Found in 2008, We are professional industrial valves manufacturers on wellhead control.

24 hours at your service: admin@sgvalves.com

pressure relief device capacity determination.

by:Sino Valves     2021-02-20
Introduction: sufficient mass flow carrying capacity is required for the decompression device (capacity)
Limit the pressure rise of the protected part to prevent its catastrophic failure.
The required minimum pressure relief unit capacity will depend on the specific components being protected and the scenario in which the overvoltage is generated.
Although the maximum discharge capacity is not limited by code and standards, excessive
In order to prevent unstable operation of the discharge device, the size of the discharge valve should be avoided.
ASHRAE standard 15 provides a simple method of capacity determination for pressure vessels and positive displacement compressors (ASHRAE 2007).
Since Standard 15 is not a design guide or manual, it does not provide any additional information about other devices that can apply safety valves.
In this paper, we propose methods for capacity determination of decompression devices for equipment that are not currently discussed in specifications and standards, including: oil separator, housing-and-
Plate tube heat exchangerand-
Frame heat exchanger, oil cooling heat exchanger, product storage tank and steam condenser.
Our principle objective here is to record the basis for the relief device capacity determination to ensure that these other types of protected components remain safe during abnormal operation that may result in high pressure.
While the methodology presented in this paper is intended to be applicable to a variety of refrigeration equipment and operating conditions, it is likely that the dimensions and selection criteria of the decompression device are neatly specified to cover all cases.
Therefore, we should always look forward to the use of reasonable engineering principles and the application of engineering judgment.
Pressure vessel ASHRAE 15 provides specification requirements for determining the minimum capacity of the decompression device that protects the pressure vessel.
The basis for the capacity determination of ASHRAE pressure relief device is the fire condition, the fire condition is the heat generated from the explosion on the projected area of the container (i. e.
The ship was not completely swallowed up).
If there are other heat sources in the container or help to produce steam, these heat sources need to be considered separately. Equation (1)
Displays the formula used to determine the minimum capacity required for the pressure relief valve to protect the pressure vessel (ASHRAE 15: [section]9. 7. 5). [C. sub. r]= [Florin]. D. L (1)where [C. sub. r]
= Minimum required emission capacity [Florin]
= Therefore, the size of the drain valve expressed with lbm/min Air is the same as the basis for protecting the liquid refrigerant
For oil separators, it is recommended to use containers with containers. [
Figure 1 slightly][C. sub. r,os]= [Florin]*D*L (2)where [C. sub. r]
, OS = minimum required emission capacity for reliable devices that protect oil separators (in [lm. sub. m]/min of air)[Florin]
= The Capacity coefficient of the decompression device D = the outer diameter of the oil separator (ft)
L = oil separator length (ft)
Heat exchanger if the heat exchanger is constructed in accordance with the requirements of the ASMEB & PV specification (
Section 8 Section 1
And it is physically stamped, it requires UG-
125 of the B & PV code.
In the case of no need to transfer decompression protection, it is usually necessary to adjust the size of the appropriate \"process\" decompression device to prevent excessive pressure on the heat exchanger in abnormal situations.
The first step in determining the minimum mass flow required for disaster relief protection is to determine the mass flow that may result in over-voltage conditions.
For ships, it is worth mentioning that the scene is due to the external heat load caused by heat dissipation under fire conditions.
Unlike ships, heat exchangers are prone to excessive
Pressurize through the internal thermal load of the product or other secondary fluid flow (e. g. clean-in-place systems).
In both cases, the key consideration for the size of the decompression device is to determine the rate at which the refrigerant is generated by evaporation, which will depend on the heat load and refrigerant properties (
Saturated pressure-
Temperature relationship and steam heat).
It must be emphasized that for all cases under consideration, the rate of refrigerant steam generation needs to be converted to air mass flow, because all decompression devices are rated according to the air base discussed by Reindl and Jekel (2008).
The following sections provide methods for determining the decompression capacity of different types of heat exchangers based on scenarios including external heat sources (
Consistent with the assumption of pressure vessels)
And alternative sources of internal heating.
The actual rescue capacity should be based on one of the larger of these two situations. Shell-and-Tube Shell-and-
Tube-type heat exchangers are used in many refrigeration applications, from packaged coolers (
Evaporator and condenser as watercooled units)
Chillers in industrial refrigeration systems.
