the fin spacing increases. This result is observed for both the wet and the dry cases. The convective heat transfer coefficient increases as the number of rows of tubes decreases. This is due to entrance effect. The convective heat transfer coefficient is higher for the triangular tube arrangement comparig to the square tube arrangement This is due to the higher local turbulence that can be achived under triangular arrangement. Similar results hold for the frlcion factor. When condensation occurs the friction factor is 70 % with respect to the dry case. The droplets which form between fins increase the resistance to air flow. Friction factors for the tube banks with plate fins increase as the fin spacing increases. This effect holds for the both dry and wet surfaces. However, the difference between the friction factors for the wet and dry surfaces decrease as the fin spacing increases. Since the increase in fin spacing increases the turbulance, the increase in the friction factor for the dry surface is greater than that for the wet surface. Flutes on fins increase the friction factor for the both dry and wet surfaces. This is due to the resistance to the flow caused by the flutes. -XV-
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