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Levhalı ısı değiştiricilerinde yaygın olarak kullanılan periyodik dalgalı yüzeyli kanal için taşınımla ısı geçişi ve sürtünme faktörü sayısal incelenmiştir. Sayısal çalışmada Sonlu Hacimler Yöntemi (Finite Volume Method – FVM) kullanılmıştır. Realize edilebilir k-ε (Realizable k-ε), k-ω, Kayma Gerilmesi Taşınımı (Shear Stress Transport – SST) ve geçiş SST (transition SST) olarak adlandırılan, dört farklı Reynolds Ortalamalı Navier-Stokes (Reynolds Averaged Navier-Stokes Simulations – RANS) tabanlı türbülans modelleri kullanılarak elde edilen sonuçlar karşılaştırılmıştır. Önceki deneysel çalışmalardan elde edilen sonuçlar da sayısal sonuçları doğrulamak için kullanılmıştır. Çalışmalar, dalga eğim açısının 30° olduğu dalgalı yüzeyli kanallarda hava akışı için yapılmıştır. Prandtl sayısı 0.70 sabit kalırken, Reynolds sayısı 2000 ile 11000 arasında değiştirilmiştir. Nusselt sayısı, Colburn faktörü, sürtünme faktörü ve ısı geçişi iyileştirme performansı Reynolds sayısına bağlı olarak verilmiştir. Dalgalı yüzeyli geometrinin ve kanal yüksekliğinin etkileri ele alınmıştır. Deneysel ve sayısal sonuçlar arasında en iyi uyum SST modeli ile sağlanırken, en çok hata ise realize edilebilir k-ε modelden elde edilmiştir.

Ortam ısı kazancı, verilen bir zamanda ortam içinde üretilen veya o ortama giren ısı miktarıdır. Ortam soğutma yükü ise bir ortamda sabit bir hava sıcaklığı sağlamak için ortamdan çekilmesi gerekli olan ısı miktarıdır. Öyle ki herhangi bir zaman için toplam ısı kazancı mutlaka aynı zamandaki soğutma yüküne eşit olmak zorunda değildir.Bir ortamdaki cisimlerin ve yüzeylerin ısı depolama kabiliyeti ortamın anlık ısı kazancı ve aynı andaki soğutma yükü üzerinde etkilidir. Bir ortamda ısının ne şekilde depolandığı ve depolanan ısı miktarının tespiti, soğutma yükü hesaplarında gerçekçi değerlere varmak açısından önemlidir.Soğutma yükünün hesabının doğru yapılması ve bir sistem için uygun olması sistemin performansı açısından etkilidir. Soğutma yükü hesaplarına etki eden değişkenlerin sayısı çok fazladır ve genellikle bu değişkenlerin kesin olarak tanımlanması zor olup, bunların daima birbiri ile karmaşık bir iliksisi vardır. Soğutma yükünü meydana getiren pek çok bileşenin değeri, 24 saatlik zaman dilimi içinde önemli seviyede değişir.Soğutma yükü hesaplarında yapı bileşenlerinden, pencerelerden, sızma ve havalandırmadan, aydınlatmadan, insan ve cihazlardan kaynaklanan ısı kazançları taşınım ve ışınım oranlarına bağlı olarak duyulur soğutma yüklerine dönüştürülür ve bunlara gizli soğutma yükleri de eklenerek toplam soğutma yükü belirlenir. Tüm yapı tek bir zon olarak ele alınabilir fakat cihaz seçimi ve tasarımı oda oda yapılan hesaplara dayanmalıdır.Bu tez çalışması kapsamında Ashrae Temel El Kitabı'nda konut dışı yapılar için önerilmiş olan soğutma yükü hesaplama metotlarından CLTD/CLF ve RTS metotları kapsamlı bir şekilde incelenmiş, Visual Basic programlama dilinin kullanıldığı soğutma yükü hesaplama programı oluşturularak, incelenen bir örnek oda için bu programdan elde edilen sonuçlar, elle hesaplanan sonuçlar ve aynı örnek oda için Carrier programından elde edilen sonuçlarla karşılaştırılmıştır. Space heat gain is the amount of heat that is generated in the space or given to the space at the specified time. However, space cooling load is the required amount of heat that should be dissipated from the space in order to keep the space temperature constant. Thus, total heat gain at an arbitrary time should not be equal to the cooling load at that time.Heat storage capacity of the body and surfaces in the space affects the instant heat gain and the cooling load. In order to obtain reasonable values in the cooling load calculations, heat storage characteristics of the space and the amount of stored heat should be taken into account.The correct calculation of the cooling load and the availability of the cooling load for the system is important for the system performance. There are significant number of variables that affect the cooling load calculation and generally those variables can not be exactly defined, besides they have complicated interactions between each other. The values of the components that constitute the cooling load change considerably in a period of 24 hours.In cooling load calculations, heat gains that are supplied by the building materials, windows, infiltration, ventilation, lighting, occupants and devices are converted to the sensible cooling loads depending on portions of radiant and convective. Thus, total cooling load is obtained by adding latent cooling load to the sensible cooling load. Entire building can be considered as a single zone but the selection and design of the devices must be referred to the calculations that are done for each room.In this study, CLTD/CLF and RTS methods, the methods of cooling load calculation, that are recommended for nonresidential buildings in Ashrae Fundamental Handbook were examined comprehensively. For this purpose, a sample room was selected and cooling load calculation program was constructed by utilizing Visual Basic programming language. The results obtained from the code were compared with the results both calculated manually and the commercial program used by Carrier. 139

Tez (Yüksek Lisans) -- İstanbul Teknik Üniversitesi, Fen Bilimleri Enstitüsü, 2010, Thesis (M.Sc.) -- İstanbul Technical University, Institute of Science and Technology, 2010, Bu tez çalışması kapsamında Ashrae Temel El Kitabı’nda konut dışı yapılar için önerilmiş olan soğutma yükü hesaplama metotlarından CLTD/CLF ve RTS metotları kapsamlı bir şekilde incelenmiş, Visual Basic programlama dilinin kullanıldığı soğutma yükü hesaplama programı oluşturularak, incelenen bir örnek oda için bu programdan elde edilen sonuçlar, elle hesaplanan sonuçlar ve aynı örnek oda için Carrier programından elde edilen sonuçlarla karşılaştırılmıştır. Örnek oda için yılın tüm zamanlarında CLTD/CLF VE RTS metotları için elle, SYHP_v2.0 programı, RTS_CL programı ve Carrier programı ile yapılan hesaplamalarda maksimum soğutma yükü Ağustos ayında çıkmıştır. En düşük toplam soğutma yükü SYHP_v2.0 programı ve CLTD/CLF metodu için el ile yapılan hesaplamalara göre elde edilmiştir. En yüksek toplam soğutma yükü RTS metoduna göre el ile yapılan hesaplamaya göre elde edilmiştir., In this study, CLTD/CLF and RTS methods, the methods of cooling load calculation, that are recommended for nonresidential buildings in Ashrae Fundamental Handbook were examined comprehensively. For this purpose, a sample room was selected and cooling load calculation program was constructed by utilizing Visual Basic programming language. The results obtained from the code were compared with the results both calculated manually and the commercial program used by Carrier. For sample room, while maximum cooling load was obtained manually for CLTD/CLF and RTS methods in all time of year. It was determined in calculations that is made by SYHP_v2.0, RTS_CL and Carrier programs in August. Minimum total cooling load was gotten from cooling load that has been calculated manually by SYHP_v2.0 program and CLTD/CLF method. Maximum total cooling load was acquired from calculations that have been made manually by RTS method., Yüksek Lisans, M.Sc.

Tez (Yüksek Lisans) -- İstanbul Teknik Üniversitesi, Fen Bilimleri Enstitüsü, 2010, Thesis (M.Sc.) -- İstanbul Technical University, Institute of Science and Technology, 2010, Bu çalışmada, klima santralleri veya havalandırma kanalları içerisinde ses yutumu amacıyla kullanılan, susturucular içerisindeki akış modellenmiştir. Susturucular içerisinde ses yutumunu sağlayan eleman olarak gözenekli bir malzeme olan cam yünü kullanılmaktadır. Hücre içerisinde susturucular arasından geçen hava gözenekli ortam ile temasta olmaktadır. Susturucu alın yüzeylerinin uygunsuz tasarımı, susturucu içerisinde gereksiz basınç kayıpları ve akıştan kaynaklı olarak ses oluşumuna neden olmaktadır. Tez çalışmasında susturucu alın yüzeyleri düz sac kaplı, üçgen sac kaplı ve silindirik sac kaplı olmak üzere üç farklı şekilde tasarlanmıştır ve fluent programı ile akış analizi yapılarak hücre içerisinde meydana gelen basınç düşümleri elde edilmiştir. Elde edilen sonuçlar karşılaştırılmıştır ve susturucu alın yüzeylerinin silindirik sac ile kaplanması hücrede meydana gelen basınç kayıplarını yüzde yirmi altı mertebesinde azalttığı sonucu elde edilmiştir., In this study, the flow is modeled within silencers which is used for air handling units or air ducts that perform voice absorbing. The glass wool is used for sound absorbing within the silencers which is a porous material. The air which is passing from between the silencers in sections is contact with porous media. Silencer surfaces that are perpendicular to the airflow causes unnecessary loss of pressure because of improper design and are caused by the formation of the sound source. Silencer surfaces that are perpendicular to airflow covered with three different shapes in thesis, rectangle, triangular and cylindrical steel sheet are designed and modeled with Fluent and obtained pressure drops that occur within the silencer section. Results compared and surfaces are covered with cylindrical sheet in silencer section that occurs in order to reduce pressure losses as a result twenty six percent were obtained., Yüksek Lisans, M.Sc.

Bu çalışmada dünyada gemilerde yapılan ısı yalıtımı uygulamaları üzerine yapılmış araştırmalar, çalışmalar ve kullanılan yönetmelikler incelenmiş, önemli yönleri derlenmiş, uygulama detayları ve ülkemiz şartlarına uyarlanmış örnekler ve hesap değerleri ile bir kılavuz oluşturulmaya çalışılmıştır. Gemilerde ısı yalıtımı uygulamaları; gemilerin yapı bileşenlerinin yalıtılması, mekanik tesisatın yalıtılması ve soğuk depolar ile konteynerlerin yalıtılması alt başlıklarında incelenmiştir. Gemilerin yapı bileşenlerinden ısı kaybında önemli bir etkisi olan mukavemet elemanlarının yalıtımı da gemilerin yapı bileşenlerinin yalıtılması konusu altında ayrı bir başlıkta işlenmiş, bu elemanlardan olan ısı geçişinin hesap yöntemi örneklerle açıklanmıştır. Bunun yanında gemilerde ısı yalıtımında kullanılan malzemelerde istenilen özellikler, uygulamada kullanılan malzemeler ve özellikleri işlenmiş ve yönetmeliklerde yer alan malzeme sınırlamaları verilmiştir. Bir diğer bölümde tesisat ve yapı elemanlarında oluşacak terleme ve yoğuşmanın yalıtım uygulamaları ile önlenmesi, buhar kesicilerin kullanılması, özellikleri ve sınıflandırılması üzerinde durulmuştur. Son bölümde konunun daha iyi anlaşılması amacıyla çeşitli örneklere yer verilmiştir. In this study, researches, studies and regulations worldwide about insulation of ships, applications have been analyzed, important parts have been gathered and selected, a guide was compiled, by using the application details, examples and calculation values which are adapted to our country?s conditions. The subject of insulation of ships have been analyzed under three main sub titles such as; insulation of ships structural compounds, insulation of the mechanical installation, insulation of cold rooms and containers. The insulation of stiffeners which have a significant share in the heat loss has been analyzed as a subject under the subtitle of ?insulation of ships? structural compounds? and the calculation method for heat transfer caused by these elements has been explained with examples following. Furthermore; the properties of the materials which are used for insulating ships, the materials which are used in applications and their properties have been analyzed and the legal regulative limitations for these materials have been presented. Preventing the sweating and condensation problems within installation and structural compounds with the help of insulation applications; usage of vapor barriers, their properties and classification have been emphasized in another chapter. There are examples in the last part for the purpose of topic to be understood better. 160

Tez (Doktora) -- İstanbul Teknik Üniversitesi, Fen Bilimleri Enstitüsü, 2006, Thesis (PhD) -- İstanbul Technical University, Institute of Science and Technology, 2006, Bu çalışmada, gaz türbin kanadı üzerinde film soğutmanın sayısal ve deneysel incelenmesi yapılmıştır. Bu kapsamda deneysel olarak, düz bir levha üzerine akış yönü ile 30 o açı yapacak şekilde tek sıralı silindirik, tek ve iki sıralı dikdörtgen delikler ele alınmıştır. Deliklerden farklı hız ve sıcaklıklarda hava enjeksiyonu ile levha üzerinde oluşan sıcaklıklar ölçülmüş ve film soğutma etkenlikleri hesaplanmıştır. Film soğutma etkenlikleri her delik geometrisi ve sıra için ayrı ayrı çıkartılmıştır. Etkenlik karşılaştırmaları farklı üfleme ve momentum akı oranları ve farklı üfleme sıcaklıkları için yapılmıştır. Ayrıca tek sıralı silindirik ve iki sıralı dikdörtgen delik geometrileri film soğutma etkenlik alanları sıvı kristal kullanılarak bulunmuştur. Deneysel olarak yapılan diğer bir çalışma ise herbir delik geometrisi ve 1.75 üfleme oranı için iki boyutlu hız ve türbülans şiddetleri ölçümüdür. Hız ve türbülans ölçümleri için jet ve ana akışın karışım bölgesinde sıcak tel(hot-wire) anemometresi kullanılarak yapılmıştır. Deneysel çalışmalar sayısal olarak modellenmiş, üfleme oranları dikkate alınarak film soğutma etkenlikleri levha yüzeyinde iki boyutlu olarak karşılaştırılmıştır. Ayrıca farklı enjeksiyon delik geometrilerinin ve yüzey eğriliğinin film soğutmasına etkileri sayısal olarak incelenmiştir. Sayısal ve deneysel çalışmaların sonucunda delik geometrisi, üfleme oranı, yüzey eğriliği, momentum akı oranı film soğutma etkenliği üzerindeki etkiler elde edilmiştir., In this study, experimental and computational investigation of film cooling on the gas turbine blade have been done. Cylindrical and rectangular holes with a single row and rectangular holes with two rows have been examined on a flate plate. The holes slope are 30o by flow direction on the flate plate. The air with different velocities and temperatures has been enjected from the holes and the temperatures on the plate have been measured and the film cooling effectiveness have been calculated. Film cooling effectiveness for every hole geometry and row has been evaluated. Comparings of effectiveness for different blowing and momentum flux ratios and blowing temperatures have been done. In addition, the region of film cooling effectiveness for cylindrical hole with a single row and rectangular hole with two rows have been measured by using liquid cyristal. The other experimental study is two dimensional velocity and turbulence intensity measurements for every hole geometry and 1.75 blowing ratio. The measurements of velocity and turbulence intensity have been done by using hot wire anemometer at the mixture region of jet and main flow. The experimental studies have been modeled as computational and two dimensional film cooling effectiveness on the flate plate have been investigated for blowing ratios. Moreover, the effects of the different jet hole geometries and the surface curvature have been investigation as computational. As a results of experimental and computational studies, hole geometry, surface curvature, blowing and momentum flux ratio effects on the film cooling effectiveness have been found., Doktora, PhD

Tez (Yüksek Lisans) -- İstanbul Teknik Üniversitesi, Enerji Enstitüsü, 2006, Thesis (M.Sc.) -- İstanbul Technical University, Energy Institute, 2006, Bu çalışmada konut olarak kullanılan binalar için enerji kayıp ve kazançları geliştirilen iki ayrı bilgisayar programı ile hesaplanmıştır. Enerji kayıplarının hesabında Türk Standartları Enstitüsünün TS825 ?Binalarda Isı Yalıtım Kuralları? standardında önerilen metot kullanılmıştır. Hesaplama sürelerini kısaltmak ve hesap doğruluğunu sağlamak için Microsoft Visual C# programlama dili kullanılarak bir program geliştirilmiş, ele alınan örnek bir proje ile bilgisayar programının sonuçları kontrol edilmiştir. Enerji kazançlarının hesabında ASHRAE'nin önerdiği ?Konutlarda Soğutma ve Isıtma Yüklerinin Hesabı? baz alınarak Microsoft Visual C# programlama dili ile enerji kazancı hesaplayan bir program geliştirilmiş, ele alınan örnek proje için değişik yalıtım senaryoları kullanılarak bilgisayar programının sonuçları kontrol edilmiştir. Bilgisayar kullanılarak gerçekleştirilen hesaplar, hesap süresini kısaltmakta ve kullanıcının hata yapmasını önlemektedir. Çalışmada geliştirilen bilgisayar programları ile binalarda enerji kayıp ve kazanç hesapları pratik ve hatasız olarak gerçekleştirilebilmektedir. Anahtar Kelimeler: Enerji kaybı, enerji kazancı, yalıtım Bilim Dalı Sayısal Kodu, In this study, the energy gains and losses will be calculated by two separate programs developed. For the energy gain calculations, TS825 ?Thermal Insulation in Buildings? of the Turkish Standards Institution is utilized. A computer program is developed in order to shorten the time consumed during calculations and reduce the miscalculations. It is written using Microsoft Visual C# 2005 programming language. The example considered in this study is initially prepared manually and afterwards re-prepared utilizing computer. Following this, the end results are compared. For the cooling load calculations of buildings, ASHRAE?s Fundamentals suggested method ?Residential Load Calculations? is utilized. Again, a computer program is written with Microsoft Visual C# 2005 programming language that uses this method. An example project is considered with different scenarios and load calculations are made both manually and by the computer program, later the results are compared. The calculations that are made using a computer provide users with speed and also reduce chances of miscalculation. With the programs that are developed in this study, residential energy calculations can be carried out more practically/more efficiently. Keywords: Energy loss, energy gain, insulation Science Code, Yüksek Lisans, M.Sc.