Except for the direct-expansionshell-and-
Tube evaporator, cooled fluid (or heated)
Move through the tube
Side of the heat exchanger when refrigerant evaporates (or condenses)on the shell-side.
Almost all the shells todayand-
ASME for tubular heat exchangers-
Rated pressure vessel. When the shell-
The sides of the heat exchanger are configured to contain liquid refrigerant and are able to be isolated from other parts of the system through a stop valve, which complies with the provisions [section]9. 7.
2 of ASHRAE standard 15 (2007)
This requires the application of pressure relief protection dimensions according to the equation (1).
Industrial housing in general-and-
The tube heat exchanger design includes a surge drum located above the shell that cannot be isolated from the heat exchanger (chiller barrel).
Figure 2 shows the on-site installation of this type of housingand-
Tubular chillers common in industrial refrigeration systems.
In this case, the capacity required due to the external heat load is calculated by the sum of the capacity determined from each heat exchanger housing and housing.
If the surge drum can be separated from the barrel, each cylinder should be installed separately with a pressure relief protection device to determine the respective capacity according to their respective dimensions. [
Figure 2:
Figure 3 shows the key dimensions of the submerged enclosureand-
Tubechiller is equipped with a horizontal surge drum that forms the basis for determining the capacity of the release unit due to external heat load, as shown below :[
Figure 3 slightly][C. sub. r]= [Florin]*([D. sub. v]*[L. sub. v]+ [D. sub. s]*[L. sub. s])(3)where: [C. sub. r]
= Minimum required discharge capacity of relief device ([lm. sub. m]/min of air)[Florin]
= Capacity coefficient of discharge device ([lm. sub. m]Air/minuteft. sup. 2]
Estimated area)[D. sub. v]
= Outer diameter of the Hell main container part-and-
Tube heat exchanger (ft)[L. sub. v]
= The length of the main container part of the shell-and-
Heat exchanger (ft)[D. sub. s]
= Outer diameter of surge drum (ft)[L. sub. s]
= Length of surge drum (ft)
In addition to determining the pressure relief capacity of the shell --and-
Tubular chillers estimated by equations (3)
, It is important to consider other situations where refrigerated steam may be generated internally, which may result in high pressure conditions in the heat exchanger.
In most other cases, when the refrigerant side of Shearer is isolated from the refrigeration system, but the second fluid side remains active, an alternative way for the heat energy to be input into the heat exchanger.
Example of heat load that may produce excessive pressure in the Shell-and-
Tube Heat Exchangers may include, but are not limited to, product load and cleaning-in-place (CIP)loads.
The main concern is those heat sources that exceed the saturation temperature corresponding to the maximum allowable working pressure of the heat exchanger (MAWP)
The pressure release device sets the pressure.
If the maximum fluid-
The side supply temperature is less than the saturation temperature corresponding to the heat exchanger MAWP, and the pressure relief capacity can pass the equation (3).
If the maximum fluid-
The side temperature is greater than the saturation temperature corresponding to the heat exchanger MAWP, and the steam generation rate based on the \"internal load\" should be estimated to determine whether the capacity requirements of large decompression devices will produce results.
Since it is not possible to specify a relief device size strategy to address all operational scenarios that may result in over-voltage conditions, good engineering analysis and judgment are essential to address the relief protection of such equipment.
The approach proposed here is intended to illustrate the process that can be followed in determining the pressure release requirements in a particular case.
In the process of considering an internal heat load scenario that may produce an over-voltage situation, the first step is to evaluate the normal capacity of the heat exchanger.
The next step is to estimate the capacity of the heat exchanger under unfavorable load conditions and determine the corresponding rate of refrigerant steam generation.
Finally, the predicted rate of refrigerant steam generation is converted to an equivalent air mass flow rate in order to select a decompression device.
Determining the refrigerant steam generation rate can be done by solving an equation system that characterizes the heat transfer performance of the device, as given by the equation (4)
Balance of two refrigerantsside and fluid-
The side energy flow given according to the equation (5)and (6), respectively.
