In this tip of the month (TOTM) we will describe the dense phase of a pure compound, what it is, and how it impacts processes. We will illustrate how thermophysical properties change in the dense phase as well as in the neighboring phases. The application of dense phase in the oil and gas industry will be discussed briefly. In next month TOTM, we will discuss the dense phase behavior of multi-component systems.
When a pure compound, in gaseous or liquid state, is heated and compressed above the critical temperature and pressure, it becomes a dense, highly compressible fluid that demonstrates properties of both liquid and gas. For a pure compound, above critical pressure and critical temperature, the system is oftentimes referred to as a “dense fluid” or “super critical fluid” to distinguish it from normal vapor and liquid (see Figure 1 for carbon dioxide). Dense phase is a fourth (Solid, Liquid, Gas, Dense) phase that cannot be described by the senses. The word “fluid” refers to anything that will flow and applies equally well to gas and liquid. Pure compounds in the dense phase or supercritical fluid state normally have better dissolving ability than do the same substances in the liquid state. The dense phase has a viscosity similar to that of a gas, but a density closer to that of a liquid. Because of its unique properties, dense phase has become attractive for transportation of natural gas, enhanced oil recovery, food processing and pharmaceutical processing products.
The low viscosity of dense phase, super critical carbon dioxide (compared with familiar liquid solvents), makes it attractive for enhanced oil recovery (EOR) since it can penetrate through porous media (reservoir formation). As carbon dioxide dissolves in oil, it reduces viscosity and oil-water interfacial tension, swells the oil and can provide highly efficient displacement if miscibility is achieved. Additionally, substances disperse throughout the dense phase rapidly, due to high diffusion coefficients. Carbon dioxide is of particular interest in dense-fluid technology because it is inexpensive, non-flammable, non-toxic, and odorless. Pipelines have been built to transport natural gas in the dense phase region due to its higher density, and this also provides the added benefit of no liquids formation in the pipeline.
In the following section we will illustrate the variation of thermophysical properties in the dense phase and its neighboring phases. Methane properties have been calculated with HYSYS software for a series of temperatures and pressures. Table 1 presents, the pressures and temperatures and their paths used in this study.
The calculated thermophysical properties are plotted as a function of pressure and temperature in Figures 2 to 9. The thermophysical property is shown on the left-hand side y-axis, temperature on the x-axis and pressure on the right-hand side y-axis.
Table 1. Pressure-Temperature combination and the paths chosen for methane
Density:
Figure 2 presents the variation of density in different phases as a function of pressure and temperature. In the isobaric subcooling path of AB, liquid density increases gradually. However, in the isothermal compression of BC path, a small increase of density is observed. In the isobaric CD path, compressed liquid density decreases gradually as temperature is increased well into the dense phase region. However, as the temperature increases further in the dense phase, density reduction is accelerated. Reduction of density is further accelerated during isothermal expansion of DE. Isobaric cooling of vapor along EF path corresponds with a gradual increase in density. It can be noted the values of dense phase density are close to the liquid phase density in some areas of the dense phase region, and is overall significantly higher than the vapor phase densities.
Viscosity:
Figure 3 presents the variation of viscosity in different phases as a function of pressure and temperature. In the isobaric subcooling path of AB, liquid viscosity increases rapidly. However, in the isothermal compression of BC path, a very small change of viscosity is observed. In the isobaric CD path, compressed liquid viscosity decreases linearly and sharply as temperature is increased well into the dense phase region. As the temperature increases further in the dense phase, viscosity reduction becomes gradual and approaches the gas phase values. Reduction of viscosity is quite small during isothermal expansion of DE. Isobaric cooling of vapor along EF path corresponds with no appreciable change in viscosity.
Compressibility Factor:
In general, the compressibility factor Z, calculated by an equation of state is not accurate for the liquid phase. Therefore, Figure 4 which presents compressibility factor as a function of pressure and temperature should be considered for qualitative study only. In the isobaric subcooling path of AB, Z remains almost constant. However, in the isothermal compression of BC path, Z increases drastically. In the isobaric CD path, Z increases gradually as temperature is increased well into the dense phase region. As the temperature increases further in the dense phase, the increase in Z is accelerated. The increase in Z is further accelerated during isothermal expansion of DE. Isobaric cooling of vapor along FF path corresponds with a gradual decrease in Z.