Tez (Yüksek Lisans) -- İstanbul Teknik Üniversitesi, Fen Bilimleri Enstitüsü, 2003, Thesis (M.Sc.) -- İstanbul Technical University, Institute of Science and Technology, 2003, Bu çalışma sırasında araçlarda kullanılan kompakt ısı değiştiricisi üzerinde ısıl performans testi yapılmıştır. Bu çalışmanın amacı, deneysel olarak elde edilen giriş ve çıkış sıcaklık farkı ile teorik olarak elde edilen giriş ve çıkış sıcaklık farkının karşılaştırılması ve radyatör performans karakteristik eğrisinin elde edilmesidir. Hava tünelinde dokuz ayrı test yapılmıştır. Testler sırasında kontrol edilebilen değerler sırasıyla; hava debisi, hava giriş sıcaklığı, soğutucu akışkan debisi ve soğutucu akışkan giriş sıcaklığıdır. Her bir test sırasında hava ve soğutucu akışkan çıkış sıcaklıkları ölçülmüştür. Üç adet testin çıkış sıcaklık değerleri kullanılmak suretiyle hava tarafının ısı taşınım katsayısı Reynolds sayının bir fonksiyonu olarak bulunmuştur. Bir sonraki adım olarak ta, deneysel olarak elde edilen ısı taşınım katsayısı ve diğer kalan test verileri kullanılarak hava ve soğutucu akışkan çıkış sıcaklıkları NTU (Geçiş Birimi Sayısı Yaklaşımı) prosedürü yardımıyla hesaplanmıştır. Deneysel ve teorik olarak elde edilen giriş ve çıkış arasındaki sıcaklık farklarının karşılaştırılması göstermektedir ki, her durum için de elde edilen sıcaklık farkları neredeyse birbiriyle aynıdır. Bunlara ek olarak, radyatör performans karakteristik eğrisi hava ve soğutucu akışkanın bir fonksiyonu olarak elde edilmiştir., In this study, heat performance test is implemented for the compact heat exchanger used on vehicles. The aim of the study is to compare the experimental difference between inlet and outlet temperatures to the theoretical one and to obtain radiator performance characteristic curves. Nine different experiments are performed at wind tunnel. During the test period, controlled items are the following; air mass flow rate, air inlet temperature, coolant mass flow rate and coolant inlet temperature. During each test period, air and coolant outlet temperatures are measured. Using the measured outlet temperatures of three of the experiments, air side convection coefficient is obtained as a function of Reynolds number. In the next step, air and coolant outlet temparatures are obtained theoretically by NTU (Number of Transfer Units) procedure, using experimental convection coefficient and controlled items of the remaining experiments. The comparison of experimental and theoretical temperature difference between inlet and outlet shows that the temperature difference is almost the same in both cases. Additionally, radiator performance characteristic curves are obtained as a function of air and coolant mass flow rates., Yüksek Lisans, M.Sc.

Isı değiştiricileri altı grupta toplanır; Isı transferi sürecine göre, yüzey yoğunluğuna göre, geometrisine (konstrüksiyona) göre, akış şekline göre, akışkan sayısına göre ve ısı transferi mekanizmasına göredir. Isı transferi sürecine göre doğrudan temaslı veya dolaylı temaslı olarak ikiye ayrılır. Yüzey yoğunluğuna göre ısı değiştiricileri kompakt ve kompakt olmayan şeklinde iki grupta toplanır. Isı transferi yüzey alanının ısı değiştiricisi hacmine oram o ısı değiştiricisinin kompaktlığını gösterir. Eğer bir ısı değiştiricisinde bu oran 700 m2/m3'den fazla ise bu tip ısı değiştiricilerine kompakt ısı değiştiricileri denir. Bunun tam tersi olarak, oran 700 m2/m 'den az ise bu ısı değiştiricisi kompakt olmayan bir ısı değiştiricidir. Geometrisine göre ısı değiştiricileri dört grupta toplanır: Borum ısı değiştiriciler, levhalı ısı değiştiriciler, kanatlı yüzeyli ısı değiştiriciler ve rejeneratif ısı değiştiricileridir. Isı değiştiricisinin içerisinde paralel, ters ve çapraz akış şekilleri vardır. Bunlardan ters akışlı olanda en fazla verim elde edilir. Isı transferi mekanizmasına göre ısı değiştiricileri dört grupta toplanır. Tek faz -Tek faz taşınım, İki faz - Tek faz taşınım, İki faz - İki faz taşınım ve Taşınım + ışınım' dır. Isı değiştiricisi içerisinden geçen akışkan sayısı çoğunlukla ikidir. Ancak bazı özel durumlarda üç veya daha fazla olabilir. Soğutma sistemi, motoru en verimli bir şekilde çalışabileceği sıcaklık bölgesinde tutabilecek şekilde dizayn edilmektedir. Bunu gerçekleştirebilmesi içinde motordan yüksek bir miktarda ısıyı dışarı atmalıdır. Soğutma sistemi bu ısıyı üç şekilde dışarı atar: İletim, taşınım ve radyasyon' dur. Soğutma sistemi bu üç yolu kullanarak ısının dışarı atılmasını maksimum düzeye getirebilecek şekilde dizayn edilmektedir. Çeşitli soğutma sistemleri otomobiller üzerinde kullanılmaktadır. Hava soğutmalı motorlarda havanın kanatlar üzerinden geçmesi suretiyle motor bloğu ve silindir kafaları soğutulmaktadır. Isı motor çevresinden geçmekte olan havaya atılmaktadır. Su soğutmalı sistemlerde ise motor çevresindeki kanalların içerisinden su ve antifreze karışımı geçirilerek motor soğutulmaktadır. Soğutucu akışkan bir pompa vasıtasıyla motor bloğuna gönderilir. Buradan su kanalları yardımıyla akışkan silindir bloğu çevresinden geçer oradan da radyatöre ulaşır. Radyatöre ulaşan soğutucu akışkan radyatör içerisinden geçen hava yardımıyla da soğutulur. Bir soğutma sisteminin dizaynını ve çalışma şeklini anlayabilmek için soğutucu akışkanın özelliklerini belirlemek çok önemlidir. Soğutucu akışkanın donmasını önlemek ve kaynama sıcaklığım arttırmak için içerisine antifreze katılır. Ayrıca antifreze katılmak suretiyle soğutma sisteminde oluşabilecek olan korozyonda önlenmiş olur. Soğutma sisteminin verimli bir şekilde çalışabilmesi için, sistemi oluşturan su pompası, su ceketleri, termostat, radyatör, basınç kapağı, genleşme tankı, fan, xııdavlumbaz, kayış ve sıcaklık göstergeleri gibi parçaların her an çalışır durumda ve bakımlı olması gerekmektedir. Radyatör, soğutucu akışkanın ısısını alan soğutma sisteminin en önemli elemanıdır. Soğutucu akışkan radyatör içerisindeki ince borulardan geçer. Bu arada havada radyatör dışından geçerek soğutucu akışkanın sıcaklığı düşürülmüş olur. ' Günümüzde aşağı veya çapraz akışlı radyatörler kullanılmaktadır. Soğutma sistemininde ortalama 1,5 barlık basıncın korunabilmesi için radyatör üzerinde basınç kapağı bulunmaktadır. Basınç, soğutucu akışkanın kaynama noktasını arttırır, kavitasyonu azaltır, radyatör borularını korur. Isı değiştiricilerinin bir çok uygulama alanına bulunmaktadır. Dolayısıyla ısı değiştiricisi araştırma ve geliştirmelerinin uzun bir tarihçesi vardır. Araştırma ve geliştirmeler halen günümüzde büyük bir hızla sürdürülmektedir. Bu çalışmalar ısı transferini daha iyi hale getirebilmek için yapılmakta ve bunu sağlamak içinde yeni ve özel ısı transferi yüzey alanları dizayn edilmektedir. Bu çalışmada ısı değiştirici hesaplamalarında kullanılabilecek bazı yaklaşım methodları verilmiştir. Dizayn aşamasında kullanılabilecek olan hesap yöntemleri ile ilgili olarak, bütün ısı değiştiricileri için tek bir prosedür ortaya koymak mümkün değildir. Çünkü parametreler her bir durum için farklılık göstermektedir. Isı değiştirici hesaplamalarında kullanılan en genel yöntemler LMTD (Logarithmic Mean Temperature Difference) ve NTU (The Number of Transfer Units) 'dır. Hesaplamalar sırasında her iki yöntemde kullanılabilir. Ancak NTU yöntemin diğerine göre bazı avantajları vardır. Örneğin ty, th0, rhc ve riıh değerleri biliniyor ve A yüzey alanı hesaplanması gerekiyor. U toplam ısı transferi katsayısının hesaplanabilmesi için her iki yöntem -kullanılarak ısı transferi katsayısı hesaplanmalıdır. NTU Yöntemi kullanılarak; l-Verilen değerler yardımıyla 8 verim ve Cmin/Cmax. değerleri hesaplanır. 2-Akış şekline göre 8 -NTU eğrileri kullanılarak NTU değeri hesaplanır. 3-A=NTU(Cmin/U) formülü yardımıyla da alan hesaplanır. LMTD Yöntemi kullanılarak; l-Verilen sıcaklık değerleri yardımıyla P ve R değerleri hesaplanır. 2- Uygun eğriler kullanılarak F düzeltme faktörü saptanır. 3-LMTD değeri hesaplanır. 4-q=Cc(tco-tcpCh(thi-tho) eşitliği kullanılarak, A=q/U(F)(LMTD)'dan alan hesaplanır. Ayrıca A, U, rhc, riıh, ty, ta değerleri bilinseydi ve çıkış sıcaklıkları olan th0 ve tco'nun hesaplanması gerekseydi, NTU Yöntemi kullanılarak; l-Verilen değerler yardımıyla NTU=UA/Cmin hesaplanır. 2-NTU, Cmin/Cmax. değerleri kullanılarak akış şekline göre uygun eğriler kullanılmak suretiyle 8 değeri hesaplanır.4-qK^Wtci)^h(thi-tho) formülü yardımıyla da ikinci akışkanın çıkış sıcaklığı hesaplanır. LMTD yöntemi kullanılarak; l -R=Cc/Ch'dan R değeri hesaplanır. 2-Akışkanlardan birinin çıkış sıcaklığı için bir değer verilir ve bunun yardımıyla da bir P değeri belirlenir (Birinci iterasyon) 3-Uygun eğriler yardımıyla F değeri bulunur. (Birinci iterasyon) 4-LMTD değeri belirlenir. 5-q=UAF(LMTD) formülü yardımıyla q değeri saptanır. 6-İkinci akışkanın çıkış sıcaklığı hesaplanır ve ikinci adımda birinci akışkan için verilen değer ile karşılaştırılır. 7-Uygun bir değer elde edilene kadar iki ve altı numaralar arasındaki adımlar tekrarlanır. Açıkça anlaşılabileceği üzere, NTU yöntemi ile daha hızlı bir şekilde sonuca ulaşılmaktadır. Bu sebepten dolayıdır ki; tezin hazırlanması sırasında NTU yöntemi kullanılmıştır. Yukarıda verilen bilgilere ek olarak, bu çalışmada bir araç radyatörü için ısıl performans testi yapılmıştır. Sonuçlar NTU yöntemi kullanılarak irdelenmiştir. Sonuç olarakta, ısı transferi hesaplama yöntemleri kullanılarak, mevcut radyatör için, (Rehava-j) Reynolds sayısı-radyatör hava tarafındaki ısı taşınım katsayısı j faktörü ve radyatör performans karakteristik eğrileri çıkartılmıştır. The heat exchangers are divided into six grups; Heat transfer process, surface compactness, geometry (construction), flow type, number of fluids and heat transfer mechanism. According to the heat transfer process classifıcation, the direct contact and the indirect contact type of heat exchangers are available. Classification to the surface compactness has two types; The ratio of heat transfer surface area to the volume of the system in heat exchangers is the measure of compactness. If the ratio is more than 700 m2/m3, this heat exchanger is called as a compact heat exchanger. in contrary, the heat exchanger is not a compact öne if this value is less than 700 m2/m3. Classification to the geometry has four types: Tube type heat exchangers, plate type heat exchangers, extended surface type heat exchangers and regenerative type heat exchangers. Paralel flow, counter flow and cross flow are the types of flow in heat exchangers. Counter flow gives the highest effectiveness. in the classifîcation according to the heat transfer mechanism, there are four items: öne phase flow convection in both sides, two phase flow convection in öne side and öne phase flow convection in the other side, two phase flow convection in the both'Sides and finally convection and radiation together. The number of fluids are mostly two in heat exchangers but in special procosses three ör more number of fluids may be used. The cooling system is designed to maintain the most efficient operating temperature within the engine. To do this, the cooling system must remove large quantities of heat from the engine. The cooling system uses three means of heat transfer to accomplish this: conduction, convection and radiation. it is designed to maximize the principles of heat transferring heat by these three means. Several types of cooling systems are now being used on the automobile. Air-cooled engines use air passing över fins to cool the engine block and cylinder heads. Heat is dissipated to the air passing around the engine. Liquid-cooled engines circulate a mixture of antifreeze and water through the engine to remove heat. The coolant is pumped with a water pump into the engine block, through internal passages, into the cylinder head, and into a radiator. Air flows through the radiator to remove the excess the heat captured by the coolant, which is either recirculated without being cooled ör sent to the radiator to be cooled. To understand the design and operation of the cooling system, it is important to identify coolant characteristics. Antifreeze is added to the coolant to protect it against freezing. Antifreeze also raises the coolant's boiling point, and it reduces corrosion within the cooling system. The cooling system has many parts that enable it to operate effectively. These include the water pump, water jackets, thermostat, radiator, pressure çap, expansion tanks, fan, shrouds, belts and temperature indicators.The main part of the cooling system is the radiator used to remove heat from the coolant. The coolant passes through small tubes inside the radiator. Air flow across the tubes and removes the heat. Both down-flow and cross-flow radiators are used in vehicles today. A pressure cap is placed on the radiator to maintain approximately 1,5 bars of pressure on the cooling system. The pressure increases the boiling point of coolant, reduces cavitation, protects radiator hoses and reduces surging. Because there are many important. applications, heat exchanger research and development has a long history. Such activity is by no means complete, however, as many talented workers continue to seek ways of improving design and performance. In fact, with heightened concern for energy conservation, there has been a steady and substantial increase in activity. A focal point for this work has been heat transfer enhancement, which includes the search for special heat exchanger surfaces through which enhancement may be achieved. In this study we have attempted to develop tools that will allow you to perform approximate heat exchanger calculations. More detailed considerations of the subject are available in the literature, including treatment of the uncertainties associated with heat exchanger analysis. Regarding design procedure for sensible heat transfer, it is diffucult to devise one procedure for designing all heat exchangers because the given parameters vary from situation to situation. All of the terminal temperatures may be known or only the inlet temperatures may be given. The mass flow rates may be fixed in some cases and variable in others. Usually the surface area is not given. LMTD (Logarithmic Mean Temperature Difference) and NTU (Number of Transfer Units) are the general heat exchanger design procedures. Either method may be used, but the NTU method. ha certain advantages. Consider the design problem where thi, th0, mc and mh are known and the area A is to be determined. With either approach the heat transfer coefficients must be determined so that the overall coefficient U can be computed. The NTU approach then proceeds as follows; 1 -Compute the effectiveness £ and Cmin/Cmax. from the given data. 2-Determine the NTU for the particular flow arrangment from the S-NTU curve. 3-Compute A from A=NTU(Cmir/U) The LMTD approach is as follows; 1 -Compute P and R from the given terminal temperatures. 2-Determine the correction factor F from the appropriate curve 3-CalcuJate the LMTD for an equivalent counterflow exchanger 4-Calculate A from A=q/U(F)(LMTD) where q=Cc(tCo-tci)=Ch(thi-tho) The NTU approach requires somewhat less effort in this case. Consider the design problem where A, U, rhc, mh, thi and t^ are given, and it is necessary to find the outlet temperatures th0, tc0. The NTU approach is as follows; 1 -Calculate the NTU=UA/Cmin from given data 2-Find £ from teh appropriate curve for the flow arrangment using NTU and ^min'^max- 3 -Compute one outlet temperature 4-Compute the other outlet temperature from, q=Cc(tco-tCi)=Ch(thi-tho) The LMTD approach requires the following iteration; xvi2- Assume one outlet temperature in order to compute P (first approximation) 3 -Find F from the appropriate curve (first approximation) 4-Evaluate LMTD (first approximation) 5-Determine q=UAF(LMTD) (first approximation) 6-Calculate outlet temperature to compare with the assumption of step 2 7-Repeat steps 2 through 6 until a satisfactory agreement is obtained. It is obvious that the NTU method is much more straight forward in this latter problem so in this study NTU method is used. In additional to all given above, in this study heat performans test is implemented for a vehicle radiator. The test results are checked by using NTU method in order to compare the diffrence of outcome between the test and calculation results. Consequently, by using the heat transfer design procedures, for the specified radiator; (Retava-j) Reynolds number-j factor of heat transfer coefficient for air side and radiator performance characteristic curves are obtained. xvii 96