The system for managing equality is as follows. Q = UA*LMTD (4)LMTD = ([T. sub. return]-[T. sub. supply])/ln[[T. sub. return]-[T. sub. refrigerant]/[T. sub. supply]-[T. sub. refrigerant]](5)Q = [m. sub. fluid]*[c. sub. p,fluid]*([T. sub. return]-[T. sub. supply])(6)Q = [m. sub. refrigerant]*([h. sub. vapor,sat]-[h. sub. liquid,sat])(7)
Among them: Q = heat flow of heat exchanger (Btu/min)
UA = total heat transfer coefficient-area product(Btu/min-[degrees]F)
LMTD = log average temperature difference ([degrees]F)[T. sub. return]= load-
Secondary fluid return temperature ([degrees]F)[T. sub. supply]= load-
Secondary fluid supply temperature ([degrees]F)[T. sub. refrigerant]
= Refrigerant saturation temperature ([degrees]F)[m. sub. brine]
= Mass flow of secondary fluid ([lm. sub. m]/min)[c. sub. p,fluid]
= Thermal capacity of secondary fluid (Btu/lbm-[degrees]F)[m. sub. refrigerant]
= Refrigerant steam generation rate ([lb. sub. m]/min)[h. sub. vapor, sat]
= The enthalpy of saturated steam refrigerant at the set pressure of the fully accumulated decompression device (Btu/[lb. sub. m])[h. sub. liquid, sat]
= The enthalpy of saturated liquid refrigerant at the set pressure of the fully accumulated release device (Btu/[lb. sub. m])In a liquid-
Heat-carrying exchanger, refrigerant temperature ([T. sub. refrigerant])
The assumption is that the saturation temperature record of the response decompression device group (opening)pressure.
Enthalpy of evaporation ([h. sub. vapor,sat]-[h. liquid,sat])
For therefrigerant-
Side energy balance is also evaluated at pressure release device set pressure.
Return fluid temperature of heat exchanger ([T. sub. return])
Estimates are made based on the flow rate of the fluid and the load of the process return fluid, CIP set temperature, etc.
Mass flow rate of fluid on load-
Side of the heat exchanger ([m. sub. fluid])
In addition to the load, also need-
Heat capacity of side flow ([c. sub. p,fluid]).
Nominal value of overall heat transfer of heat exchanger-area product (UA)
Run conditions based on design. Equation (4)
Used to estimate a nominal or design UA.
Once the nominalor design UA has been established, it can be adjusted or corrected to estimate the rate of refrigerant steam generation generated without pressure.
For example, if the fluid-
Side flow rates will vary depending on the design conditions, the following relationships are based on Dittus-
Boelter turbulent heat transfer associations can be used to predict improved UA based on alternative fluidsside flow rate. UA\' = UA. [{[m\'. sub. fluid]/[m. sub. fluid]}. sup. 0. 8]. [{Pr\'/Pr}. sup. 0. 4](8)
Where: UA = nominal total heat transfer coefficient-area product(Btu/min-[degrees]F)
Ua \'= total heat transfer coefficient corrected-area product(Btu/min-[degrees]F)[P. sub. r]
= Prandtl number used to establish the fluid of nominalUA (-)[P. sub. r]
Prandtl number used to establish the fluid of the modified ua (-)
In addition, the equation (8)
When transitioning from a design load condition to a different working fluid that may occur and producing an over-voltage condition, adapt to the change in the working fluid (e. g.
Changes from fluid drinks under load conditions to CIP solutions during cleaning-up)
This forms the basis for pressure relief protection for heat exchange. The above-
Known information mentioned ([T. sub. refrigerant],[h. sub. vapor,sat], [h. sub. liquid,sat],[T. sub. return][m. sub. fluid], [c. sub. p, fluid], and UA)
Can be used to solve equations simultaneously (4), (6), and (7)
Find the remaining three unknown variables :[m. sub. refrigerant], [T. sub. supply]and Q.
The quantity of interest is the refrigerant steam flow rate ,[m. sub. refrigerant]
, Represents the mass flow of steam generated in the case of over-voltage.