Surface Tension:
Figure 5 shows that in the liquid phase, surface tension is a strong function of temperature but independent of pressure. Above the critical temperature, surface tension is not applicable and its value is zero.
Heat Capacity:
Generally, heat capacity is applicable in a single phase region and should not be used when there is a phase change. Figure 6 presents the variation of density in different phases as a function of pressure and temperature. In the isobaric subcooling path of AB, liquid heat capacity decreases. In the isothermal compression of BC path, a small decrease of heat capacity is observed. In the isobaric CD path, compressed liquid heat capacity increases gradually as temperature is increased well into the dense phase region. As the temperature increases further in the dense phase, heat capacity reaches a maximum value and then starts to decrease. This is strange behavior and surprisingly high values are calculated. Similar results were obtained using ProMax software. Reduction of heat capacity is further noticed during isothermal expansion of DE. Isobaric cooling of vapor along EF path corresponds with a gradual increase in heat capacity.
Thermal Conductivity:
Figure 7 presents the variation of thermal conductivity in different phases as a function of pressure and temperature. In the isobaric subcooling path of AB, liquid thermal conductivity increases. In the isothermal compression of BC path, no change is observed. In the isobaric CD path, compressed liquid thermal conductivity decreases gradually as temperature is increased well into the dense phase region. However, as the temperature increases further in the dense phase, thermal conductivity reduction is accelerated. Reduction of thermal conductivity is further noticed during isothermal expansion of DE. Isobaric cooling of vapor along EF path corresponds with a small decrease in thermal conductivity.
Enthalpy and Entropy:
Figures 8 and 9 present the variation of enthalpy and entropy in different phases as a function of pressure and temperature. As shown in these figures, their qualitative variations are similar. In the isobaric subcooling path of AB, liquid enthalpy and entropy decrease. In the isothermal compression of BC path, no change is observed. During the isobaric CD path, compressed liquid enthalpy and entropy values increase gradually as temperature is increased well into the dense phase region. However, as the temperature increases further in the dense phase, the enthalpy and entropy increase becomes larger. The increase in enthalpy and entropy is further noticed during isothermal expansion of DE. Isobaric cooling of vapor along EF path corresponds with a decrease in enthalpy and entropy.
Conclusions:
Dense phase behavior is unique and has special features. The thermophysical properties in this phase may vary abnormally. Care should be taken when equations of state are used to predict thermophysical properties in dense phase. Evaluation of equations of state should be performed in advance to assure their accuracy in this region. Many simulators offer the option to use liquid-based algorithms (e.g. COSTALD) for this region.
As shown in Figure 1, there is a gradual change of phase transition from gas-to-dense and dense-to-liquid phases or vice versa. Dense phase is a highly compressible fluid that demonstrates properties of both liquid and gas. The dense phase has a viscosity similar to that of a gas, but a density closer to that of a liquid. This is a favorable condition for transporting natural gas in dense phase as well as carbon dioxide injection into crude oil reservoir for enhanced oil recovery.
John M. Campbell Consulting (JMCC) offers consulting expertise on this subject and many others. For more information about the services JMCC provides, visit our website at www.jmcampbellconsulting.com, or email your consulting needs to consulting@jmcampbell.com.
In the August 2009 Tip of the Month (TOTM), it was shown that pumping power requirement varies as the crude oil °API changes. Increasing °API or line average temperature reduces the crude oil viscosity. The viscosity reduction caused higher Reynolds number, lower friction factor and in effect lowered pumping power requirements. Since the objective of the August 2009 TOTM was to study the effect °API and the line average temperature have on the pumping power requirement, the effect of crude oil viscosity on pump performance was ignored and in the course of calculation a constant pump efficiency of =0.75 was used for all cases. In this TOTM, we will consider the crude oil viscosity effect on a selected pump performance. The Hydraulic Institute Standards [1] procedures and the guideline presented in the August 2006 TOTM written by Honeywell were applied to correct the pump efficiency.