Bu çalışmada birinci bölümde konuya bir giriş yapılmıştır. İkinci bölümde kompresörlerin standartlaştırılmış sınıflandırması hakkında kısa bir bilgi verilerek, kompresör deneylerinin daha anlaşılabilir olmasının sağlanması amaçlanmıştır. Kompresörlerin tiplerine göre, hesaplanması gereken düzeltmeler farklılık göstermektedir, bu durum kompresör deneylerinin biraz daha karmaşık bir hal almasına neden olmaktadır. Üçüncü bölümde kompresör tip deneyine geçilmiştir, kompresör tip deneyi, belirli sayıda üretilen kompresörün bir modelinin tipik performansını belirlemek için standartlara uygun olarak yapılan bir tam performans deneyidir. Bu bölüm, bu deneyle ilgili bazı tarifler, ölçme cihaz ve metotları, deney metodu, hata analizi, tolerans hesaplamaları ve sonuç bölümlerinden oluşmaktadır. Dördüncü bölümde sadeleştirilmiş kompresör deneyinin ana hatları anlatılmıştır. Bu deney seri üretim kompresörlerine uygulanabilen bir performans deneyidir. Sadeleştirilmiş kompresör deneyi, kompresörün hacim debisi ile kompresörü tahrik eden güç ölçmelerinin imalatçının normal deney alet ve cihazlarını kullanarak yapabildiği deneylerdir. Bu bölüm de deneyle ilgili bazı tarifler, ölçme cihaz ve metotları, deney metodu, hata analizi ve sonuç bölümlerinden oluşmaktadır. Beşinci bölümde dört çeşit kompresör deney örneği verilerek ( Pistonlu-gel git hareketli, komple yağ yüzmeli döner vidalı, portatif yağ yüzmeli döner vidalı ve sıvı segmanlı basitleştirilmiş kompresör deney örnekleri) konunun daha iyi anlaşılabilir olması amaçlanmıştır. Bu çalışma ile kompresör deneylerinin standartlara uygun olarak yapılabilmesi için bir yöntem oluşturulmuştur. In this study, it is given some acknowledgements about of the compressors. In the third chapter, it is mentioned type tests of compressors. According to the type of compressors, corrections required is differed from each other. This situation causes the test of compressors to be more complicated. Type test of compressor is a exact performance test that specify the performance of a model compressor produced specific amount. This chapter includes a specifications about this tests, instruments about measuring, error analysis, calculations of tolerance and conclusions. In fourth chapter, in the simplified compressor test, flow rate of compressor and compressor power are used. In the fifth section, it is given four kind of test examples and aimed to be more understanding relation to the issue. With this study, it is taken place a method to fit to the standards. It is specified in TS 7765 how acceptance tests for displacement (displacing) type compressor (displacing) are conducted. The thing to be carried out is to find deviation amount by error analysis and similar methods and compare found deviation amounts with allowed limits. Table 3.1 indicates fluctuations according to the average of highest deviations among previously determined values and found values. According to these tests, deviation value of the values measured by test is specified as %. In order to accept the tests, deviation amounts must be within such values. Test results are controlled according to Table 4.1. Capacity and specific energy consumption values of compressor being tested are calculated according to these results. If the values guaranteed by manufacturer or the party ordering the tests are within the deviation values given in table 4.1 as percent, the compliance of compressor is confirmed. Otherwise, the compliance of compressor is not confirmed. Volumetric capacity measurements are carried out according to TS 7253. 86

Tez (Yüksek Lisans) -- İstanbul Teknik Üniversitesi, Fen Bilimleri Enstitüsü, 2000, Thesis (M.Sc.) -- İstanbul Technical University, Institute of Science and Technology, 2000, Vantilatör istenen gaz debisini gazı hareket ettirebilecek kadar enerji vererek sağlayan cihazdır. Vantilatör tiplerinin parametrik tasarımı üzerine yapılan bu çalışma üç ana bölümde incelenebilir. Birinci bölümü öne eğik kanatlı, geriye eğik kanatlı ve radyal merkezkaç vantilatör tipleri ve boru tipli yönlendirici kanatsız eksenel vantilatörler için elde edilmiş teorilere dayanılarak kullanıcının istediği özelliklere uygun olarak parametrik vantilatör tasarımı yapan QBASIC tabanlı programların hazırlanması oluşturmaktadır. İkinci bölümü ise, bir CAD-CAM-CAE programı olan I-DEAS üzerinde, QBASIC programlarında hesaplanan vantilatör boyutlarının kullanılması ile, tasarımın bilgisayar ortamında modelenmesini sağlayan programların hazırlanması oluşturmaktadır. Bu programlar I-DEAS üzerinde çalıştırıldığında merkezkaç ve eksenel vantilatörler için parçaları çizmekte, montajı oluşturmakta ve iki boyutlu teknik resmini çıkartmaktadır. Üçüncü bölümü ise, vantilatör tasarımda rol alan teoriler henüz tam olarak geliştirilemediği (özellikle merkezkaç vantilatörler) için gerçekleştirilen tasarımların Fluent adlı yazılım yardımıyla akış analizin yapılarak vantilatörün karakteristik eğrilerinin ve çalışma bölgesinin belirlenmesi oluşturmaktadır., Fan is the tool which gives energy required to move the gas in order to obtain mass flow. This study can be divided in three groups. First part covers QBASIC based program used for user dependent parametrical design of a fan. Theory used in this program is about forward curved, backward curved and radial bladed centrifugal fans; pipe type axial fans without guide vanes. Second part covers the solid modelling of the fan with a CAD-CAM-CAE software named I-DEAS. QBASIC output including exact fan dimensions is used as input of I-DEAS software. When user runs a program file I-DEAS draws impeller, scroll, inlet nozzle of fan, assembles all those parts and creates 2D drawings. Besides, 3D model for impeller of axial fan and related 2D drawings are created.In the third part, flow analysis of the related design is generated by the help of software named FLUENT as theories supporting the design of fans (especially centrifugal fans) haven’t been competely developed. Software creates characteristical curves and operational region of the fan., Yüksek Lisans, M.Sc.

ÖZET BİLGİSAYAR YARDIMIYLA VANTİLATÖR TASARIMI Fuat Hakan DOLAY Vantilatör istenen gaz debisini gazı hareket ettirebilecek kadar enerji vererek sağlayan cihazdır. Vantilatörler merkezkaç ve eksenel tiplerde olmak üzere sanayi, işyeri ve konutlarda; iklimlendirme, soğutma, ısı ekonomisi ve ısının geri kazanılması gibi çok çeşitli amaçlarla yaygın olarak kullanılmaktadır. Bu kadar yaygın kullanılan vantilatörlerin doğru seçimi işletmenin verimli, sürekli ve ekonomik çalışmasını sağlayacaktır. Yapılan çalışma üç ana bölümde incelenebilir. Birinci bölümü öne eğik kanatlı, geriye eğik kanatlı ve radyal merkezkaç vantilatör tipleri ve boru tipli yönlendirici kanatsız eksenel vantilatörler için elde edilmiş teorilere dayanılarak, kullanıcının istediği özelliklere uygun olarak parametrik vantilatör tasarımı yapan QBASIC tabanlı programların hazırlanması oluşturmaktadır. Programlar kullanıcının girdiği debi, basma yüsekliği, motor devir sayısı gibi akışı belirleyici özelliklere uygun olarak vantilatör tasarımını gerçekleştirmektedir. İkinci bölümü ise, bir CAD-CAM-CAE programı olan I-DEAS üzerinde, QBASIC programlarında hesaplanan vantilatör boyutlarının kullanılması ile, tasarımın bilgisayar ortamında modelenmesini sağlayan programların hazırlanması oluşturmaktadır. Bu programlar I-DEAS üzerinde çalıştırıldığında merkezkaç vantilatörler için çark, salyangoz ve difüzörü çizerek montajını kurmakta ve iki boyutlu teknik resmini çıkartırken, eksenel vantilatörün çarkının üç boyutlu modelini ve teknik resmini çıkartmaktadır. Üçüncü bölümü ise, vantilatör tasarımda rol alan teoriler henüz tam olarak geliştirilemediği (özellikle merkezkaç vantilatörler) için gerçekleştirilen tasarımların Fluent adlı yazılım yardımıyla akış analizin yapılarak vantilatörün karakteristik eğrilerinin ve çalışma bölgesinin belirlenmesi oluşturmaktadır.Bu şekilde bazı ampirik formüller aracılığı ile geliştirilen vantilatör modellerinin istenen değerleri karşılayıp karşılamadığı incelenebilir. Yapılan çalışmalarla, istenilen özellikleri sağlayan vantilatörler için tasarım tamamı ile otomatikleştirilerek bilgisayar ortamına taşınmıştır. İleride yapılacak çalışmaların, XIşu ana kadar sadece akış üzerinde gelişen çalışmaları mukavemet analizleri ile desteklenmesi uygun olacaktır. SUMMARY COMPUTER AIDED VANTİLATÖR DESIGN Fuat Hakan DOLAY Fan is the tool which gives energy required to move the gas in order to obtain mass flow. Fans, grouped as centifugal and axial, are used in industry, office and resident; for climatization, cooling, heat economy, heat recycling. Being widely used, it's very important to make the right choice for efficient, continuous and economical corporations. This study can be divided in three groups. First part covers QBASIC based program used for user dependent parametrical design of a fan. Theory used in this program is about forward curved, backward curved and radial bladed centrifugal fans; pipe type axial fans without guide vanes. User inputs data like mass flow rate, pressure rise and rpm of the motor. Second part covers the solid modelling of the fan with a CAD-CAM-CAE software named I-DEAS. QBASIC output including exact fan dimensions is used as input of I- DEAS software. When user runs a program file I-DEAS draws impeller, scroll, inlet nozzle of fan, assembles all those parts and creates 2D drawings. Besides, 3D model for impeller of axial fan and related 2D drawings are created. In the third part, flow analysis of the related design is generated by the help of software named FLUENT as theories supporting the design of fans (especially centrifugal fans) haven't been competely developed. Software creates characteristical curves and operational region of the fan. The aim of this part is to check if fan models created by the help of some emprical formulas meet the required values. Fans, meeting the required values, are completely automated and moved to computer environment. From now on, vantilatör design can be examined in terms of strength and structure. 130

Tez (Yüksek Lisans) -- İstanbul Teknik Üniversitesi, Fen Bilimleri Enstitüsü, 1999, Thesis (M.Sc.) -- İstanbul Technical University, Institute of Science and Technology, 1999, Bu çalışmada, gövde boru tipi ısı değiştiricileri, soğutucu ve nem alıcı serpantinler ile ısıtıcı serpantinlerin tasarımı ve boyutlandırılması Visual Basic 5.0 bilgisayar programı yardımıyla gerçekleştirilmiştir. Veri olarak girilen değerler ile gövde boru tipi ısı değiştiricisinin matematiksel modeli oluşturulmuş, ısıl ve maliyet analizleri çeşitli gövde ve boru çapları için yapılarak optimum ısı değiştiricisi boyutu elde edilmiş ve son olarak AutoCAD R14 çizim programında teknik resmi çizdirilmiştir. Soğutucu ve nem alıcı serpantin için benzer bir ısıl ve maliyet analizi yapılmış ve örnek bir resmi çizdirilmiştir. Isıtıcı serpantin için ise belli olan boyutların ısıl hesaplamalar sonucundaki kontrolü gerçekleştirilmiştir. Gövde boru tipi ısı değiştiricileri için elde edilen sonuçlar pratikte kullanılan değerlerle karşılaştırıldığında kabul edilebilir bir yakınsaklığın sağlandığı gözlemlenmiştir., In this study, designing and sizing of the shell and tube type heat exchangers, air cooling and dehumidifying coils and heating coils with the computer program Visual Basic 5.0 are examined. The mathematical model of the shell and tube type heat exchanger is formed with the values entered as datas, thermal and cost analysises are done for the different shell and tube diameters, optimum exchanger dimensions are found out and finally the technical drawing of the exchanger is obtained in drawing program AutoCAD R14. Similarly, thermal, cost analysises and a sample drawing are attained for the air cooling and dehumidifying coils. For the heating coils, definite dimensions are controlled by the help of heat transfer calculations. When the results of the shell and tube type heat exchangers are compared with the ones being used in the industry, agreeable proximities are observed., Yüksek Lisans, M.Sc.