After obtaining, the refrigerant mass flow rate generated must be converted to the equivalent mass flow rate of the air using the following relationship (
ASHRAE month AppendixF): [C. sub. r]= [m. sub. refrigerant]*[C. sub. air]/[C. sub. refrigerant]*[Square root]T. sub. refrigerant]*[M. sub. air]/[T. sub. air]*[M. sub. refrigerant]](9)
Appendix F of ASHRAE 15 (2007)
Assuming a refrigerant temperature of 510 [degrees]R [283 K]
The temperature is 520 [degrees]R [289 K].
Appendix F lists the values of constants ,[C. sub. air]and[C. sub. refrigerant]
For many different refrigerants.
The air mass flow calculated based on the estimated refrigerant steam flow represents the minimum release capability required for the internal load scenario. Plate-and-
Frame technology is rapidly penetrating the refrigeration market for process fluid coolingand-
Frame heat exchanger.
These heat transfer devices can be configured as refrigerant-fed bydirect-
Expansion, gravity flooding recycling, or excess liquid recycling.
Figure 4 shows the on-site installation of the heavy water plate-and-
Frame heat exchanger in industrial cooling application.
Please note that both the liquid supply and the steam returned from the plate heat exchanger have global isolation valves that allow the \"valve out\" heat exchanger from the fluctuation cylinder.
In almost all cases, the surge drum will be ASME-
Usually, the plate heat exchanger will also be ASME-stamped.
In this case, both the fluctuating drum container and the heat exchanger require separate pressure relief protection.
The problem then becomes the problem of determining the pressure relief protection capacity of each device. [
Figure 4 slightly]
ASHRAE 15 [methods outlined]section]9. 7. 5 (
Is the equation (1)in this paper)
Suitable for adjusting the size of the pressure release device of the surge drum.
The next question is: what is the appropriate way to protect the discharge valve size of the plate heat exchanger?
The dimension pressure relief protection of the plate heat exchanger set shall take into account the nature of the heat energy input into the heat exchanger under isolation conditions.
As mentioned earlier, the external heat load due to fire or internal load may result in a heat load of excessive pressure.
By extending the analogy of the external heat load under the fire conditions used on the container, the maximum projected area of the plate assembly can be defined, as shown in figure 5. [
Figure 5 Slightly]
Based on this situation, the capacity of the minimum decompression device required for the plate heat exchanger becomes :[C. sub. r,plateHX]= [Florin]. [Square root]L. sup. 2]+[W. sup. 2]]. H (10)where: [C. sub. r,plate]
Hu x = the minimum discharge capacity required for plate heat exchangers ([lb. sub. m]/min of air)[Florin]
= Capacity coefficient L of discharge device = length of plate package (ft)
W = width of the board (ft)
H = height of the board (ft)
For cases where an \"internal\" heat load from a heat exchanger causes over-voltage, one must evaluate the size of the expected load and convert that load into refrigerant mass flow, this is produced by evaporation of liquid refrigerant in the flat package.
The refrigerant flow rate must then be converted to an equivalent air mass flow rate to maintain consistency with the selection of dimensional processes and decompression devices on other devices.
Due to the internal thermal load of the plate, the subject used to determine the release capacity --
Type heat exchanger and forshell-and-
Tube heat exchanger.
Screw compressors used in industrial refrigeration systems are often equipped with hot siphon oil coolers.
Usually, the heat exchanger is the shell-and-
As shown in Figure 6, a tube heat exchanger that inhibits heat from oil to high pressure refrigerant.
Oil cooling heat exchanger receives return heat from the oil separator of the screw compressor at a temperature close to the discharge temperature of the compressor (~185[degrees]
F for advanced ammonia machine)
And reject its heat in the high-pressure steam flowing through the pipe.
Reduce the oil to the required supply temperature (~130[degrees]F)
Before being sent back to the compressor. [
Figure 6 slightly]
Capacity of refrigerant safety valve-
The side of the thermosiphon oil cooler should be determined based on the maximum rate of refrigerant steam generated inside the oil cooler, which comes from situations where the refrigerant pressure may be too high
Side of the oil cooler.
At the very least, it is recommended to consider the following two situations to protect the oil cooler (refrigerant-side): a.
External heat input: equation (1)for shell-and-
Tube oil cooler type (10)
For plate type oil cooler. b.
Internal heat input: refrigerant steam generated by oil cooling load under normal or elevated oil supply temperature and under designed compressor working conditions.