As in the August 2009 TOTM, we will study crude oil °API and the pipeline average temperature and how these effect the pumping requirement. For a case study, we will consider a 160.9 km (100 miles) pipeline with an outside diameter of 406.4 mm (16 in) carrying crude oil with a flow rate of 0.313 m3/s or 1,126 m3/h (170,000 bbl/day or 4958 GPM). The pipeline design pressure is 8.963 MPa (1300 psia) with a maximum operating pressure of 8.067 MPa (1170 psia). The wall thickness was estimated to be 6.12 mm (0.24 in). The wall roughness is 51 microns (0.002 in) or a relative roughness (e/D) of 0.00013. The procedures outlined in the March 2009 TOTM were used to calculate the line pressure drop due to friction. Then corrected pumping efficiency was used to calculate the required pumping power. Since the objective was to study the effect °API and the line average temperature have on the pumping power requirement, we will ignore elevation change. The change in pumping power requirements due to changes in crude oil °API and line average temperature for this case study will be demonstrated.
Viscosity Effect on Centrifugal Pump Performance
There are several papers investigating and presenting procedures for correcting centrifugal pump curves [2-3]. According to Turzo et al. [2], three models are available for correcting performance curves: Hydraulic Institute, Stepanoff, and Paciga. Turzo et al. [2] also presented a computer applications for correcting pump curves for viscosity effect. In this review, the Hydraulic Institute [1], HI, procedure was applied and is described briefly here.
HI uses a performance factor, called Parameter B which includes terms for viscosity, speed, flow rate and total head. The method uses a new basis for determining the correction factors CH, CQ, and C. The basic equation for Parameter B is given as Equation 1.
B = Performance factor K = 16.5 for SI units
= 26.5 for USCS (FPS) vis = Viscous fluid Kinematic viscosity – cSt HBEP-W = Water head per stage at BEP – m (ft) QBEP-W = Water flow rate at BEP – m3/h (gpm) N = Pump shaft speed – rpm
Correction factors are applied to capacity (CQ), head (CH), and efficiency (C). Calculation of these Correction Factors is dependent on the calculated value of Parameter B. For the cases considered in this study, the B values were less than 1; therefore, based on the HI guideline, the correction factors for head and capacity were set equal to 1 and the correction factor for efficiency, C, was calculated by Equation 2.
BEP-W = Pump efficiency at BEP Vw = Water kinematic viscosity – cSt
Figures 1 and 2 present the water-based pump curves used in this study. For computer calculations, these two curves were fitted to polynomials of degrees 3 and 2 for head vs capacity and efficiency vs capacity, respectively.
In Equations 3 and 4, H is in m (ft) and Q is in m3/h (GPM). For this pump:
Case Study 1: Effect of Line Average Temperature (Seasonal Variation)
To study the effect of the line average temperature on the pumping power requirement, an in house computer program called OP&P (Oil Production and Processing) was used to perform the calculations outlined in the March 2009 TOTM. For a 35 °API crude oil in the pipeline described the required pumping power was calculated for line average temperature ranging 21.1 to 37.8 °C (70 to 100 °F). For each case, the parameter B was calculated by Equation 1 and since its value was less than 1, the efficiency correction factor was calculated by Equation 2. Then, the pump efficiency calculated by Equation 4 was multiplied by the correction factor for the subsequent calculations. The corrected efficiency ranged from 0.70 to 0.72. The required pumping power was compared with an arbitrary base case (85 °F or 29.4 °C and constant = 0.75) and the percentage change in the pumping power requirement was calculated. Figure 3 presents the percent change in power requirement as a function of line average temperature. There is about 5% change (for constant =0.75) and more than 8% change (for corrected efficiency) in the pumping power requirement for the temperature range considered.
Note that as the line average temperature increases the power requirement decreases. This can be explained by referring to Figure 4 in which the oil viscosity decreases as the temperature increases. Lower viscosity results in higher Reynolds (i.e. Reynolds number which is the ratio of inertia force to viscous force); therefore, the friction factor decreases (refer to the Moody friction factor diagram in the March 2009 TOTM).