BİNALARDA ISI KAZANCINA BAĞLI SOĞUTMA YÜKÜNÜN BİLGİSAYARLA HESABI ÖZET Ortam ısı kazancı, verilen bir andaki zamanda ortama doğru üretilen ısı girişi oranıdır. Isı kazancı şu şekilde sınıflandırılabilir; 1. ortama giren ısı kazancı tarzına bağlı kazanç, 2. ısı kazancı duyulur veya gizli kazanç mı olduğu. Bu sınıflamalardan ilki, enerji transferindeki farklı tarzları hesaplamada kullanılan farklı temel prensipler ve eşitlikler nedeniyle gereklidir. Aşağıdaki enerji transferi şekilleri ile ısı kazancı meydana gelir. 1. Yan saydam yüzeylerden güneş ışınımı, 2. Dış duvar ve çatıdan ısı iletimi, 3. iç bölümler, tavan ve zeminden ısı iletimi, 4. Ortamdaki kişiler, ışıklar ve cihazlardan üretilen ısı, 5. Dış havanın, havalandırma ve enfiltrasyonun sonucu enerji transferi, 6. Çeşitli ısı kazançtan. ikinci sınıflandırma, duyulur veya gizli ısı kazancı, soğutma donanımının doğru seçilmesinde önemlidir. Herhangi veya tüm mekanizmalardan, iletim, taşınım veya ışınım tarafından şartlandınlmış ortama doğrudan ilave ısı olduğu zaman, ısı kazana duyulur ısı kazancıdır. Ortama nem ekleniyorsa, ısı kazancı gizli ısı kazancıdır. Ortam soğutma yükü, ortam hava sıcaklığını sabit bir değerde tutmak için ortamdan ısıyı atmak zorunda olduğu orandır, öyle ki, herhangi verilen bir zaman için toplam ısı kazancı mutlaka aynı zamanda ki soğutma yüküne eşit değildir. Işınım tarafından ortam ısı kazancı, kısmen ortam bileşenleri ve yüzeyler tarafından yutulur ve belli bir süre kadar oda havasına tesir etmez. Işınım enerjisi ilk olarak ortam içindeki malzeme ve ortam etrafındaki yüzeyler tarafından yutulmak zorundadır. Yakın bir zamanda yüzeyler ve nesneler, ortam havasından daha sıcak olur ve ısılarının bir kısmı ortam havasına taşınım yoluyla iletilmiş olacaktır. vıı1972 yılında ortam soğutma yükü hesabında iki method kullanılıyordu. Bunlardan ilk yöntem Toplam Eşitlik Sıcaklık Farkı (TETD) methodudur. Bu method da, çeşitli bileşen ısı kazançlan, anlık toplam ısı kazancı oranını oluşturmak için toplanılır. Bu, önceki saat periyotlanndaki değerlerle bağlantılı bileşenlerin ışınım yükleri ortalaması tekniği Zaman Ortalaması (TA) tarafından anlık soğutma yüküne çevrilir. İkinci method Transfer Fonksiyonu Methodudur (TFM). İlk methoda benzer prensipleri olmasına rağmen, bu method, ısı kazancının soğutma yüküne dönüşmesinde ıs depolanması etkisini açıklamak için, şimdiki zaman kadar önceki birçok zamandaki soğutma yükü ve ısı kazancı değerlerine uygulamak üzere bir ağırlık faktörü serisi kullanır. TETD/TA ve TFM arasındaki çelişkileri ortadan kaldırmak için yeni bir method geliştirilmiştir. Bu method için, araştırmacılar, güneşli duvar ve çatılardan ve camlardan iletim tarafından olan soğutma yükünün direkt tek adımda hesaplanması için Soğutma Yükü Sıcaklık Farkı (CLTD) değerlerini çıkarmak için, yöntem bilimi ve transfer fonksiyonu methodunun basit eşitliklerini kullandı. Aynca, camdan güneş tesiri ile ve iç ortam kaynaklanndan olan yükleri benzer tek adımda hesaplanması için Soğutma Yükü Faktörü (CLF) değerlerini geliştirdi. CLTD'ler ve CLF'lerin her ikisi de ısısal method tarafından kullanılan çift adımlı yöntem yerine, ısısal depolanmanın sebep olduğu zaman gecikmesi etkisini içerir, uygun olarak kullanıldığı soğutma yükü hesaplannın elle yapılan ortalamalann oluşmasına müsaade edilmesi, TFM formundaki değerlere uygun sonuçlar üretir. Ortam soğutma yükünü hesaplamak için, detaylı yapı dizayn bilgileri ve seçilen dizayn koşullarındaki hava bilgileri gereklidir. Genellikle aşağıdaki işlemler yapılmalıdır; a) Binanın özelliklerinin elde edilir. b) Binanın yerleşiminin, yön tayininin ve dış gölgelenmelerinin tayin edilir. c) Dış ortam dizayn koşullan seçilmeli ve uygun hava verileri elde edilir. d) İç ortam dizayn koşullan seçilir. e) Aydınlatma, ortamdaki kişiler, iç donanımlar ve cihazların çalışma aralıkları elde edilir. f) Soğutma yükü hesabı yapmak için ay ve saat seçilir. VIIIBu işlemler tanımlandıktan sonra, tüm enerji ileten kaynakların ısı kazançtan hesaplanır. Daha sonra, ısı kazançları, seçilen saat ve aya için uygun CLTD'ler ve CLF'ler kullanarak duyulur veya gizli soğutma yüküne dönüştürülür. Son olarak toplam soğutma yükünün bulunması için toplam duyulur soğutma yükü ile toplam gizli soğutma yükü beraber toplanır. Ancak, maksimum soğutma yükünün bulunmasını istiyorsak bunun için kritik zamanın tayin edilmesi gerekir. Kritik zaman tayinin doğru bulunması için tüm zamanlann ve bu zamanlardaki ortam bileşenlerinden olan ısı kazançlannın göz önüne alınması gerekir. Bu işlem çok vakit alacağından dolayı bazı kabuller ve basitleştirmeler yapılarak maksimum soğutma yükü bulunur. Bulunan bu değer gerçek maksimum soğutma yükü değerine eşit değildir, fakat yakın bir değerdir. Doğru değerin bulunması için çok hızlı işlem yapabilen bilgisayarlar bu işlem için kullanılabilir. Böylece, tüm veriler yılın 365 günü için karşılaştırılarak maksimum soğutma yükü zamanı kesin olarak bulunabilir. IX CALCULATING SPACE COOUNG LOAD WITH COMPUTER PROGRAMMING SUMMARY Space heat gain is the rate which heat enters into is generated within a space at a given instant of time. Heat gain is classified by 1. the mode in which it enters the space and, 2. whether heat gain is a sensible or latent gain. The first classification is necessary because different fundamental principles and equations are used to calculate different modes of energy transfer. Heat gain occurs with energy transfer like following forms. 1. Solar radiation through transparent surfaces, 2. Heat conduction through exterior walls and roofs, 3. Heat conduction through interior partitions, ceilings and floors, 4. Heat generated within the space by occupants, lights and appliances, 5. Energy transfer as a result of ventilation and infiltration of outdoor air, 6. Miscellaneous heat gain. The second classification, sensible or latent heat gain, is important for proper selection of cooling equipment The heat gain is sensible when there is a direct addition of heat to the conductioned space by any or all mechanism of conduction, convection or radiation. The heat gain is latent when moisture is added to the space. The space cooling load is the rate at which heat must be removed from the space to maintain room air temperature at a constant value. Note that the sum of all space instantanous heat gains at any given time does not necessarilyequal the cooling load for the sapce at that same time. The space heat gain by radiation is partially absorbed by the surfaces and contents of the space and does not affect the room air until later. The radiant energy must first be absorbed by the surfaces that enclose the space and the material in the space. As soon as these surfaces and objects become warmer than the space air, some of their heat will be transferred to the air in the room by convection. Two methods of calculating space cooling load were using in 1972. The first procedure was the Total Equivalent Temperature Differential (TETD) Method. In thismethod, various components of space heat gain are added together to get an instantaneous total rate of space heat gain. This is converted to a instantaneous space cooling load by the Time Averaging (TA) technique of averaging the radiant load companents with related values from a period of preceding hours. The second procedure is the Transfer Function Method (TFM). Altough similar in principle to the first method, it employs a series of weighting factors to apply to heat gain and cooling load values from several previous hours as well as the current hour, in order to account for the thermal storage effect in converting heat gain to cooling load. To eliminate any discrepancy between the TETD/TA and TFM, the new method was generated. For this method, investigators used to methodology and basic equations of the Transfer Function Method to generate Cooling Load Temperature Differential (CLTD) data for direct one-step calculation of cooling load from conduction heat gain through sunlit walls and roofs and conduction through glass exposures. Also developed were Cooling Load Factors (CLF) for similar one- step calculation of solar load through glass and loads from internal sources. Both CLTDs and CLFs include the effect to time delay caused by thermal storage, instead of the two-step procedure used by thermal methods described, allowing the creation of a manual means of calculating cooling loads that when appropriately used, produces results consistent with those form the TFM. To calculate a space cooling load, detailed building design information and weather data at selected design conditions are required. Generally, following procedures should make; a) Obtain characteristics of the building. b) Determine building location, orientation and external shading. c) Obtain appropriate weather data and select outdoor design conditions. d) Select indoor design conditions. e) Obtain a proposed schedule of lighting, occupants, internal equipment and appliances. f) Select the time of day and month to do the cooling load calculation. When these procedures are defined, heat gains of all energy transfer supplies are calculated. Then, the heat gains are coverted to sensible or latent cooling load with using appropriate CLTDs and CLFs for selected time of day and XImonth. Finally, the total sensible cooling load and the total latent cooling load are added together to find totat space cooling load. On the other hand, if we want to find maximum cooling load, the critical time was determined. To determine the critical time, we have to calculate heat gain from components in the space at all time of year. So, the calculation takes more time, the maximum cooling load can be calculated with some eliminations and simplifyings. This result does no equal the real maksimun cooling load, but it is a near value. To find true value, for this calculation, we can use the computers that makes calculation quickly. Thus, the maximum cooling load time was found corectly with all data are compared for 365 days of year. XII 65

BİLGİSAYAR YARDIMIYLA ISI DEĞİŞTİRİCİSİ TASARIMI ÖZET Bu çalışmada genel anlamda ısı değiştiricilerinin tasarımının bilgisayar yardımıyla gerçekleştirilmesi incelenmiştir. Öncelikle ısı değiştiricilerinin tanımı, sınıflandırılmaları, çalışma şekilleri ve kullanım alanları belirtilmiştir. Visual Basic 5.0 bilgisayar diliyle dört ayrı bölümden oluşan bir program hazırlanmıştır. Amaç, bilgisayarda programlama yoluyla ısı değiştiricilerinin ısıl hesaplarının yapılması ve boyutlandınlmasıdır. Programın ilk bölümü matematiksel modeli incelenmiş olan gövde boru tipi ısı değiştiricisi ile ilgilidir. Akışkanların giriş, çıkış sıcaklık değerleri, kütlesel debileri, ekonomik faktörler programa veri olarak girilir ve maliyet analizi ile birlikte optimum ısı değiştiricisi boyutları bulunup boru dizilişlerinin resmi çizdirilir. Yöntem olarak şöyle bir yol izlenmiştir: Değişik gövde ve boru çapları için tek tek ısıl ve maliyet hesapları yapılmıştır. İçlerinden en ucuz maliyete sahip olan ısı değiştiricisi çözüm olarak kabul edilmiş ve çözümler sanayide kullanılan ısı değiştiricileri ile karşılaştırıldığında bir yaklaşıklık gözlemlenmiştir. İkinci bölümde soğutucu ve nem alıcı serpantin tasarımı incelenmiştir. Serpantinin ısıl hesapları yapılıp, kullanılan akışkanlar (su ve hava) için yapılıp soğutma ile birlikte neni alıcı olarak çalışıp çalışmayacağı belirlenmiştir. Islak, kuru ve toplam yüzey alanları bulunduktan sonra akışkanların çıkış sıcaklıkları ve ısı yükleri hesaplanır. Basınç düşümleri tespit edildikten sonra maliyet analizi yapılır ve son olarak örnek bir resim bilgisayar ortamında çizdirilir. Üçüncü bölüm ısıtıcı serpantin ile ilgilidir. Boyutları veri olarak girilen ısıtıcı serpantinin ısı transferi hesaplan yapılır ve bunun neticesinde bulunan boyutlar ile ilk girilen boyutların uygunluğu kontrol edilir ve bu değerlerin kabul edilecek yakınlıkta olması sağlanır. Son bölümde ise akışkan özelliklerinin bulunduğu program hazırlanmıştır. Bu özellikler ilk üç bölümde gerekli olmuş ve o bölümlerde kullanılmıştır. Bu özellikler ayrı olarak dördüncü bölümde bir kez daha incelenmiştir. Su için sıcaklık değerlerine bağlı olarak yoğunluk, ısı iletim katsayısı, Prandtl sayısı, özgül ısı ve dinamik viskozite gibi değerler bulunmuştur. Hava için ise yine sıcaklığa bağlı olarak yoğunluk, ısı iletim katsayısı, ısı yayılım katsayısı, Prandtl sayısı, dinamik viskozite ve özgül ısı değerlerinin yanında havanın psikrometrik özellikleri belli bir iterasyon sonucunda bulunmuştur. Kullanılan tüm bağıntılar amprik ifadelerdir ve belli bir oranda hata vermektedir. Ancak bu hata oranları ısı değiştiricilerinin tasarım sıcaklığı için oldukça düşüktür ve hesaplamalarda kullanılır. xı COMPUTER AIDED DESIGN OF HEAT EXCHANGERS SUMMARY In this study, computer aided design of heat exchangers is examined generally. At first the heat exchangers' definitions, classifications, working manners and usage areas are determined. A computer program which is formed from four parts is prepared by using Visual Basic 5.0. The object is the obtaining of thermal calculations and sizing of the heat exchangers. The heat exchangers are classified to six parts: Heat transfer process, surface compactness, geometry (construction), flow type, number of fluids, heat transfer mechanism. According to the heat transfer process classification, the direct contact and the indirect contact type of heat exchangers are available. In direct contact type of heat exchangers the heat transfer is obtained from the contact of hot and cold fluids to each other directly which are used in cooling towers and jet type of condensers mostly. In indirect contact type, the thermal energy transfers from a surface or a wall between the fluids that do not touch each other. The ratio of heat transfer surface area to the volume of the system in heat exchangers is the measure of compactness. If this ratio is more than 700 m2/m3, this heat exchanger is called as a compact heat exchanger. In contrary, the heat exchanger is not a compact one if this value is less than 700 m2/m3. The concept of compactness is very important when the replacement area of the heat exchanger is in the first consideration. Classification to the geometry has four types: Tube type heat exchangers, plate type heat exchangers, extended surface type heat exchangers and regenerative type heat exchangers. The tube type of heat exchangers are double pipe, shell and tube type and spiral tube type. The shell and tube type of heat exchanger is the most frequently used one and its thermal design and the sizing is calculated in the first part of the program. The plate type of heat exchangers are gasketed, spiral and lamella. Especially, the gasketed plate type of heat exchangers are being used widely after the shell and tube type heat exchangers. This type is preferred when there are too much disadvantages of the shell and tube type heat exchanger in that application. The heat exchangers that have extended surfaces are plate-fin types and tube-fin types. The tube-fin type heat exchangers are examined in the second part of the program. They are especially used in xnthe air conditioning systems. The regenerative type of heat exchangers are mostly used in steam boilers as air heaters and also they have a very large compactness. Paralel flow, counter flow and cocurrent flow are the types of flow in heat exchangers. Counter flow gives the highest effectiveness. In the classification according to the heat transfer mechanism, there are four items: one phase flow convection in both sides, two phase flow convection in one side and one phase flow convection in the other side, two phase flow convection in the both sides and finally convection and radiation together. The number of fluids are mostly two in heat exchangers but in special processes three or more number of fluids may be used. The shell and tube type heat exchangers are used in heating, refrigeration, air conditioning, power plants, chemical processes and so many similiar areas. They are the structures which the tubes are placed in a shell. They provide high heat transfer surface area in respect of their volumes and the weights. They can be cleaned easily, can be designed for high pressure applications, have a great flexibility for the service requirements and they can be designed and manufactured without any difficulties. This type heat exchanger is most widely used one. The most important parameter in the design of heat exchangers is the shell type. The shell type is chosen according to the TEMA (Tubular Exchanger Manufacturera Association) standarts. Tube arrangements and the baffles are also very important parameters in heat exchanger design. Square and triangle tube arrangements are available but the triangle tube arrangement are mostly preferred because of its tube density, low cost and higher heat transfer per surface area. Also the distance between the tubes is very considerable for the strength of the tube sheet. The distance between the baffles is the function of the shell diameter and it plays a great role in heat transfer so the designer should give great importance to this parameter during the design. In the basic design of heat exchangers, at first the problem should be defined clearly. The flow rates, arrangements (like condensation or boiling), inlet and outlet temperatures and pressures and the other extra datas necessary for the design should be given. In design process, the basic construction of the heat exchanger is selected as a trial form. This trial design is evaluated and the thermal performance and the pressure drops are found. If these found values are in satisfied ranks, this design is accepted. Otherwise, the trial form is changed and and the calculations are done till reaching the desired values. It is preferred to do these calculations by the help of the computers. There are two main problems encountered in the design of heat exchangers: Rating (performance analysis) and sizing problems. In rating problem, an available heat exchanger is examined. Heat exchanger type, dimensions, surface geometry, flow rates of the fluids, inlet temperatures and the fouling factors are known. According to these datas, outlet temperatures, total heat transfer amount and the pressure drops for the both sides can be found out. If an acceptable thermal performance that supplies a value under the maximum pressure drop is found, this is the solution of the design. The second problem is the sizing problem. In this problem, inlet and outlet temperatures, flow rates, surface geometries, allowed pressure drops and the material properties are the datas of xinthis kind of design. Consequently the sizes of the heat exchanger is obtained. Since the necessity of the selection of the heat exchanger type before the thermal analysis, second problem is more difficult than the first one. The cooling and the dehumidifying coils are used in many fields like refrigeration, air conditioning systems, drug and food industry and product storing. The heat of the air transfers to a metal surface, then transfers to the cooling fluid. In this study the cooling fluid is selected as water. Also the moisture in the air, condenses on a cold surface and constitutes wet layer. So it is possible to take the moisture from the air besides cooling the air. Usually fins are used on these coils. The fins are located to the air side of the tubes since the heat convection coefficient here is lower than then the value on the other side. In the design of shell and tube type heat exchangers and the cooling and the dehumidifying coils, the properties of the fluids (usually water and air) are used repeatedly. The amprical equations which are related with the temperature of the fluids are used. The result of these equations are a little bit different form the real values read from the fluid tables. But these differences are acceptable in the design range of the heat exchangers. For water, viscosity, density, specific heat, heat conduction coefficient and Prandtl numbers are found with the data of temperature. For air, heat conduction coefficient, Prandtl number, viscosity, density, specific heat and psychrometric values like wet bulb temperature, relative humidity are found out. As a computer program, Visual Basic 5.0 is used in this study. It is possible of writing computers with Visual Basic which run under Windows operation system. Since widely known software programs that are used in PC's are prepared to be run under Windows, the other conventional computer programs like GWBasic, Qbasic, Fortran, C, Pascal are all outdated. Creating new softwares with these programs are very troublesome because of very long command lines. But with visual computer programs (Visual Basic, Delphi, Visual C) it is very suitable to make programs with rich graphical media. This graphical user interface facilitates the usage and learning of the different kind of applications. In Visual Basic program, the known abilities of the usual Basic are combined with the visual design tools. Menus, fonts, dialog boxes, text boxes, scroll bars and the other graphical elements can be designed fast and easily. It has an event driven structure. Namely, instead of programming softwares in sequences with constant steps, the programmer clicks a window, moves the mouse or selects a command. Creating databases, interaction with the other popular software like CAD programs and data exchange applications are possible. Another advantage of Visual Basic is that the programmed software can be compiled and it can be converted to an executable file. For the heat exchanger program, a main menu is prepared at the beginning. There are four numbered command buttons on a form for the four program parts. Clicking one of these buttons make the connected program part run. Also five labels exist. The color of the forms, the font types and the font sizes on the boxes and buttons are set for the main menu and for the other forms which are used for the other parts of the program. xivThe first part of the program is related with the shell and tube type heat exchangers that at first its mathematical model is found out. Fluids' inlet and outlet values, mass flow rates, fouling factors, total operating time, economical factors like materials unit cost, electrical cost and the rate of increase for fuel are entered as datas and after the cost analysis, the optimum exchanger dimensions and the drawing of the tube arrangements are found out. As a method, the thermal and cost analysis is done for the different diameters of the shells and the cheapest one is accepted as the optimum solution. Besides these results, the number of tubes is also determined. This program part is divided to three sections. The first section is for the water in the shell side and the water in the tube side. The second section is for the steam in the shell type and water in the tube side. Finally the third section is for the water in the shell side and the steam in the tube side. The two phase flow analysis is done for the second and for the third sections. When these solutions are compared with the exchangers used in the industry, a great approximation is seen. In the second part of the program, the design of air cooling and dehumidifying coils is studied. The principal logic is same with the former program part. Air inlet pressure, temperature, specific humidity, mass flow rate, air velocity, water inlet temperature, mass flow rate, water flow velocity are the datas which are entered to the computer. According to these entered values, the thermal calculations of the cooling and the dehumidifying coils run. After the thermal calculations, it is established whether the air cooling coil is working as dehumidifier or not. After wet, dry and total heat transfer areas are determined, the outlet values of the fluids and total amount of heat are found out. Following the pressure drop obtention, cost analysis is done and a sample drawing is created in the computer medium. The third part is about the heating coils. The former program parts' logic continues in this part. The dimensions of the coil, heat load, fluids' inlet and outlet temperatures and the properties and dimensions of the fins are entered as the datas. After the heat transfer calculations, the dimensions found ultimately are compared with the ones entered before. It is tried to get these values have acceptable nearness to each other. The last part of the program is prepared for obtaining the properties of the fluids (water and air) which are used in the first, second and third parts of the heat exchanger program. For shell and tube type heat exchanger design, the water properties; for the cooling and dehumidifying coils the water and air properties with the psychrometric properties and for the heating coils, the water and air properties are necessary. For water, as the function of the temperature, density, heat conduction coefficient, Prandtl number, specific heat, dynamical viscosity values are found. For air, besides these values, psychrometric values are found by using iteration method. These properties are found with the emprical equations and these equations have a certain error rate. But these error rates are acceptable at the range of the design of heat exchangers and after the comparison of tese values with the real ones in the diagrams and tables it can be considered that they can be used reliably. xvConsequently, the design, optimisation and the sizing of the shell and tube type heat exchangers, cooling and dehumidifying coils and the heating coils are determined with the computer. For creating this software, MS Visual Basic 5.0 is used due to its flexibilities, facilities in working and the graphical user interface properties. In the applications, there are very different types of the heat exchangers for different fields. Because of this, the heat exchangers should be designed one by one for the pointed out values. To obtain the optimum design, variable economical factors should be always taken into consideration. This program provides the reducing of design time, the minimizing the design errors and the facilitating of the sizing. Briefly, the computers and the softwares plays the most significant role in the engineering fields recently. It is certain that the importance of the computers will be further in the engineering, industrial and technological fields in the future. xvi 111