The first condition is related to the fire condition, in which the refrigerant inside the external heat load evaporation oil cooler causes the overvoltage condition.
Since the oil cooler is often a small heat exchanger (i. e.
Their projection area is small)
, Method-based capacity (a)
It will be low.
The second scene ,(b)
When hot siphon oil-
The Cooledscrew refrigerant compressor package was at the beginning-
Side isolation of oil cooler (valved-out).
In this case, the compressor will operate and exclude heat from the oil cooler, and the supply temperature of the oil cooler will rise over time.
As the compressor continues to operate and a portion of the compressed heat is repelled to the oil cooling heat exchanger through its oil, a point will be reached when turned on
The board compressor safely shuts off the uniton high oil temperature.
A typical screw compressor assembly
About 205 [degrees]F [96[degrees]C].
The saturated pressure corresponding to the refrigerant temperature is equal to the oil at high temperature cuttingout of 205[degrees]F [96[degrees]C]is 825 psig (for ammonia).
Since this pressure is significantly greater than the maximum allowable working pressure of the oil cooling heat exchanger, the oil cooler will be subject to over-voltage in this case.
In the second case ,(b)
, Mass flow rate of refrigerant steam generated on refrigerant-
In exceptional circumstances, the side of the oil cooler is given by the following persons :[m. sub. refrigerant,OC]= [Q. sub. OC]/{[h. sub. vapor,sat]-[h. sub. liquid,sat]}(11)where: [Q. sub. OC]
= Oil cooling heat load generated by the compressor during operation at design suction pressure and discharge pressure, supply oil temperature response at compressor high temperature output limit (Btu/min)[m. sub. refrigerant,OC]
= Mass flow rate of refrigerant evaporation generated by oil cooler ([lb. sub. m]/min)[h. sub. vapor,sat]
= The enthalpy of saturated steam refrigerant under the opening pressure of the fully accumulated decompression device (Btu/[lb. sub. m])[h. sub. liquid,sat]
= The enthalpy of saturated liquid refrigerant under the opening pressure of the fully accumulated decompression device (Btu/[lb. sub. m])
Determine the optimal source oil cooling load for over-voltage conditions ,[Q. sub. OC]
, Is the information provided by the compressor manufacturer.
Computerized selection programs for some compressor manufacturers provide this information based on user input design suction and discharge pressure and oil supply temperature.
After the program is modified (
High oil supply temperature)conditions.
Although there are no convenient rulesof-
Thumb used to determine the cooling load of these AC generator oil, the following is an observation of the general behavior of the oil cooling load at a high supply oil temperature.
As the oil supply temperature increases, the oil cooling load will decrease.
Under the high refrigerant suction pressure, the decrease of oil cooling load is more significant when the oil supply temperature rises.
Table 2 shows examples of oil cooling loads within the suction pressure and supply oil temperature range from the normal range (130[degrees]F)
The degree of temperature reduction in high supply oil-
Shut down the compressor (205[degrees]F)
Typical screw compressor for operation in industrial ammonia refrigeration system.
Under overvoltage conditions, the fraction of the designed oil cooling load is within the range of 74% from a medium low suction pressure (
24 psia or 9 psig)
To reach 49% under higher suction pressure (
49 psia or 34 psig).
Boosterloads showed a similar trend.
Conservatively, the size of the safety valve can be determined according to the oil cooling load designed by the oil cooling heat exchanger (
No load reduction due to high oil return temperature).
Oil cooling load generated under elevated operating conditions ([Q. sub. OC])
The refrigerant mass flow rate can then be estimated by using the equation (11).
Alternatively, the full oil cooling load can be taken to adjust the size of the decompression device.
The refrigerant mass flow rate is then converted to an air base using the equation (9);
This allows the selection of relief devices.
Product storage tanks are sometimes called \"silos\" and are common in food plant applications for industrial refrigeration systems.
The device often uses a jacket heat exchanger to indirectly cool fluid products.
Refrigerant-
The side of the product storage tank is often fed directly-
Expand or submerge;
However, there is also the possibility of excessive liquid supply.
In cases where pressure relief protection is required for this equipment category, capacity determination needs to consider potential sources that may produce refrigerant steam resulting in excessive pressure.