Case Study 2: Effect of Variation of Crude Oil °API
In this case, the effect of crude oil °API on the total pump power requirement for three different line average temperatures was studied. For each line average temperature, the crude oil °API was varied from 30 to 40 and the total pumping power requirement was calculated and compared to the base case (35 °API and average line temperature of 29.4°C=85°F).
For each case the percent change in total power requirement was calculated and is presented in Figure 5. As shown, when °API increases the total power requirement decreases. This also can be explained by referring to Figure 4 in which the crude oil viscosity decreases as ° API increases. The effect of viscosity is more pronounced at lower line average temperature (i.e. 21.1 °C or 70°F). Figure 5 also indicates that there is about 30 % change in total power requirement as °API varies from 30 to 40 °API. This is a significant variation and suggests that it should be considered during design of crude oil pipelines.
Discussion and Conclusions
The analysis of Figures 3 and 5 indicates that for the oil pipeline, the pumping power requirement varies as the crude oil °API changes. Increasing °API or line average temperature reduces the crude oil viscosity (see Figure 4). The reduction of viscosity results in higher a Reynolds number, lower friction factor and in effect lowers pumping power requirements.
For the cases studied in this TOTM, the effect of crude oil viscosity on the performance of pump was considered. It was found that no correction was required for the capacity and head but a correction factor in the range of 0.95 to 0.98 was required to adjust the pump efficiency for crude oil applications.
A sound pipeline design should consider expected variations in crude oil °API and the line average temperature. In addition, the pump performance curves should be corrected for the effect of viscosity.
During the life cycle of a crude oil pipeline the properties of transported oil change, because in gathering systems the produced oils come from different wells. New wells may be added or some wells may go out of production for maintenance and repair. Production rates during the life of wells vary, too. In addition the properties of crude oil change during production. Due to seasonal variation, the average line temperature may also change. As it is shown in the proceeding sections, viscosity of crude oil is a strong function of API gravity and temperature.
In the March 2009 tip of the month (TOTM), procedures for calculation of friction losses in oil and gas pipelines were presented. The sensitivity of friction pressure drop with the wall roughness factor was also demonstrated.
In this TOTM, we will study crude oil °API and the pipeline average temperature and how they effect the pumping requirement. For a case study, we will consider a 160.9 km (100 miles) pipeline with an outside diameter of 406.4 mm (16 in) carrying crude oil with a flow rate of 0.313 m3/s (170,000 bbl/day). The pipeline design pressure is 8.963 MPa (1300 psia) with a maximum operating pressure of 8.067 MPa (1170 psia). The wall thickness was estimated to be 6.12 mm (0.24 in). The wall roughness is 51 microns (0.002 in) or a relative roughness (e/D) of 0.00013. The procedures outlined in the March 2009 TOTM were used to calculate the line pressure drop due to friction. Then assuming 75 % pumping efficiency, the required pumping power was calculated. Since the objective was to study the effect °API and the line average temperature have on the pumping power requirement, we will ignore elevation change. The change in pumping power requirements due to changes in crude oil °API and line average temperature for this case study will be demonstrated.
Case Study 1: Effect of Line Average Temperature (Seasonal Variation)
To study the effect of the line average temperature on the pumping power requirement, an in house computer program called OP&P (Oil Production and Processing) was used to perform the calculations as outlined in the March 2009 TOTM. For a 35 °API crude oil in the pipeline described in the preceding section, the required pumping power was calculated for the line average temperature ranging from 21.1 to 37.8 °C (70 to 100 °F). For each case, the required pumping power was compared with an arbitrary base case (85 °F or 29.4 °C) and the percentage change in the pumping power requirement was calculated, accordingly. Figure 1 presents the percent change in power requirement as a function of line average temperature. There is about 5% change in the pumping power requirement for the temperature range considered.
