Fin HVAC and refrigeration coils have a wide range of applications in various settings, including:

    • Residential: Homes, apartments, condominiums (HVAC systems).
    • Commercial: Office buildings, shopping malls, schools, hospitals (HVAC systems).
    • Industrial: Factories, warehouses, manufacturing plants (industrial process cooling, heating).
    • Automotive: Car air conditioning and heating systems.
    • Food and beverage: Refrigeration systems for food preservation and storage.

4. Design and Construction:

    • Coil size and capacity: Determined by the required heat transfer rate for the specific application.
    • Fin spacing and geometry: Optimized to balance airflow resistance and heat transfer efficiency.
    • Tube arrangement: Can be staggered or in-line, impacting airflow patterns, heat transfer and air static pressure.
    • Coil circuiting: Designed to optimize heat transfer, maintain appropriate fluid velocity, minimize pressure drop,
      and ensure even distribution across all tubes.
    • Material selection: Based on factors like operating pressures, temperatures, corrosion resistance, and cost.

Fin attachment methods:

    • Mechanical bonding:

     Edge tension wound:



     L-foot attachment:



    • Extruded fins:



    • Brazing: Creates a strong and durable bond between fins and tubes using a high-temperature filler metal.

    • Welding: Employed for specific applications and materials requiring a permanent joint.

Cooling coils are basically fin and tube heat exchangers used in any air conditioning installations.  The cooling media is usually water from 42 F, also Ethylene or Propylene glycol water solutions of various concentrations can be used as the cooling media.  Cooling fluid runs in tubes and air runs over fins. Low coolant temperature could result in a coil fin surface temperature that is below dew point temperature of air going through finned surface of a coil, which leads to condensation of moisture in air. The condensation in cooling coils complicates design and selection of cooling coils comparing to heating coils, because we have to account latent heat which is necessary to condense moisture in air. In this case the coil is Wet. When no condensation happens during cooling down air passing through a coil, the coil is Dry. The program considers a coil -Wet, when at least one of two conditions is met: the relative humidity of air is more the 30% and/or Dew Point Temperature of incoming air is more than Entering Water/Glycol Temperature.

There are two methods of calculating heat transfer load in heat exchangers: LMTD and NTU.

The “Logarithmic Mean Temperature Difference “(LMTD) is a logarithmic average of the temperature difference between the hot and cold streams at each end of the heat exchanger. The larger the LMTD, the more heat is transferred.

LMTD method works very well for dry coil runs (one-phase streams), but it is very inaccurate for two-phase streams, in our case – air and moisture in the air, which needs to be condensed. To use LMTD method, we have to know all four temperatures for both steams. Usually we either know or assume outlet temperatures. Then we calculate LMTD with a correction factor (derived as per geometry of heat exchanger and directions of both streams), corrected LMTD or CLMTD. Mass flow of at least one stream is known. Knowing the temperature difference, mass flow and specific heat for the stream, we can calculate total heat transfer rate for the process, overall heat transfer coefficient and heat transfer area of the coil. Based on this data the program is trying to match the calculated coil heat transfer area to the given or assumed coil geometry (given heat transfer area), gradually changing outlet temperature of the stream in question in a loop until the heat transfer areas, given and calculated are matched.

The Effectiveness-NTU method takes a different approach to calculating heat exchange analysis by using three dimensionless parameters: Heat Capacity Rate Ratio (HCRR), Effectiveness (ε), and Number of Transfer Units (NTU). The relationship between these three parameters depends on the type of heat exchanger and the internal flow characteristics. 

The NTU method is useful when minimum data is provided for calculating heat transfer. In our case: the amount of moisture is to be condensed during the process and additional heat load which is necessary to do so and outlet dry and wet bulb temperatures are unknown.

First, the program calculates Heat Capacity Rate for each stream in the process. Then it selects the minimum and maximum Heat Capacity Rates for both steams. Usually the stream with a bigger temperature difference has minimum capacity rate. HCRR is calculated as a ratio between minimum and maximum capacity rates. NTU is calculated as heat transfer area of a coil multiplied by overheat transfer coefficient and divided by minimum capacity rate.

 Effectiveness (ε) can be derived  from graphs or equations for certain types of heat exchangers using NTU and HCRR data.

Heat transfer load for a cooling coil is calculated as maximum possible heat transfer load for Minimum Capacity Rate stream multiplied by Effectiveness (ε). Effectiveness (ε) value is in the range between 0 and 1.

F & T Steam Traps.

F & T traps are the common trap type used for process applications: S&T heat exchangers, steam generators, waters heaters, heating coils, etc.  
F & T traps can handle both light and heavy condensate loads at different pressures and temperatures.
F & T steam trap basically is normally closed mechanical valve, where the valve opens when condensate enters the trap’s chamber, lifts a float, and at that moment the condensate discharges and continues to discharge in a piping system while the float is up. When the trap is filled with steam, the internal pressures in the chamber equalizes and float drops down, closing the valve until the steam in the trap’s chamber turns to condensate.
Thermostatic element (air vent) in the trap is to vent air out of trap.
The air vent is normally open and its function is to pass air and other non- condensable gases into the return line. Once steam arrives at the trap, this element snaps shut preventing steam from passing into the return lines, and its job is finished until the next cycle.
To size properly F & T trap, the following conditions must be known:
• operating inlet pressure
• operating outlet pressure or back pressure
• differential pressure (i.e. the difference between the inlet pressure and the outlet pressure)
• flow capacity of condensate to be removed.