ÖZET Plakalı ısı değiştiricilerindeki sinüzoidal dalgalı levhalar arasındaki ısı taşınım katsayısının ve basınç düşüşünün incelendiği bu çalışmada, çeşitli parametrelere göre önceden yapılmış deney sonuçları veri olarak kullanılmış ve sayısal yöntemler yardımıyla incelenmiştir. Elde edilen değerler tablo ve grafiklerle ortaya konulmuştur. Veriler değerlendirilerek levhalar üzerindeki yerel ısı geçişinin, ele alınan geometri üzerindeki değişimi grafikler halinde verilmiştir. Sonuçların irdelendiği bölümde, daha önce yapılan çalışmalarda bulunan değerler elde edilen sonuçlarla karşılaştırılmış ve uyum içerisinde oldukları görülmüştür. Ayrıca, bu çalışma sırasında dikkate alınan ve ihmal edilen diğer noktalar üzerinde durulmuş ve bunların etkisinin ne yönde sisteme ne yönde etki edeceğine ilişkin değerlendirmeler yapılmış ve bundan sonra yapılacak çalışmalar yönünde fikir verilmiştir. SUMMARY This study primarily centers the effect of various physical parameters on the heat transfer rate and heat transfer coefficient of the fluids which flow between horizontal wavy walls. Heat exchangers are used to heat recovery in industry and to obtain different temperatures which is needed in process. Plate heat exchangers is used commonly because of various usage possibilities and their compact structures. A plate exchanger consist of pack of rectangular pressed plates, suspended vertically, and clamped together in a frame by tie bars or screws. The plates are corrugated to increase strength and to improve heat transfer, and are fitted with peripheral sealing gaskets. Four corner ports, which communicate with appropriate connections mounted on the frame, are arranged so that the two liquids flow through alternate passages between adjent plates, usually in countercurrent flow, with heat transfer taking place across the plate surfaces. Plate heat exchangers find their main applications in liquid-liquid heat transfer duties. They are most common in the dairy, beverage, general food processing, and pharmaceutical industries where their ease of cleaning and the thermal control required for sterilization/pasteurization makes themideal. They are also used in synthetic rubber industry, paper mills and petrochemical plants, and in process heaters, coolers and closed-circuit cooling systems. Over 60 different plate patterns have been developed worldwide. Since the geometry of present study has not been studied previously in literature, firsly two adjacent flat walls channels were analyzed and then two adjacent sunizoidal wavy walls channels were analyzed. In the further researches can be also employed to examine graphically dimensionless numbers with mass transfer and velocity distribution. Besides this.various channel geometries can be analyzed. In this study, numerical method was chosen due to limited time and possibility of easy apply. The problem has been modeled and solved using finite difference scheme. Present results are compared to previous experimental studies and industrial usage. In the first step, the finite difference scheme model is arranged 400 grid in the x direction and 24 grid in the y direction. Boundary conditions in the entrance and exit areas can be velocity and pressure for the flow analyses and temperature, heat flux for the heat transfer analyses. The boundary conditions for the geometries in figure 8. 1 can be express as below: > Entrance velocity is constant and horizontal and vertical velocity component is zero. > The velocity on the channel wall is zero. > Gravitational effects are neglected. xii> The physical properties of fluid in the channel such as density, viscosity, specific heat are assumed constant. > Fluid is assumed as isotropic. > The pressure gradian in the exit is zero. > The temperatures in the entrance and exit of the channel are zero. > The entrance and exit effects and secondary flows are neglected. if ``v.iA ?Wrt. (a) (b) Figure 1 The grid domain of the problem. (c) These analyses are compared to the each other and the results of the previous research, [4], and the results of the present study are showed using tables and graphics. The heat transfer rate between flowing fluids in two paralel horizontal sinuzoidal wavy wall is examined and the results are analyzed The results shows that heat transfer coefficient primarily depends upon fluid flux and channel geometry. It is determined minumum and maximum values at certain points on sunizoidal wave using graphical results. Fig. 2 Numax-Re corelation at case (a) Especially, as shown Fig. 2 when the fluid flux is high, this stuation can be seen clearly. The results of the problem for various Reynolds numbers are as below: It can be seen that the higher Reynolds number cause the higher Nusselt number and the higher heat transfer coefficient. On the other hand, when Reynolds number increase, the pressure drop between entrance and exit also increase. It could be seen clearly at Fig. 3. In the high Reynolds number values, there is a big difference between minumum and maximum local heat transfer coefficients on the wavy channel. (see Fig. 4) XUl.7```. ı;` i Kanal geometrisi H_3.8rrs p/h2 hs H_38rre p/h2 ta H_19ms p/h2 te 0.78 Kanal Boyu Fig 3. Pressure distribution along a sinusoidal wavy channel Kanal Geometrisi Pres_3.8ms p/h2 hava Pres_19ms p/h2 hava Pres_38ms p/h2 hava Kanal boyu Fig 4. Local heat transfer coefficient in a sinusoidal wavy channel xiv 88

ÖZET Endüstriyel tüm sistemlerde borulama elemanlarından emilmesi gereken bir ısıl gerilme bulunmaktadır. İdeal verimle çalışan sistemi elde etme çabaları içinde bu ısıl gerilmenin giderilmesi önemli bir sorun oluşturmaktadır. Bu çalışmada bir boru- esnekliği probleminin matrissel formda nasıl çözüleceği verilmiştir. Ama asıl ilgilenilen konu çevresel şartların değişimi ile oluşan bu ısıl gerilmelerin nasıl giderileceğidir. Bu çalışmada kompansatörlerin bu amaca nasıl hizmet ettikleride anlatılmaktadır. Borulama sistemlerinde, ısıl genleşme analizi yaparken teoriyi iyi anlamak ve konuyu anlamayı kolaylaştırmak için, Bölüm 2 de temel bağıntı ve kavramak açıklanmıştır. Bu bölümde borulama problemlerinin matrissel formda nasıl çözüleceği gösterilmiştir. Matrissel çözüme gerek duyulmasının sebebi ise, çok karmaşık olan borulama sistemlerinde elle çözüm elde etmenin neredeyse olanaksız olmasıdır. Matrissel çözüm çözüm yöntemini bir sisteme sokma açısından bir kolaylıktır. Yapılan çalışmalar sonucu bunca karmaşık olan sistemlerin çözümünde sonlu elemanlar yönteminin kullanılmasının uygun olduğu görülmüştür. Bu çalışmada uygulamalar kısmında yer alan problemler sonlu elemanlar yöntemini kullanan ANSYS programı ile çözülmüştür. Ve basit sistemler için elle yapılan çözümler ile ANSYS ile yapılan çözümler karşılaştırılmıştır. Bölüm 3 de kompansatörlerin nasıl kullanılacağını, seçileceğim ve kompansatör kullanırken gereken yardımcı elemanlarının neler olduğu bu yardımcı elemanları seçerken dikkat edilmesi gereken unsurları anlatılır. Bölüm 4 uygulamalardan ibarettir. Buradaki örnekler,düz bir borunun incelenmesinden başlar,bir boru sisteminde Q parçasının anlamını açıklar ve sistemde yapılan konstrüksiyon değişiklikler ile sistemin nasıl iyileştirildiğini anlatır. XVI SUMMARY THERMAL STRESS ANALYSIS OF THE PIPING SYSTEMS Piping systems form one of the most commonly used elements in today's industry and daily use. Pipes are used to carry various sorts of fluids or gases from one place to another with the basic principle of flow from a high pressure to a lower pressure environment. From nuclear energy field to plumbing at our homes, to the natural gas transporting miles long piping systems, they are used everywhere. All different kinds of materials and technology is used to make the piping systems convenient for the individual purpose, to reduce the cost and to provide the longest service life and the reliability possible. Of course, all these require extensive engineering study and research. No material or structure is perfectly rigid. Small or large, every configuration change shape or move under both internal and external loading conditions. Piping systems are no exclusion to this universal law. Under temperature variations, internal pressurization, weight of the structure itself or the material transported, external stressing and because of many other possible reasons, the pipes change their dimensions or their states of stressing. A piping system is a system that travels a distance between two specific points (usually stationary locations). This implies the distance between the start and the end of the pipe being a constant value. Considering the fact that the pipe will change its dimensions during its service, this change in dimension becomes an important issue to be determined since the pipe may elongate or contract more than the acceptable limits and induce extremely high compressive or tensile stresses at the connections (usually named as anchors or terminal points). These stresses may cause the cracking or failure of the pipe itseld or the destruction of the anchors or terminal points. This may not seem to be a considerable problem for a small piping system such as a pipe connecting the furnace in your home to your radiator. However, considering a nuclear plant using a piping system for transporting toxic waste, it sure would be fatal if these pipes cracked and all the toxic waste leaked through the pipes. Due to the above stated importance of the problem, extensive studies are made in the last century on the expansion and flexibility of piping systems. The basic aim of these studies were to determine the magnitude of the stresses or expansions created by expanding pipes, internal to the piping system, and the reactions they exert upon the terminal equipment or anchors, and decide whether or not they are tolerable. Theoretical, as well as experimental and numerical studies are made. Generally, deep theory is avoided due to the time such studies take and their inconvenience to the engineer in the field who is supposed to use them. Tabular and simplified geometry solutions are rather preferred. With the invent of fast computers, matrix form solutions became possible, thanks to the computation speed provided by the computers. This study focuses on the expansion and flexibility analysis of piping systems. Theoretical and practical solution methods are summarized and used in example problems to give the reader a better feel of the application. Deep theory is avoided wherever it is felt that doing so would disturb the focus of the study. Appropriate references are given under these circumstances. XVHChapter 1 is the introduction of the study and its contents. The importance of the problem is emphasized and the possible uses of this study are mentioned. Chapter 2 includes a thorough explanation of piping expansion and flexiblity analysis. Tables that list the mechanical and thermo-mechanical properties of various engineering materials are given. Matrix methods to determine the flexibility of piping systems are explained. The transformations of coordinates to make displacement and stresses in different directions be compatible toeach other are given. The relations between the stresses and deflections at any point in the pipes are derived and presented. The formation of the Compliance and Stiffness matrices and the determination of their components using the normalization and the base point approaches are included. Looped and unlooped systems are both considered. Effects of non-uniform temperature distribution and internal pressurization are lsited. Individual components (such as an elbow or a straight portion of a pipe system) are treated individually and their compliances are derived. Expansion Stress Range and Stress Intensification concepts are introduced. Cold-drawing and its importance is explained. Finally, simplified methods for desk-top calculations are given. These include the simple conservative approach, the graphical and tabular approaches and many other simplified methods. Chapter 3 introduces the bellow joints, ball joints and swivel joints, these element are used either alone or within compensators to absorb the expansion a piping system undergoes. The theory and the studies for these elements are scarce because there is no standart geometry or material while there are many manufacturers which produce them. The best source of information related to these elements would be the individual manufacturers. Bellows are classified in two major groups: 1- That involve sliding - in which there is relative motion of parts 2- That involve flexing - in which there is no relative motion of parts The main purpose is to allow the motion to absorb the expansion while avoiding any leakage of the transported material. To achieve this, many lubricants and materials are developed specific for this industry. In addition, since there is cyclic loading of these elements, fatigue life considerations are made. Chapter 3 gives the sizing methodology for bellow joints to absorbaxial, lateral or angular expansions. This methodology is intended to be as simple as possible to allow the piping engineer do the calculations easily. Fatigue life calculation procedure is explained with an example. Buckling and bending are also mentioned. Chapter 4 is dedicated to compensators. Compensators absorb the expansions, contractions and vibrations that occur during the service life of a piping system and provide a continuous and efficient operation. They include bellow joints, ball joints and swivel joints which have very high flexibilities. They increase the total flexibility of a piping system and are most commonly used where high expansions or stress concentrations are expected in a pipe. The compensators are grouped under three major categories 1- Axial Compensators 2- Lateral Compensators 3- Angular Compensators By using axial compensators, the flow does not have to change its direction, which minimizes the losses within the pipe, therefore pressure is not lost. The structure is nor affected by lateral expansion and it can freely expand. The important point is that XIXthe structure has to withstand the pressure thrust that is exerted by the material flowing inside the system. Axial compensators require very small assemblies and thus, the cost is low. Their drawback is their need for extra strong guides to keep the compensator in line which impose potential economical and technical risks. For a long pipeline, many axial compensators are needed which increase the total cost although they are relatively cheap elements individually. Lateral compensators can absorb both the axial and the lateral expansions. In a system where there are elbows that change the direction of the pipe, the use of lateral compensators reduce the cost appreciably. Their advantage is that they require less guiding as compared to axial compensators and expansion in two directions can be absorbed with a single lateral compensator. Their darwback is that they can be only used where there is a turn in the pipe. In addition, while expanding laterally, the lateral compensator contracts axially and can induce axial stresses on the pipe. Therefore, angular compensators should be used in long piping systems. Angular compensators require change in pipe direction for absorbing expansions. They are the most suitable elements to absorb expansions in two or three directions. Like the lateral compensators, they minimize the need for guides. One major advantage is that unless the pipe is extremely long, they do not require guiding. Any expansion in any direction can be absorbed, threfore they are crucial elements in complicated piping systems. Their drawback is the requirement for large spaces and the turns in the pipes. Chapter 4 focuses on different compensator applications. The bellow dynamics, service life, effects of the level of temperature, expansion, internal pressure, preset, stress frequency, pressure and thermal shock, corrosion and improper installation on the servicability of the compensators are studied. Guides, terminal fixtures (anchors and terminal equipment) and bellow mounts and their importance in piping flexibility is emphasized. Design, piping line planning, guiding, isolation and installation of axial, lateral and angular compensators are explained with individual examples. Finally, three examples including complete design processes are presented followed by rules of thumb of compensator selection. These examples are shown below, Example. 1 A pipe which is subjected to a change intemperature will change it's physical dimensions if it is free, if it is not, will be placed in condition of stree and will exert reactive forces and moments on the equipment at its ends. The basic problem is to determine the magnitude of these stresses in the piping system and reactions, it exerts upon the terminal equipment and decide whether or not are tolerable. First consider the which is fixed at two ends, X X L=lm Basic equation of unit expansion is: aAT=- (1) E Consider a steam pipe of, a =12.5 e -9 1/°C, E=2.1 e5 N/mm2 xxand taking temperature rise 220 °C, we get 12.5 1(T. 220 2.110s normal stress at both ends, as o = 577.5 N/mm2 Acting force, for pipi DN 150 (A=3206.3 mm2) is, F=oA= 1.851 638 N If one and is not fixed, thermel expansion is Ax=aAT =2.75 mm The results are, Example 2. L shapped construction which is fixed two ends c Hf 4*> c Using theory the following give equations are obtained Deformation at B Ay = - EI Fya 3 Mba2 Fh -^ + abAT EA (2) Ax = EI F,b3 Mbb' F a = ^ + aaAT EA (3) Force at B 3EI (4ab + a2)Ax+3b2Ay x ~ ab(a+b) b2 (4) XXIMoment at B 3EI (4ab + b2)Ay + 3a2Ax y ~ ab(a + b) a2 Mb=~c7 - ut (aAx + bAy) ab(a+b) (5) (6) and five unknows Ax, Ay, Fx, Fy, Mb are solved from these five equations for Ay and Ax where interia moment is, *(d;-df) 64 and using these parameters, Moment at A MA=a Fy-MB Moment at C Mc= b Fx-Mc M. d, (7) m (9) (10) (11) Numerical example, Pipe size DN 150, a=10 m, b=10 m Temperature rise AT=220 °C Material steel E=2.1 105 N/mm2, a -12.5 10`6 1/°C The results can be found from hand calculation and program. Some contents of program output are, SBEND TORSIONAL MOMENT 't I ? SBEND SP SDIR SDIR : Direct axial stress SBEND : Maximum bending stress ST : Shear stressdue to torsion SP : Hoop stress SIG 1 : Max. principal stress at outer surface SIG 3 : Min principal stress at outer surface SHEAR FORCE xxuST : Max. equivalent stress intensity at outer surface SEQV : Max equivalent stress at outer surface deformations, forces and moments. The results are XXUl 224