Note as the line average temperature increases, the power requirement decreases. This can be explained by referring to Figure 2 in which the oil viscosity decreases as the temperature increases. Lower viscosity results in higher Reynolds (i.e. Reynolds number is the ratio of inertia force to viscous force); therefore, the friction factor decreases (refer to the Moody friction factor diagram in the March 2009 TOTM).
Case Study 2: Effect of Variation of Crude Oil API
In this case, the effect of crude oil °API on the total pump power requirement for three different line average temperatures was studied. For each line average temperature, the crude oil °API was varied from 30 to 40 and the total pumping power requirement was calculated and compared to the base case (35 °API and average line temperature of 29.4°C=85°F).
For each case the percent change in total power requirement was calculated and is presented in Figure 3. As shown in this figure, when °API increases the total power requirement decreases. This also can be explained by referring to Figure 2 in which the crude oil viscosity decreases as ° API increases. The effect of viscosity is more pronounced at lower line average temperature (i.e. 21.1 °C or 70°F). Figure 3 also indicates that there is about 25 % change in total power requirement as °API varies from 30 to 40 °API. This is a big change and should be considered during design of crude oil pipelines.
Discussion and Conclusions
The analysis of Figure1-3 indicates that for the oil pipeline, the pumping power requirement varies as the crude oil °API changes. Increasing °API or line average temperature reduces the crude oil viscosity (see Figure 2). The reduction of viscosity results in higher Reynolds number, lower friction factor and in effect lower pumping power requirements.
In practical situations, an originating station takes crude out of storage and the midline stations taking suction from the upstream section of pipeline. In some parts of the world, the suction temperature to the originating pumps is +38 °C (+100 °F) but the temperature to the midline station is ground temperature (this assumes a buried line below the frost line) approximately 18 °C (65 °F). The originating station will always be more affected by temperature because storage will follow ambient – whereas the midline station will operate at notionally constant temperature +/- 5.5 °C (+/- 10 °F) in the lower 9 °C (48 °F). For the case studied in this TOTM, the number of pumping stations varied from 2.5 to 3.2.
In light of the above discussion, a sound pipeline design should consider expected variation in crude oil °API and the line average temperature.
To many people involved in the Oil and Gas production and refining industry, the terms monitoring and inspection are used interchangeably when referring to corrosion issues. However, this lack of differentiation can lead to misunderstandings and errors. It is our contention that a clear differentiation is needed and that engineers should strive to use the correct terminology. In order to achieve that differentiation it is necessary to first define these terms ‘corrosion monitoring’ and ‘inspection’. A review of some of the named techniques and methods used in these areas will help to consolidate an understanding of which terms fall into the inspection bracket and which are viewed as corrosion monitoring devices.
Definitions
The following definitions may not be exactly scientific in nature, but they do help to show two major differences between the two sets of valuable corrosion management tools.
‘Corrosion monitoring’ – is a way of determining how corrosive the fluids are within a specific environment. The various techniques available are typically used to give frequent, short time interval measurements, thereby allowing the day-to-day control of corrosion mitigation / prevention approaches such as corrosion inhibition. (The one exception to the ‘short time interval’ description is the weight-loss coupon).
‘Inspection’ – is the means by which corrosion (and other) damage may be located in a structure, as well as gaining insight to the amount and severity of that damage. Usually inspection tools are used less frequently than corrosion monitoring devices, often on an annual or even longer basis. However, the frequency of measurement should be determined via a process of risk based analysis to give a programme of ‘risk based inspection’ (RBI).
The Methods and Techniques
It is important to point out at the outset that the use of any corrosion monitoring device or inspection tool should be within the bounds of prudent process safety engineering. In the first place only trained personnel should be allowed to operate and maintain the various pieces of equipment. Secondly, they should learn about the system to be monitored / inspected, so that they clearly understand what risks are involved with respect to carrying out their monitoring / inspection activities.