Similar to other heat exchanger types, at least two types of hot addition should be considered, one of which is larger for capacity determination.
The first is the external heat added due to fire conditions.
In this case, the projected area used to release the capacity determination of the unit is only for the heat exchanger--
Not the whole product tank.
Figure 7 shows the vertical product storage tank with jacket heat exchanger located near the bottom of the tank. [
Figure 7 Slightly]
The discharge capacity determination based on the external heat load is given by the following personnel :[C. sub. r,tank]= [Florin]*D*H (12)where: [C. sub. r,tank]
= Minimum required emission capacity of relief device for protection product tank heat exchanger indicated [lb. sub. m]/min of air [Florin]
= Capacity coefficient D = outer diameter of product tank (ft)
H = height of heat exchanger (ft)
In almost all cases, the heat exchanger is shielded by the outer sheath of the product tank, which makes it difficult to accurately determine its height.
Therefore, the height of the heat exchanger should be used as the distance between the two refrigerant connections on the tank (
Supply and return).
The second case of refrigerant steam generation in the heat exchanger is generated during cleaningin-place (CIP).
Ideally, the designer should work with the silo or tank manufacturer to obtain detailed information about the CIP cycle to determine the required decompression capability.
In the absence of this information, the rate at which refrigerant steam is produced during CIP can be estimated as follows :[m. sub.
Refrigerant, tank]=[epsilon]. [m. sub. fluid,CIP]. [cp. sub. fluid,CIP]. ([T. sub.
Fluid, CIP, supply]-[T. sub. ref,sat])/([h. sub. vapor,sat]-[h. sub. liquid,sat])(13)where: [m. sub.
Tank Ray Grant
= Mass flow rate of refrigerant evaporation generated in heat exchanger ([lb. sub. m]/min)[epsilon]
= Refrigerant performance on product tanks (estimated as0. 2)[m. sub. fluid,CIP]
= CIP fluid mass flow rate ([lb. sub. m]/min)[CP. sub.
Fluid, CIP = CIP fluid thermal capacity (
Approximate 1Btu /[lb. sub. m]-[degrees]F)[T. sub.
Fluid, CIP, supply]
= Maximum liquid supply temperature during Cip ([degrees]F)[T. sub. ref,sat]
= Therelief valve sets the saturation temperature of refrigerant under pressure ([degrees]F)[h. sub. vapor,sat]
= The enthalpy of saturated steam refrigerant at the set pressure of the fully accumulated decompression device (Btu/[lb. sub. m])[h. sub. liquid,sat]
= Enthalpy of saturated liquid refrigerant at the set pressure of the fully accumulated release device (Btu/[lb. sub. m])
After determining the mass flow of refrigerant, the capacity of the decompression device (on an air-
Equivalent base)
By using the equation (9).
One of the larger of these two capabilities is the basis for the selection of product tank decompression devices.
Conclusion The safety decompression system is an important part of maintaining the safety of the refrigeration system.
Standards such as ASHRAE standard 15 include the prescribed method of adjusting the discharge valve size for the container and positive displacement compressor;
However, it does not include coverage of other related equipment that may have a decompression device installed.
In this paper, we propose an oil separator, shell-and-
Tube heat changers, boardand-
Frame heat exchanger, oil cooler and jacketedproduct storage tank.
The proposed dimension measurement method takes into account the overvoltage source both outside the system and inside the system.
It is important to evaluate these two sources in order to be able to apply a relief device of the right size.
The pressure relief device with insufficient size causes the steam flow to be limited;
Therefore, due to the higher build, the risk of failure of components during operation is increased
Increased internal pressure.
Excessive Pressure Relief Devices will lead to a rapid cycle of the operation of the pressure relief device (chatter)
Relief equipment may fail if allowed to last.
NOMENCLATURE [C. sub. p,fluid]
= Thermal capacity of secondary fluid (Btu/[lb. sub. m]-[degrees]F)[CP. sub.