ÖZET Bütün cisimler sıcaklıklarından dolayı elektromagnetik dalgalar şeklinde ışınım yayarlar. Yayılan ışınım maddesel ortam olmaksızın her yönde yayılır. Işınım enerjinin bir kısmı diğer bir cisme geldiğinde cisim tarafından yutulur. Net ısı akısı yüksek sıcaklıktaki cisimden düşük sıcaklıktaki cisme doğru olur. Enerjinin bu şekilde geçişi ısıl ışınım olarak adlandırılır. Bu çalışmada ısı ışınımı problemlerini çözmek için sonlu elemanlarla çözüm yapan ANSYS yazılım programı kullanılmıştır. Ele alınan problemlerin sonlu eleman modeli oluşturulmuş, gerekli olan sınır şartlan uygulanarak ısı akısı ve sıcaklık dağılımı hesaplanmıştır. İlk örnekte sıcaklıkları bilinen iki nokta arasında geçen ısı akısı hesaplanmıştır. Son iki örnek gerek yüzeyler arasında ve yüzeylerden çevreye ışınımla ısı geçişinin olduğu gerekse de ısı akısının olduğu farklı geometrilerde genel ısı ışınımı problemleri için sıcaklık değişimleri hesaplanmıştır. vıı SUMMARY Thermal radiation is one of basic mechanisms for the transfer of energy between two bodies or regions at different temperatures, the energy being conveyed by electromagnetic waves. Therefore, radiation does not require an intervening medium for the transfer of energy, as is needed for conduction and convection. A portion of this energy flux when impinging other bodies is absorbed. As a result, net energy flow occurs from a body of bigger temperature to a body having lower temperature. This mode of energy transfer is termed heatradiation[13,14]. Espanding the Stefan- Boltzmann Law to two- surface radiation equation, the heat transfer rate between two surfaces I and J is : Ql = a.el.cl}.Fl.(0l4-0s4) (1) This equation is not linear and cannot be solved using the linear equation solver. Therefore the equation is expanded as: Q!=ö:£1.ç),j.Fi.(Öi2+Öj2).(Öi+Öj).(ö,-Öj) or (2) 01=1^.(0,-0,) (3) K1 = asl.n).ct.04 (5) S,, = Kronecker delta JI 0, when J*I This equation can be used to construct a single row in the following matrix equation: such that: [C]{Q} = [D]{04} (6) each row J in /C] = cp ` - 1 J Ul *I J ^,1=1,2, N (7) Fi each row J in [D] = (ân-¥n)a,I=1,2,....N (8) solving for {q} : {qMk*]^4} (9) [K* ] = [C ?>] [D] (10) IXEquation (9) is analogous to equation (1) and can be set up for sandard matrix equation solution by the process similar to the steps shown in equation (l)tru (3). {qJ-PC'Kt} (11) [K ] now includes T terms and is calculated iteratively. To be able to include radiationeffects in elements other than LINK31, The ANSYS Program uses MATRIX50 (the substructure elements) to bring in the radiation matrix. MATRIX50 has an option that instructs yhe solution phase to calculate [K1]. The AUX12 utility is used to create the substructure matrix. AUX12 calculate the effective conductivity matrix [Kte]. The view factor, Fu, is defined as the fraction of total radiant energy that leaves surface I which arrives directly on surface J. It can be expressedby the following equation: 1,, cos/?/. cos/?,,.`. __,_..,_. ^J FiFj TC.t Two methods are available to calculate the view factors: the hidden procedure and the non - hidden procedure. The non- hidden procedure calculates a view factor for every surface to every other surface whether the view is blocked by an element or not, the following equation is used and the integration is performed adaptively. 0u=^SI v-2 Fip-Fjq (13) F! P=lq=l n.x The hidden procedure is a simplified method which uses below equation and assumes that all the variables are costant, so that the equation becomes[9]: 7ZT~ Ç'U = -J2-COSy!?I.COSyffJ (14)A space node may be defined to absorb all energy not radiated to other elements. If the model is not closed system then the user must define a space node with its appropriate boundary conditions. The ANSYS program provides three methods for radiation analysis. A brief description of two method is given below[8,10]: 1. LINK31, the radiation link element, is useful for simple problems involving radiation between two points or several pairs of points. LINK3 1 calculates heat flow caused by radiation between two points. The element requires you to specify in the form of real costant:. An effective radiation surface area. Form factor. Emissivity. The Stefan- Bolzmann constant. 2. The AUX12 radiation matrix generator is meant for more generalized radiation problems involving two or more surfaces. The method involves generating a matrix of form factors (view factors ) between radiating surfaces and using the matrix as a superelement in the thermal analysis. You also can include hidden or non- hidden surfaces, as well as a space node that can absorb radiation energy. The AUX12 analysis method consist of three steps[8]:. Define the radiating surfaces. To define the radiating surfaces, you create a superimposed mesh of LINK32 alements in 2-D (2- Dimensional) and SHELL57 elements in 3-D (3- Dimensional ) models. The orientation of the superimposed elements is important: AUX12 assumes that the viewing direction (i.e., the direction of radiation is along +Ze for SHELL57 lements and along +Ye for LINK32 elements. Therefore, you must mesh the superimposed elements in such a way that the radiation occurs from (or to) the proper face. The element orientation is controlled by the order in which the elements nodes are defined (based on the rigt- hand rule ), as shown below: XIThe findings were compared with the analitical solution for heat flow rate and temperature distribution. The error was less than 0.4 %. The comparisons of temperature distributions and heat fluxes have all shown good agreement. XlllThe findings were compared with the analitical solution for heat flow rate and temperature distribution. The error was less than 0.4 %. The comparisons of temperature distributions and heat fluxes have all shown good agreement. Xlll 49

ÖZET Günümüzde üretim ve imalatta kalite kontrolün öneminin artması ile birlikte, metroloji ve kalibrasyon laboratuvarlarına olan ihtiyaç oldukça artmıştır. Laboratuvarların iklimlendirilmesi ile sağlanan ortam şartları, yapılan ölçümlerin doğruluğu ile direk olarak ilgilidir. Bunun yanısıra, iklimlendirme sisteminin dizaynında nem ve sıcaklık değişkenlerinin birbirlerine olan etkilerinin gözönünde bulundurulması, enerji kayıplarını azaltacağı gibi sistemin optimum kontrolünü de sağlayacaktır. Bu çalışmada bu iki konu ayrı ayrı incelenmektedir. Tezin 1. kısmını oluşturan 2. bölümde, ölçme ve kalibrasyon laboratuvarlarının iklimlendirme sistemleri tasarlanırken, hava kalitesi, hava hızı, hava akış şekli, sıcaklık, nem gibi paremetrelerin seçilmesi konusu incelenmiştir. Bu amaçla, labaratuvar iklimlendirilmesinde çok sık kullanılan temiz oda standartlarından da bahsedilmiş ve ihtiyaç duyulan temiz oda klasının elde edilebilmesi için kullanılması gereken filtrelerin seçimi de incelenmiştir. 3., 4. ve 5. bölümlerde Ulusal Metroloji Enstisünde kurulu olan kütle labaratuvarı kontrol hacmi olarak seçilerek, kontrol hacmi ve kurulu olan klima sistemi tanıtılmış ve daha sonra kontrol hacminin matematik modeli kurularak çözülmüş ve bilgisayar çözümleri deney sonuçları ile karşılaştırılmıştır. Kontrol hacmi fanının iki farklı devirde çalıştırılmasıyla elde edilen bilgisayar çözümlerinin, deneysel olarak bulunan çözümlere oldukça yakın eğriler izledikleri görülmüştür. SUMMARY AIR CONDITIONING OF LÂBORATORDES AND OPTİMUM CONTROL Part 1. Introduction : With growing technology, the assurance of measurement in production and quality has become universally influential. The accuracy of the measurement is crucial to the condition and quality of air in the place where it is done. in the air conditioning of offıces and houses, it is enough for people to feel comfortable; neither too höt nor cold. However, in the calibration and measurement laboratories, a lot of parameters, primarily temperature and humidity control should be obtained to achieve accurate measurement. in addition to this, it is desirable to accomplish the design of air conditioning system with minimal energy too. When we consider that, at least %25 of the world energy consumption is used for heating and cooling of the buildings [1], we can understand how important it is to economize on energy. As it is known, temperature and relative humidity are related to each other in the process of air conditioning. The affection of these two parameters with each other should be considered in the design of air conditioning systems. Otherwise, it can cause vibration and energy loses in the system. This thesis has göt two main stages including the following two subjects. in the first stage, the criteria of designing air conditioning systems for laboratories, which consists of temperature, humidity, air quality, air flow patterns, air velocities, noise, pressurization ete. is examined. Although there isn't any standard for measurement and calibration laboratories, the clean room classes can partially assist. Therefore, the clean room classes and also filtering systems to attain. these classes are examined. xiin the second stage, the Mass laboratory which was established in Ulusal Metroloji Enstitüsü was selected as the control chamber and the mathematical model, which consists of temperature and humidity relations, is verified. in the installation of the model, models established by Parmaksızoğlu [1] and Öznergiz [2] facilitate me. Following this, the mathematical model is solved by computer and the results are compared with the experiments from the real system. Part 2. Air Conditioning Systems of Measurement and Calibration Laboratories : in this part of the thesis, parameters f ör designing air conditioning system of laboratories are examined. The main parameters are temperature and relative humidity. A lot of devices in laboratories are sensitive to these parameters. For example, metal parts expand and shrink, zener diode voltage value decrease ör increase with changing temperature. in addition to this, when relative humidity is very Iow, it can cause static electric loads for electronic devices. Air flow patterns and air velocities are other important parameters which should be considered in the design of air conditioning systems. Basically, there are two different air flow patterns; non laminar flow and laminar flow. in the latter, the air moves with uniform velocity along parallel flow lines, with a minimum of eddies. However, in non laminar flow, air moves with nonuniform air patterns. Air velocities for nonlaminar air flow should be between 0.15- 0.45 m/san. Lower velocities may permit particles to settle and higher velocities may generale objectionable local turbulence and drafts. in a laminar flow system, air is introduced evenly from öne entire surface of room, such as ceiling ör a wall, with flow at constant velocity across the room and removed through the enter area of an opposite surface. To provide good dilution and sufficient air motion to prevent settling of particles, velocity should be about 0.45 m/san. [9] Noise and pressurization control are also important for the laboratories too. For normal applications, the noise level is designed to be belovv 65 decibels [9]. Laboratories should be maintained at static pressure suffıciently higher than xiiatmospheric pressure to prevent infîltration by wind ör other effects. Another essential criteria, which should be controlled, is airborne particles in the air. Clean room classes assist to define the quality of air in the laboratories. The first Clean room standard was introduced with American Federal Standard 209 in 1963. After that, this standard was improved and 209-D in 1988 and 209-E in 1992 were issued. Another standard widely used is the German VDI-2083 which is devised by considering Fed. Std. 209-D. Although 209-E is the last version of Fed. Std. 209, 209-D is commonly known and used. Classes in this standard are classifıed class l, 10, 100, 1.000, 10.000 and 100.000, which are based on partide count with maximum number of particles, 0.5 microns and larger. For example, in class 100.000, partide count must not exceed 3.530.000 partide per m3 ( 100.000 particles per ft3) of a size 0.5 microns and larger. Before selecting the clean room class for calibration and measurement laboratories, a decision must be made to how critical this particulate matter is to the process. To achieve the clean room class selected, necessary type of fılters should be used in the air conditioning system. The basic types of air fılters in use today are straining, impingement, interception, diffusion and electrostatic. The basic parameter for choosing filters is their efficiency. Standards used to classify the effıciency of the fılters are ASHRAE 52-76 and EUROVENT 4/5. in these standards, there are two definitions named Weight Arrestance and Atmospheric Dust Spot Efficiency. According to the efficiency of the filters, they are named EU2, EU3 ete. If the efficiency of the filler is more than % 98, DOP test ( Dioctyl Phthalate ) is applied. The filler which has göt DOP effıciency more than %99.97 is called HEPA filler ( High Efficiency Particulale Air Filler ) Part 3. Description of the System That is Controlled : in this part of the thesis, control chamber, which is Mass laboratory established in Ulusal Metroloji Enstitüsü and the central air conditioning system which comprises whole air conditioning units of laboratories, fresh air unit and ehiller units are described. xiiiconsidering thermal storing properties of walls. This enables one to understand the effect of this property in the computer solution. Part 4. Solving the Mathematical Model with Computer and Comparing the Solutions with Experimental Results : In this part, non linear differential equations including temperature, absolute humidity and leakage heat, are solved by using Mathematica Version 2 program and the results are compared with experimental results. Two solutions are obtained by making the fan work at stepl and step2. At the same time, the control chamber is humidified with the fresh air. These solutions are: 1- Fan works in the second step W(0) = 4 gr/kg ( absolute humidity of control chamber at t=0 ) Wa =8 gr/kg ( absolute humidity of fresh air ) T(0) = 20°C ( temperature of control chamber at t=0 ) Ta = 18°C ( temperature of fresh air ) Q =1.8kW ( supplied heat ) cia =0.15 mVsan ( fresh air quantity ) 2- Fan works in the first step W(0) = 4 gr/kg q, = 0.0925 mVsan Wa =8 gr/kg T(0) = 20°C Ta = 18°C Q = 0.8 kW The computer solutions and experiments are done in 2°C outside temperature. So it is expected that the soil temperature under the floor is 10°C and near the wall is 4°C. [16] XVIPart 7. Conclusion and Advice : In this part, firstly, the main parameters which should be considered for designing the air conditioning systems of measurement and calibration laboratories are advised. After that, the comparison of computer solutions with the experimental results is discussed. It is found that, the variation curves of temperature and absolute humidity, found by computer simulation, is very near to the experimental results. xvnconsidering thermal storing properties of walls. This enables one to understand the effect of this property in the computer solution. Part 4. Solving the Mathematical Model with Computer and Comparing the Solutions with Experimental Results : In this part, non linear differential equations including temperature, absolute humidity and leakage heat, are solved by using Mathematica Version 2 program and the results are compared with experimental results. Two solutions are obtained by making the fan work at stepl and step2. At the same time, the control chamber is humidified with the fresh air. These solutions are: 1- Fan works in the second step W(0) = 4 gr/kg ( absolute humidity of control chamber at t=0 ) Wa =8 gr/kg ( absolute humidity of fresh air ) T(0) = 20°C ( temperature of control chamber at t=0 ) Ta = 18°C ( temperature of fresh air ) Q =1.8kW ( supplied heat ) cia =0.15 mVsan ( fresh air quantity ) 2- Fan works in the first step W(0) = 4 gr/kg q, = 0.0925 mVsan Wa =8 gr/kg T(0) = 20°C Ta = 18°C Q = 0.8 kW The computer solutions and experiments are done in 2°C outside temperature. So it is expected that the soil temperature under the floor is 10°C and near the wall is 4°C. [16] XVIPart 7. Conclusion and Advice : In this part, firstly, the main parameters which should be considered for designing the air conditioning systems of measurement and calibration laboratories are advised. After that, the comparison of computer solutions with the experimental results is discussed. It is found that, the variation curves of temperature and absolute humidity, found by computer simulation, is very near to the experimental results. xvn 80