Corrosion Monitoring Techniques
The most commonly used corrosion monitoring devices are included in the following list of equipment:
– Weight-loss coupons
– Spool pieces
– Electrical resistance probes
– Linear polarisation probes
– Galvanic probes
– Hydrogen pressure probes
– Hydrogen electrochemical patch probes
– Electrochemical noise probes
– Field Signature MethodTM
– Bioprobes
The majority of these techniques are classed as ‘intrusive’, in that for internal measurement there must be an access fitting to allow the measuring probe to be inserted into the process fluids. The exceptions on this list are the hydrogen electrochemical patch probes and the Field Signature MethodTM, which are attached to the outside surface of vessels and pipes.
Weight loss coupons, spool pieces and bioprobes, give one-off readings. In order to determine the result, each of these items must be withdrawn from the system and carefully examined and tested. The other devices can be left in place for some time, measurements being obtained either manually or automatically collected via hard wire connections or radio transmission devices.
Inspection Tools
The inspection tools are either arranged on the external surfaces of structures, or inserted into tubing via wireline and into pipelines installed in ‘intelligent pigs’. The following list of these methods is not exclusive:
– Visual inspection
– Eye and magnifying glass
– Boroscopes
– Fiberscopes
– Robotic crawlers
– Cameras
– Calliper tools (on wireline or in intelligent pigs)
– Ultrasonic thickness (UT) measurements
– ‘Spot’ UT (compression mode) / straight beam (UTL)
– Pulse-echo contact method
– Shearwave mode (UTS)
– Phased array
– Automated UT (both in compression and shear modes)
– Long range UT (LRUT, or guided wave inspection – GWI)
– Radiography (RT)
– Dye penetrant (PT)
– Magnetic flux leakage (for example in intelligent pigs)
These first six methods are the most commonly used. Others include the following:
– Dry magnetic particle
– Wet magnetic particle
– Wet fluorescent magnetic particle testing (WFMP or WFMT)
– Magnetostrictive guided wave testing (MGWT)
– Eddy current
– Pulsed eddy current (PEC)
– Neutron backscatter (for CUI – corrosion under insulation)
– Tangential radioscopy
– Magnetic flux exclusion
– Acoustic emission (AE)
– Acousto ultrasonics
Specialist application of some of these inspection tools allows the detection of cracking damage, including sub-surface cracks. Early detection of the latter can obviously prevent subsequent catastrophic failures.
Finally
This ‘Tip of the Month’ has concentrated on the need to differentiate between corrosion monitoring and inspection, to show that each play a part in the overall corrosion management of an oil and gas production / processing system. For more information on these various methods there are a number of sources, including the JMC/PetroSkills training class:
“PF-22; Corrosion Management in Production / Processing Operations”.
By: Alan Foster Discipline Manager for ‘Water and Corrosion’
One of the most important physical properties of a gas is the ratio of specific heats. It is used in the design and evaluation of many processes. For compressors, it is used in the design of components and determination of the overall performance of the machine. Engineers are frequently asked to evaluate a compressor performance utilizing traditional equations of head, power and discharge temperature. While these simplified equations may not give exact results, they give useful information needed to troubleshoot a machine, predict operating conditions, or a long-term trend analysis. The accuracy of the performance information will depend on the proper selection of the ratio of specific heats. This Tip of the Month (TOTM) will investigate the application of the ratio of specific heats to compressors, its sensitivity to the determination of machine performance and give recommendations for improved accuracy.
Background of k-value
The ratio of specific heats is a physical property of pure gases and gas mixtures and is known by many other names including: adiabatic exponent, isentropic exponent, and k-value. It is used to define basic gas processes including adiabatic and polytropic compression. It also appears in many of the traditional equations commonly used to determine a compressor head, gas discharge temperature, gas power, and polytropic exponent. The k-value also influences the operating speed of a compressor, but we will simplify the present analysis by deleting speed from our evaluation. The following commonly used compressor performance equations show how the k-value is utilized in the design and evaluation of compressors.
Note: The actual Z-value will vary from the suction to discharge conditions. ZS is sometimes replaced with ZAVE to approximate the variations in compressibility value [1, 5]. See the nomenclature at the end of this TOTM.