• type of steam system:  Constant (steam flows continuously) or Modulating (steam flow controls by temperature or pressure control valve).  

If the flow rate is unknown, condensate flow rate can be simply estimated by means of the calorific power of the heating apparatus expressed in BTU/hr and by using the following formula: Condensate Flow Rate lb/hr = BTU Rating (in millions)/100.
Every trap is sized with safety factor applied to a condensate flow rate needed to be discharged.
Depending upon manufacturer’s requirements and a particular application, the safety factor might be from 1.5 to 3.

The difference between a trap's inlet (primary) pressure and back pressure is called the 'differential pressure' or ‘pressure differential’. The 'back pressure' is the pressure just downstream of the steam trap. In other words, back pressure is the outlet or secondary pressure of the trap. If the condensate is discharged to the atmosphere just after the trap, the back pressure is considered to be 0 psig. Even if the trap discharges to atmosphere, restrictions in the downstream piping such as elbows, tees and valves may add back pressure.

Vertical lifts (risers) in the condensate piping add back pressure in the form of hydraulic head.   A rule of thumb is that every foot of rise in the condensate line after the steam trap equals ½ psig backpressure on the steam trap discharge. If the condensate discharges into a flash tank and the flash tank pressure increases, the back pressure increases correspondingly.

The maximum allowable back pressure for disc traps is generally 50% to 80% of the inlet of inlet pressure. Mechanical traps on the other hand, have a relatively back pressure of over 90%. A high percentage of steam trap applications will have pressure at the discharge side of the steam trap from the condensate return system. The back pressure may be unintentional or deliberately produced by the design or by the operation of the condensate return system. Unintentional back pressure is caused by a static pressure or rise in condensate piping after the steam trap. The main condensate lines are typically installed at a high elevation and located above the steam traps; therefore, it is necessary to pipe the condensate from the steam trap location up to the higher condensate mains. Undersized condensate lines are another factor that can cause backpressure on the steam trap that must be considered when sizing steam traps. Condensate lines need to be sized for two-phase flow (condensate and flash steam). Another factor that increases condensate line pressure is failed steam traps blowing steam in a condensate line that was not designed for a high volume of steam. 

Formed Heads. Pressure on Concave Side

Introduction:

Formed heads are crucial components in pressure vessels, serving as the end closures. They are designed to withstand internal or external pressure. The selection of a specific head type depends on factors like pressure, temperature, material, and manufacturing feasibility. This article discusses the types of formed heads and the principles of calculations as per ASME Section VIII, Division 1 (UG-32 and Appendix 1-4).

Types of Heads:

Formed heads come in various shapes, each tailored for specific applications. Here are the common types:

• Hemispherical Heads: These heads offer the highest strength-to-thickness ratio, making them ideal for high-pressure applications.
• Ellipsoidal Heads: Ellipsoidal Heads are a popular choice due to their balance of strength and manufacturability. The 2:1 ratio refers to the major to minor axis ratio.
• Torispherical Heads: Commonly used for moderate pressure applications. They consist of a spherical dish and a toroidal knuckle.
• Conical Heads: Employed when a transition in diameter is needed, often used in conjunction with cylindrical shells.
• Toriconical Heads: A toriconical head is a type of pressure vessel head that is a combination of a torispherical head and a conical head. The toriconical head is a stronger and more efficient design than the torispherical head.
• Flat Heads: Suitable for low-pressure applications, typically requiring substantial reinforcement.

The design of formed heads under pressure is governed by codes such as ASME Section VIII, Division 1. Specifically, UG-32 and 1-4 provide guidelines for calculating the required thickness of heads under internal pressure. UG-32 outlines the formulas for calculating the minimum required thickness of heads of revolution, including hemispherical, ellipsoidal, torispherical and toriconical heads. The calculations are based on factors like design pressure, material allowable stress, and head geometry.

Key Concepts in UG-32:

• Design Pressure (P): The maximum pressure the vessel is designed to withstand.
• Allowable Stress (S): The maximum permissible stress for the material at the design temperature.
• Head Geometry: The dimensions of the head, such as inside radius (L), knuckle radius (r), and straight flange length.

Appendix 1-4: Ellipsoidal Heads

Paragraph 1-4 specifically addresses the design of ellipsoidal heads. It provides formulas and guidelines for determining the required thickness based on the 2:1 ellipsoidal ratio and the design conditions.

Key Concepts in Appendix 1-4:

• 2:1 Ellipsoidal Ratio: The ratio of the major axis to the minor axis of the ellipse.
• Thickness Calculation: Formulas are provided to calculate the minimum required thickness, considering the design pressure and material properties.

General Calculation Principles:

The general formulas involve calculating the required thickness (t) based on the design pressure (P), allowable stress (S), and geometric factors. The formulas vary depending on the head type. For instance, for a hemispherical head:

t = (P * L) / (2 * S * E - 0.2 * P)

Where:

• t = minimum required thickness


• P = design pressure


• L = inside spherical radius


• S = allowable stress


• E = joint efficiency

For ellipsoidal and torispherical heads, the formulas are more complex, incorporating factors like knuckle radius and dish radius. Consult ASME Section VIII, Division 1 for the specific formulas and requirements.

Disclaimer: This information is for general knowledge and educational purposes only.