ÖZET Hava şartlandırma sistemlerinde ve endüstriyel proseslerde her zaman için sistemden çekilmesi ve uzaklaştırılması gerekli bir ısı bulunmaktadır. Bir başka deyişle bir soğutma işlemine ihtiyaç vardır. Geçmişte bu işlem doğal bir su kaynağından veya şebekeden çekilen soğuk su ile gerçekleştiriliyordu. Bugün ise şebeke suyu ile bunu gerçekleştirmek pahalıya mal olmaktadır. Aynı şekilde soğutma suyunun doğal kaynaklardan çekilmeside nispeten zor olmakta ve ayrıca ekolojik dengeye zarar vermektedir. İşte bu noktada su soğutma kuleleri sistemdeki atik ısıyı uzaklaştırmak için ideal bir çözüm olmakta ve yukarıda belirtilen sorunların üstesinden gelmektedir.Bu çalışmada soğutma kulelerinden genel olarak bahsedilmekle birlikte,odak noktasını ters akımlı soğutma kuleleri oluşturacaktır. Ters akımlı soğutma kulelerinde ısı ve kütle geçişim irdelerken,teoriyi oluşturmak ve konuyu anlamayı kolaylaştırmak için3ölüm 2 ve 3 'de temel bağıntı ve kavramlar açıklanmıştır.Bu bölümlerde hava ve subuharın ideal gaz olarak kabul edeceğimizden karışımlar ve ideal gazlar anlatılmış,soğutma kulesindeki ısı ve kütle geçişini açıklayabilmek için birarada ısı ve kütle geçişi konusuna değinilmiştir. Daha sonra sırasıyla kütle geçiş katsayısı, Lewis bağıntısı, psikrometrik diyagram, bağıl ve özgül nem, adyabatik yaş termometre sıcaklıkları açıklanrmşur.Bölüm 4'de ASHRAE Handbook esas alarak soğutma kulelerinin çalışma prensipleri, tipleri ve tasarım esasları hakkında bilgi verilmiş ve kulede sis oluşumu açıklanmıştır. Son bölümde ısı ve kütle geçiş bağıntılarından faydalanarak, ters akımlı su soğutma kuleleri için ısı ve kütle geçiş teorisi çıkarılmış ve Merkel eşitliği elde edilmiştir.Buradan kule hacim hesabı için NTU(Number of Transfer Unit) değerini veren bir integral bağıntısı bulunmuştur. Daha sonra giriş ve çıkış şartlan verilen, örnek bir su soğutma kulesi için hacim ve kule yükseldiği hesabı yapılmıştır Bu hesap yapılırken NTU integrali Simpson 1/3 metodu kullanılarak sayısal olarak çözülmüştür.Bu amaçla FORTRAN 77 ile hazırlanmış basit bir bilgisayar programı bölüm sonunda verilmiştir.Aynı örnek problem daha sonra adım adım integrasyon metodu kullanılarak çözülmüş ve sayısal çözümle karşılaştırılmıştır. Daha sonra örnek kule için, kule hacminin yaş termometre sıcaklığı ve su/hava debileri oranı ile değişimi, grafik olarak sunulmuş ve açıklanmıştır. viii SUMMARY Heat and Mass Transfer Analysis of Counterflow Cooling Towers Most air-conditioning systems and industrial processes generate heat that must be removed and dissipated. Water is commonly used as a heat transfer medium to remove heat from refrigerant condensers or industrial process heat exchangers. In the past, this was accomplished by drawing a continuous stream of water from a natural body of water ot a utility water supply, heating it as it passed through the process, and then discharging the water directly to sewer or returning it to the body of water. Water purchased from utilities for this purpose has now become prohibitively expensive because of increased water supply and disposal costs. Similarly, cooling water drawn from natural sources is relatively unavailable because the increased temperature of the discharge water disturbs the ecology of the water source. Air-cooled heat exchangers may be used to cool the water by rejecting heat directly to the atmosphere, but the first cost and fan energy consumption of these devices are high. Cooling towers overcome most of these problems and, as such, are commonly used to dissipate heat from water-cooled refrigeration, air-conditioning, and industrial process systems. A cooling tower cools water by a combination of heat and mass transfer. The water to be cooled is distributed in the tower by spray nozzles, splash bars, or filming- type fill, which exposes a very large water surface area to atmospheric air. Atmospheric air is circulated by (1) fans, (2) convective currents, (3) natural wind currents, or (4) induction effect from sprays. A portion of the water absorbs heat to change from a liquid to a vapor at constant pressure. This heat of vaporization at atmospheric pressure is transferred from the water remaining in the liquid state into the airstream. The temperature difference between the water entering and leaving the cooling tower is the range. For a system operating in a steady state, the range is the same as the water temperature rise through the load heat exchanger, provided the flow rate through the cooling tower and heat exchanger are the same. Accordingly, the range is determined by the heat load and water flow rate, not by the size or capability of the cooling tower. IXThe difference between the leaving water temperature and the entering air wet-bulb temperature is the approach to the wet bulb or simply the approach of the cooling tower. The approach is a function of cooling tower capability, and a larger cooling tower produces a closer approach ( colder leaving water ) for a given heat load, flow rate, and entering air condition. Thus, the amount of heat transferred to the atmosphere by the cooling tower is always equal to the heat load imposed on the tower, while the temperature level at which the heat is transferred is determined by the thermal capability of the cooling tower and the entering wet-bulb temperature. The thermal performance of a cooling tower depends principally on the entering air wet-bulb temperature. The entering air dry-bulb temperature and relatively humidity, taken independently, have an insignificant effecet on fliermal performance of mechanical-draft cooling towers, but they do affect the rate of water evaporation within the cooling tower. The thermal capability of any cooling tower may be defined by tile fblowing parameters : 1. Entering and leaving water temperatures 2. Entering air wet-bulb or entering air wet-bulb and dry-bulb temperatures 3. Water flow rate The entering air dry-bulb temperature affects the amount of water evaporated from the water cooled in any evaporative-type cooling tower. It also affects m flow through hyperbolic towers and directly establishes thermal capability within any indirect-contact cooling tower component operating in a dry mode. The thermal capability of cooling towers for air conditioning is identified in nominal capacity, based on heat dissipation of 1.25 kw per condenser kilowatt and a water circulation rate of 54 mL/s per kilowatt cooled from 35°C to 29.4 ° C wet-bulb temperature. For specific applications, however, nominal capacity ratings are not used, and the thermal performance capability is usually stated in terms of flow rate at specified operating conditions ( entering and leaving water teperatures and entering air wet-bulb and/or dry-bulb temperatures ). According to the location of th fan corresponding to the fill and to the flow arrangements of air and water, currently widely used mechanical draft cooling towers for HVAC&R can be classified into the following categories:. Counterflow induced draft. Crossflow induced draft. Counterflow forced draft Counterflow Induced Draft Cooling Towers In a counterflow induced draft cooling tower, the fan is located downstream from the fill at the air exit. Atmospheric air is drawn by the fan through the intake louver or, more simply, an opening covered by wire mesh. Cooling water from thecondenser or recirculating water from the coil, or a mixture of the two is evenly sprayed or distributed over the fill and falls down into the water basin. Air is extracted across the fill and comes in direct contact with the water film. Because of the evaporation of a small portion o the cooling water, usualy about 1 percent of the water flow, the temperature of the water gradually decreases as it falls down through the fill countercurrent to the exracted air. Evaporated water vapor is absorbed by the airstream. Large water droplets entrained in the airstream are collected by the drift eliminators. Finally, the airstream and drift are discharged at the top exit. The evaporatively cooled water falls into the water basin and flows to the condenser. In a counterflow induced draft cooling tower, the driest air contacts the coldest water. Such a counterflow arrangement shows a better tower performance than a crossflow arrangement. In addition, air is drawn through the fill more evenly by the induced draft fan and is discharged at a higher velocity from the top fan outlet. Both higher exhaust air velocity and even velocity distribution reduce the possibility of exhaust air recirculation. Crossflow Induced Draft Cooling Towers In a crossflow induced draft cooling tower, the fan is also located downstream from the fill at the top exit. The fill is installed at the same level as the air intake. Air enters the tower from the side louvers and moves horizantally through the fill and the drift eliminator. Air is then turn upward and finally discharged at the top exit. Water sprays from the nozzles, falls across the fill, and forms a crossflow arrangement with the airstream. The crossflow induced draft cooling tower has a greater air intake area. Because of the crossflow arrangement, th tower can be considerably lower than the counterflow tower. However, the risk of recirculation of tower exhaust air increases. Counterflow Forced Draft Cooling Towers In a counterflow forced draft cooling tower, the fan is positioned at the bottom air intake, that is, on the upstream side of the fill. Cooling water sprays over the fill from the top and falls down to the water basin. Air is forced across the fill and comes in direct contact with the water. Because of the evaporation of the water, its temperature gradually decreases as it flows down along the fill in a counter-flow arrangement with air. In the airstream, large water droplets are intercepted near the air exit by the eliminator. Finally, the airstream containing drift is discharged at th top opening. Because the fan is located near the ground level, the vibration of the counterflow forced draft tower is small compared with that of the induced draft tower. Also, if the centrifugal fan blow toward the water surface, there is a better evaporative cooling effect over the water basin. However, the disadvantages of this type of cooling tower are the uneven distribution of air flowing through the fill, which is caused by the forced draft fan. In addition, the high intake velocity may xirecapture a portion of the warm and humid exhaust air. Counterflow forced draft cooling towers are often used in small and medium installations. Materials of Construction Materials found in cooling tower costruction are usually selected to resist the corrosive water and atmospheric conditions. Wood. Wood has been used extensively for all static components except hardware. Redwood and fir predominate, usually with postfabrication pressure treatment of waterborne preservative chemicals, typically chromated-copper-arsenate (CCA) or acid-copper-chromate (ACC). These microbiocide chemicals prevent the attact of wood-destructive organisms, such as termites or fungi. Metals. Steel with galvanized zinc is used for small-and medium-size installations. Hot-dip galvanizing ?fter fabrication is used for larger weldments. Hot- dip galvanizing and cadmium and zinc plating are used for hardware. Brasses and bronzes are selected for special hardware, fittings, and tubing material. Stainless steels (principally 302, 304, and 316) are often used for sheet metal, drive shafts, and hardware in exceptionally corrosive atmospheres. Cast iron is a common choice for base castings, fan hubs, motor or gear reduction housings, and piping-valve components. Metals coated with polyurethane and polyvinyl-chloride are used selectively for special components. Epoxy-coal tar compounds and epoxy-powdered coatings are also used fbr key components or entire cooling towers. Plastics. Fiberglass-reinforced polyester materials are used for components such as piping, fan cylinders, fan blades, casing, louvers, and structural connecting components. Polypropylene and ABS are specified for injection-molded components, such as fill bars and flow orifices. PVC is increasingly used as fill, eliminator, and louver materials. Reinforced plastic mortar is used in larger piping systems, coupled by neoprene O-ring-gasketed ball and socket joints. Concrete, masonry, and tile. Concrete is typically specified for cold-water basins of fielderected cooling towers and is used in piping, casing, and structural systems of the largest towers, primarily in the tower industry. Special titles and masonry are used when aesthetic considerations are important. The basic theory of cooling tower operation was first proposed in 1923 by Walker et al. However, the first practical use of the differential equations was developed by Merkel in 1925. He combined the equations for heat and water vapor transfer and used enthalpy as the driving force to allow for both sensible and latent heat transfer. Heat is removed from the water by a transfer of sensible heat due to a difference in temperature levels and by the latent heat equivalent of the mass transfer resulting from the evaporation of a portion of the circulating water. Merkel combined these into a single process based on enthalpy difference as the driving force. The theory used by Merkel requires two main assumptions, namely, that the water loss by evaporation is neglected and that the Lewis number for air/water vapor systems is unity. The theory states that all of the heat transfer taking place at any xiiposition in the tower is proportional to the difference between the enthalpy of air saturated at the temperature of the water at that point in the tower. Quantative treatment of cooling tower performance by dealing with heat and mass transfer separately is very laborious. Therefore the simplifying approximation of MerkeFs enthalpy theory has been almost universally adopted for the calculation of tower performance. MerkeFs differential equation for the cooling tower was redeveloped by Nottage and converted to a graphical method of solution by Lichtenstein. Another graphical prosedüre, for determining the air process line in a cooling tower, was suggested by Mickley. Simpson and Sherwood carried out experimental studies on several small-scale cooling towers, and examined the dependence of the mass transfer coefficient on the various air and water properties. Carey and Williamson extended MerkeFs theory to be applicable to gas cooling and humidification, and proposed the Stevens diagram (devised by W.L.Stevens ) for the solution of the cooling tower integral necessary for determining the required volume of a tower. Baker and Shryock reviewed MerkeFs work and examined the effects of some of the approximations. Further theoritical work on cooling towers has been carried out by Berman, Hsu et al., Threlkeld, Yadigaroglu and Pastor, and Whiller. Extensive sets of curves for cooling tower design, based on MerkeFs theory, have been prepared by the American Society of Heating, Refrigerating and Air Conditioning Engineers. In this study, heat and mass transfer theory of counterflow cooling towers are presented. The performance of counterflow cooling towers is analyzed by applying by so-called Merkel model. This model is based on equations developed in a paper published in German by Merkel (1925). The equations express an energy balance and describe simultaneous mass and heat transfer coupled through the Lewis relation. However, in the interest of tractibility the equations were simplified by omitting a term and as a result do not account for the mass of water lost by evaporation. Given the inlet water and air conditions the Merkel equations predict the enthalpy (hence wet-bulb temperature ) of the outlet air, but not its humidity. The equations also predict the required number of transfer units (NTU) to accomplish the process. XIII 59