Liquid, CIP = cleaning-in-
Place the thermal capacity of the fluid (Btu/[lb. sub. m]-[degrees]F)[C. sub. r]
= Minimum required discharge capacity of the relief device of the container ([lb. sub. m]/min of air)[C. sub. r,plateHX]
= Minimum discharge capacity required for plate heat exchangers ([lb. sub. m]/ min of air)[C. sub. r,OS]
= Minimum required discharge capacity for reliable devices that protect oil separators ([lb. sub. m]/min of air)[C. sub. r,tank]
= Minimum required emission capacity to protect the relief device of the product tank heat exchanger ([lb. sub. m]/min of air)
D = outer diameter of container or product tank (ft)[D. sub. s]
= Outer diameter of surge drum (ft)[D. sub. v]
= Outer diameter of the Hell main container part-and-
Tube heat exchanger (ft)[Florin]
= The capacity factor of the decompression device depending on the type of refrigerant and whether the combustible material is close to the pressure vessel (
See ASHRAE 15 for capacity factor values)
H = the height of the plate set or tank heat exchanger (ft)[h. sub. vapor,sat]
= The enthalpy of saturated steam refrigerant at the set pressure of the fully accumulated decompression device (Btu/[lb. sub. m])[h. sub. liquid,sat]
= Enthalpy of saturated liquid refrigerant at the set pressure of the fully accumulated release device (Btu/[lb. sub. m])
L = the length of the container or board assembly (ft)
LMTD = log average temperature difference ([degrees]F)[L. sub. s]
= Length of surge drum (ft)[L. sub. v]
= The length of the main container part of the shell-and-
Heat exchanger (ft)[m. sub. fluid,CIP]
= Mass flow of secondary fluid ([lb. sub. m]/min)[m. sub. fluid,CIP]= clean-in-
Placement of fluid mass flow ([lb. sub. m]/min)[m. sub. regrigerant,OC]
= Refrigerant steam generation rate ([lb. sub. m]/min)[m. sub. refrigerant,OC]
= Mass flow rate of refrigerant evaporation generated by oil cooler ([lb. sub. m]/min)[m. sub.
Refrigerant, tank]
= Mass flow rate of refrigerant evaporation generated in the tank heat exchanger ([lb. sub. m]/min)[P. sub. r]
= Prandtl number used to establish the fluid of nominalUA (-)[P. sub. r]
Prandtl number used to establish the fluid of the modified ua (-)
Q = heat flow of heat exchanger (Btu/min)[Q. sub. OC]
= Heat load of oil cooling heat exchanger (Btu/min)[T. sub.
Fluid, CIP, supply]
= Maximum liquid supply temperature during Cip ([degrees]F)[T. sub. refrigerant]
= Refrigerant saturation temperature ([degrees]F)[T. sub. ref,sat]
= Refrigerant saturation temperature at the set pressure of the safety valve ([degrees]F)[T. sub. return]= load-
Secondary fluid return temperature ([degrees]F)[T. sub. supply]= load-
Secondary fluid supply temperature ([degrees]F)
UA = total heat transfer coefficient-area product(Btu/min-[degrees]F)
Ua \'= total heat transfer coefficient corrected-area product(Btu/min-[degrees]F)
W = width of the board (ft)[epsilon]= refrigerant-
Can effect on products (estimated as0.
2 Bulk cans)
Reference to ASME, \"2007 ASME Boiler and Pressure Vessel Specification\", American Association of Mechanical Engineers, New York ,(2007).
ASHRAE standard 15, \"safety standards for refrigeration systems\", American Association of Heating, Refrigeration and Air Conditioning Engineers, Atlanta, GA ,(2007). Fenton, D. L.
And Richards, W. V.
, User manual forANSI/ASHRAE Standard 15-
2001, American Association of Heating, Refrigeration and Air Conditioning Engineers, Atlanta, GA ,(2003).
Douglas T. Reinder
And Todd B. Jackel.
Guidelines for engineering safety relief systems.
Industrial Refrigeration Association
University of Wisconsin
Madison, Madison, WI (2006). Reindl, D. T. and Jekel, T. B.
Foundation of security relief system, ASHRAE magazine 50 (2):22-29 (2008). Douglas T.
PE dl, PhD, sports researcher ASHRAE Todd B.
Dr. Jekel, MD, mp ashrae Douglas
Reindl is professor and director, Todd B.
Jackkel is an assistant director of the University of Wisconsin-Madison Federation of Industrial refrigeration.
Custom message
Chat Online
Chat Online
Chat Online inputting...