ÖZET Binalarda ısıtma sistemi ve yakıt seçimi aşamasında alternatifler arasından en uygun olanının belirlenmesinde birçok güçlükle karşılaşılmaktadır. Proje sorumlusu, uygulayıcı ve bina sahibi arasında karar vermeyi kolaylaştıracak bir çalışma yapılması bu tezin konusunu oluşturmuştur. Değişik ısıtma sistemleri, değişik yakıt kombinasyonları ile yıllık yatı n m maliyeti, yakıt ve bakım işletme maliyetleri karşılaştırmalı olarak tespit edilmeye çalışılmıştır. Ayrıca sistem ve yakıtların maliyet açısından karşılaştırılmasının yanısıra birbirleri ile konfor, çevre kirliliği, bölgeye göre genel politikalar açısından da karşılaştırmalar yapılarak en uygun sistem ve yakıtın seçilmesi kolaylaştırılmaya çalışılmıştır. Bu amaçla İstanbul U. ısı bölgesinde dört katlı bir bağımsız konut örnek olarak seçilmiş ve yakıt olarak doğalgaz, motorin ve elektrik, ısıtma sistemi olarak ise radyatöriü sistem, döşemeden ısıtma sistemi, ısıtma santrali ve hava kanallı sistem, fan coilli ve ısı pompalı sistem, split tip ısı pompalı sistemlerin karşılaştırılması yapılmıştır. ix SUMMARY Today, the energy conservation, economic conditions and the need to prevent the air and environmental pollution in large living areas make necessary to decide on the heat production system of a district according to the results of a detailed and carefully made feasbility study. in this study, six district heating system altematives are considered and analyzed by selecting the area of İstanbul, a four floors house as a sample district. The energy issue necessitates a rational and more effıcient use of the energy resources, and it points out the need for research that should be carried out in the field of new and renevvable energy resources. The economical use of energy and its savings is very important in Turkey because 48 % of the total energy consumption is imported. The consumptions of energy acording to sectors are : 41 % in building heating, 33 % in industry, 20 % in transportation, 5 % in agriculture 1 % in other sectors. Natural gaz is a contemprary kind of fuel vvhich is ashless and smokeless. it is easy to use and does not create air pollution, also does not need to be stocked. The main parameter effecting heating system selection is the fuel kind. Especially introducing natural gaz as fuel for heating systems in recently years, altemative heating system choices have been appered and previous preferences have been changed. in the thesis, the six heating systems are höt vvater boller and radiator heating system, höt water boiler and flooring heating system, höt water boiler, heating ventilation unit and air düet system, heat pump, heating ventilation untt and air düet system, heat pump and fan coil heating system, split type heat pump and air düet heating system. xC/î O ü >- < LU > o /- İLİ H > t» O < UJ X XIIIUnconscious use of energy is continuing to be one of the most important problems of today. Fuel saving in house heting constites the most important part of the energy problem of our country. According this results, hot water boiler and radiator system is cheapest than the other heating systems. And the natural gas is the most convenient fuel for the cities which is completed the piping system of the natural gas like Istanbul, Ankara and Bursa. In the other hand, some properties cooling and air ventilation are advantage for the other some alternatives. Using natural gas was very limited in Istanbul this year. Igdas should indrease it's gas sell by big compains and reach to the potantiel customers. On the other hand burning low quality coal should be prohibited by the local authorities. Serious emmision limits should be brought to coal, fuel oil and natural gas burning hot water boilers. xivC/î O ü >- < LU > o /- İLİ H > t» O < UJ X XIIIUnconscious use of energy is continuing to be one of the most important problems of today. Fuel saving in house heting constites the most important part of the energy problem of our country. According this results, hot water boiler and radiator system is cheapest than the other heating systems. And the natural gas is the most convenient fuel for the cities which is completed the piping system of the natural gas like Istanbul, Ankara and Bursa. In the other hand, some properties cooling and air ventilation are advantage for the other some alternatives. Using natural gas was very limited in Istanbul this year. Igdas should indrease it's gas sell by big compains and reach to the potantiel customers. On the other hand burning low quality coal should be prohibited by the local authorities. Serious emmision limits should be brought to coal, fuel oil and natural gas burning hot water boilers. xiv 54

Önemli mühendislik uygulamalarından olan ve birçok alanda ihtiyaç duyulan soğutma, ısıtma ve havalandırma sistemlerinde Freon adı verilen halojeni! hidrokarbonlar yaygın olarak kullanılmaktadır. Hem bu sistemlerin dizayn ve tasarımında hem mühendislik eğitiminde bu soğutucu akışkanların karakteristik değerlerinin ve termodinamik özeliklerinin bilinmesi gereklidir. Sistem tasarımlarında gittikçe artan bir uygulama alanına sahip bilgisayar kullanımı için akışkanların termodinamik özeliklerinin de bilgisayar yardımıyla hesaplanması büyük kolaylıklar sağlamaktadır. Ayrıca mühendislik eğitiminde ihtiyaç duyulan termodinamik tablo ve diyagramların yeterli ölçüde mevcut olması anlaşılırlık açısından faydalı olmaktadır. Bu çalışmada, Reynolds [l] tarafından verilen dört özelik bağıntısı ve termodinamik bağıntılar kullanılarak soğutucu akışkanların termodinamik özeliklerinin hesabı, tablo ve diyagramlarının oluşturulması ve çevrim çözümlemesine yardımcı olmak amacıyla çevrim çizimi yapılması için bir program hazırlanmıştır. Basic dilinde hazırlanan bu program herbir akışkan için çalıştırılmış ve termodinamik tablo ve diyagramlar laCl 3 karşılaştırılarak doğruluğu kontrol edilmiştir. İlerdeki çalışmalara yardımcı olması düşüncesiyle programa ait detaylı açıklamalar yapılmış ve bir kaynak oluşturması açısından herbir akışkan için program kullanılarak çizimi yapılan LnP-h ve T-s diyagramları verilmiştir. Halocarbon refrigerants are used widely in the cooling, heating and air-conditioning systems some of important engineering applications and needed in many departments. In both the design of these systems and engineering education, characteristics and thermodynamic properties of these refrigerants must be known. It will be very useful to calculate these properties by a computer playing a great role in the system designs from day to day. And also in engineering education, it is useful for understanding to be available enough thermodynamic tables and diagrams for any refrigerants. In this study, by using four property equations developed by ReynoldsC13 and thermodynamic relations, a computer program has been prepared calculating the thermodynamic properties of refrigerants Fll, F12, F13, Fİ4. F22, F23, F114, FC-318, F500, FS02 and forming the tables and diagrams and drawing cooling cycles on LnP-h and T-s diagrams. Thermodynamic properties, mostly considered in the analysis of cooling cycle are temperature, pressure, specific volume, internal energy, enthalpy and entropy. All these properties are the function of each other and define the state of a substance. It is known that the state of a simple compressible substance is completely spec i f i ed by two i ndependen t, in tens ive pr öper t i es. Therefore all the properties of a substance at any state can be calculated once two independent, intensive properties are available. Only three of the properties given above can be directly measured; temperature, pressure, specific volume. However, the important properties in engineering such as internal energy, enthalpy, entropy cannot be measured by using any experimental methods. The calculation of these properties is only possible by developing some thermodynamic relations between these properties and measurable properties by using thermodynamic relations and property equations given in the literature by many researchers. There are many ways developed for this aim. The way used in the study has been explained below: Four basic equations used in the program were formed and their coefficients were determined by ReynoldsE13 after compiling the empirical data of many researchers These four equations are: 1 ^Equation of State CP-v-T Relations}: R T s 1 r -kT/T P=PCT,v>= + E - f A. +B T+C e k 1 v-b 1 = 2 Cv-hj K J -kT/T A +B T+C e + * « £ C15 eaV Cl+caaV > 2DIdeal Gas Specific Heat Equation: c = c CT}= r G Tv_± + G /T2 C2> VÜ VO ** 1. 5 3>Saturation Pressure Equation: i 5 F12 için, F Ln P = F + - + F Ln T + F T C3-a) d i T 3 4 İİ5 F23 için, F Ln P= F + - + F Ln T + F T + F T2 +F T3 C3-b5 d i _ 3 4, 5 Ö ill) Fil. F13, F14, F22, Fİ14, FC-318, FSÛO, F502 icin F y-T Ln P = F + - +F LnT+FT+F Lnty-T5 C3-e5.d^Saturated Liquid Equation:.s < i - 1 > ya İ/2 2 ö = £ D X + D X +.D X C45 X=l- y~ C4-a5 k In the above equations, symbols are defined as follows; T : Temperature CK5 P : Pressure CPa) v : Specific volume Cm /kg> VI IIc : Specific heat at constant volume for an ideal vo gas C J/kg. K 5 P, : Saturation pressure CPa> p : Saturated liquid density Ckg/m 5 T : Critical temperature CK> K R : Gas constant CJ/kg.iO A., B.. C.. G.. D.. F.. b»k,ot,c,y : Empirical constant l I I l J. t Calculation of properties in the superheated vapor region is done as follows: Two additional equations for calculating the internal energy and entropy are developed firstly. The internal energy is usually defined as the function of T and v, u=uCT,v3. By taking its total di f f erantial, a u a u du=C ) dT +C 3 dv C55 a t v a v T Using the definition of specific heat at constant volume a u du= c dT + C - > dv C63 V - T Q v An important thermodynamic relationship, derived in most thermodynamic books [2, 3,43, is a u a p C 5 = T C > -P C7> # v T a T v Substituting Eqn. C73 in Eqn. C63 j= c dT -f v I a p. du= c dT -P - T C > dv C8> a t v J Rewriting the last equation with respect to specific volume v=l/p, dv=dp/p 1 f * P 1 du= c dT + P - T C 3 dp C9> p2 I 3T pi The change of internal energy of a simple compressible substance is associated with a change of state from Cp,T ') to Cp,T'}t shown in Figure 1. Since o o internal energy, u, is a property and total diff erantial, the integration of it is independent from the integration way. So the integration of this equation can be taken first at constant density, than at constant temperature by considering the equation of state and the ideal gas specific heat equation; IKo Cpo,T> T O ü O Sekil -.Cp.T} 1.1. The integration way of calculation of the differences of internal energy and enthalpy by using experimentel data -; c CT>dT + vo t- â p P - T C `> d T P ]dp+uo C10> where, u : internal energy CJ>'kg5 T : reference temperature CIO u : reference internal enerqy at reference state o -` P CJ/kg> density C kg/m 5 This equation can also be written with respect to specific volume replacing density as the variable; -J c CT>dT + vO Û i [ P -T C 5 1 J I 3T VJ dv + u Cll) The enthalpy property is calculated by using a simple thermodynamic relationship» h=u+Pv (. 1 2 3 And the entropy can be determined from the Gibbs equation E 23, Tds=du+Pdv C13> ds=- =- du + - =- dv C14> xSubstituting Eqn.CQ) in Eqn. £14-3 and considering v=l/p, 2 ) dp CİS> P After adding and subtracting a term RdpSp from this equation, if it is integrated in the manner shown Fiq.l, ç CvoCT5 s = j dT - R Lnp J T T O P r * r d p i where, s.- reference entropy at reference state' CJ/kq.K} o Thus for a given temperature and specific volume, the properties u, h, s in the superheated vapor region are calculated by using these developed equations above. These main equations were used in the program. It is seen that all equations are the function of tempereture and specific volume. When two different properties other than T, v are given, the equations must be written with respect to the given properties to calculate the other properties. It is highly difficult to obtain them for any given two properties because of being some logarithmic, exponential and total terms in those equations. For this aim Newton-Raphson methodESD was used by derivating these equations. Detailed explanations were given in Section 3. Calculation of thermodynamic properties in the saturation region proceeds as follows: Saturation temperature, T or saturation pressure, P is given to d d calculate the saturation properties. If T is known, P, d d is calculated from the saturation pressure equation C3> directly. If P is qiven the calculation of T, is done d d by using the Eqn. £3) and the partial derivative of it with respect to T. And by using these saturation values d T,, P, the properties of saturated vapor v,, h » s can d d ODD be found by following the procedure of the superheated vapor properties calculation. And saturated liquid specific volume v is calculated from the saturated a liquid equation. Enthalpy and entropy of saturated liquid, h, s are calculated by using the Clapeyron S B SIequationE23, & P s, -s h, -h * t d V*TV TCv -v } tl7:> 0 1 b a b s where, CdP/dT), is calculated from saturation pressure d equations C3>. So, from Clapeyron equation the enthalpy of saturated liquid is â P h = h -T (v -v K } CIS} a b b s ^ ` d and the entropy of saturated liquid is h -h s a s- - - C195 a b _ Thus all thermodynamic properties in the regions of superheated vapor and saturation are calculated by using these developed equations. The program consists of six subparts and many subroutines directed by a main menu. The parts of main menu enables to use the program easily and give the definition of special aims from the most general aim. When the program is run, firstly the refrigerant is choosen from the subroutine `Akışkan Secimi`. And for chooser» refrigerant, the coefficients of equations are read from the data directory and the name of directories will be used by the `Diyagram` part are determined in the subroutine of `Katsayılar`. After coming to the main menu, all commands and the running of the program are directed by this part. The six parts of the main menu are: 1-Hesaplama : Calculation 2-Tablo : Tables 3-Diyagram : Diagrams 4- Diyagram ve çevrim çizimi : Di agrams&Dr awing cycle 5- Yeni akışkan secimi : New refrigerant S-Cıkıs : Exit The contents and running of these parts are summar i zed below : i-Hesaplama : After determining the entrance values, the properties in the superheated vapor and saturation regions are calculated in this part. It returns again to the `Hesaplama` part for new values after showing the results on the screen. XII2-Tablo : This part enables to prepare tables in the three common form. These tables are prepared due to the changing of saturation temperature, saturation pressure and temperature and pressure. 3-Diyagram : For any refrigerants the diagrams of LnP-h and T-s or both of them can be drawn on the screen by using this part. ?4-Diyagram ve çevrim çişimi : This part allows to draw cooling cycle on the LnP-h and T-s diagrams. The definition of characteristic properties of cooling system i s done manua 1 1 y. 5- Yeni akışkan seçimi : By using this part, it is possible to choose a new refrigerant and to use the program for it, without exiting the program. 6-Cikas : Exit from the program All of these parts run effectively and interactively by calling the prepared subroutines. The flow chart of the program is given in appendix. The program was run for each refrigerant and compared with the tables and diagrams given by Reynolds £13 prepared by using the same equations. For the aim of helping to the studies in the future, equations, coefficients of equations, and variables used in the program and flow chart are given in appendix. In addition, LnP-h and T-s diagrams for each refrigerant have been included to appendix. XI XI 116