The above equations are written in terms of the adiabatic process with the exception of Equation 5, which refers to the polytropic process. Both compression processes are similar and will give the same actual results. The adiabatic and polytropic methods are extensively used by manufacturers to design compressors, and make use of k-values to calculate their performance. However, as will be seen, the effect of the k-value and the calculated results will influence both compression processes alike. For simplicity, this Tip of the Month will use the adiabatic process.
It can be seen from Equations 1-5 that the k-value has an effect on a compressor head, temperature, power, and polytropic exponent. In order to determine how small changes in the k-value can influence a compressor performance, let us first define the k-value of a pure gas. The thermodynamic definition of a gas k-value is given by Equation 6. It shows the relationship to the specific heat at constant volume, CV and specific heat at constant pressure, CP. Both values vary with temperature and pressure.
For a pure gas there are many references that give CP and CV values at various conditions. One useful source is National Institute of Standards and Technology. Their website is http://webbook.nist.gov/chemistry/fluid/
The method of determining the k-value for gas mixtures is more complex. The major difference is that a gas mixture does not behave as any one of its components but as an “equivalent” gas. Therefore, to determine the k-value of the mixture, we must know the mole fraction of each component, Yi and the molar specific heat at constant pressure for each component, M CPi. Equation 7 can be used to determine the k-value of an ideal gas mixture [1, 5]. Real gases may deviate from the calculated value.
While Equations 1-7 are applicable for manual calculations methods, it is important to note that process simulation packages determine the compressor head and discharge temperature utilizing equations of state. The results are the same but the methods are very different.
K-value Sensitivity Analysis
In the compression process the temperature and pressure of the process gas both increase. Not knowing what k-value to select for evaluating the compression process can lead to errors. For example, a typical propane compressor may have a k-value at suction conditions of 1.195. At the compressor discharge conditions the k-value is 1.254. The difference in the two values varies by 4.94 percent and can have a significant influence in the performance evaluation. The following example illustrates how minor changes in the k-value can influence the calculated compressor head, temperature, power and the polytropic coefficient.
Example 1: A natural gas compressor is operating at the conditions given below. Only the k-value is varied from 1.20 to 1.28, all other given parameters remain constant. Figure 1 illustrates how the “apparent” performance of a compressor can change by varying the k-value.
It can be seen from Figure 1 that the discharge temperature deviated over 18.8 percent by only changing the k-value by 6.7 percent. In this case the k-value varied from a value of 1.20 to 1.28; which is the typical range for natural gas. Similarly, the power changed by 2.5 percent, polytropic exponent by 9.5 percent, and adiabatic head by 2.5 percent for the same variation of the k-value. The changes in compressor performance described in Figure 1 can be much larger depending on the gas composition and the operating temperature and pressure.
Corrected k-Value Recommendations
The k-value sensitivity for a single-stage machine is not nearly the problem as a multi-stage compressor. For a single-stage machine, the pressure ratio is typically lower and the temperature and pressure changes are less. As a result the changes in k-value are not as great and accurate results can be obtained by approximating the k-value at the suction conditions. However, for multi-stage machines, where the pressure and temperature ratios are higher, the k-value sensitivity is more of a factor in evaluating compressor performance. Most compressor manufacturers calculate the k-value for each stage of compression and avoid errors introduced by utilizing an overall k-value. Without their software, we are left with a corrected k-value by empirical methods.
There are many useful approximations that will correct for changes in the k-value as the process gas passes through the compressor. Normally the k-value will decrease during compression but not always. Utilizing the suction conditions to estimate the k-value will generally give higher values of temperature, heat, and power. The polytropic exponent generally decreases as the adiabatic exponent decreases. To avoid potential discrepancies, a k-value correct may be warranted. The following are six methods of determining the corrected k-value commonly used in industry.
At TS and PS: This method determines the k-value at suction conditions and is useful for single stage compressors or applications where there is little change in the k-value. The k-value is easy to determine and tends to overestimate results, especially if the temperature and pressure do not change significantly. For greater values of RP the results may become so conservative they become useless.k = ks at suction conditions
At TD and PD: This method determines the k-value at discharge conditions. The k-value is less conservative and tends to underestimate results. The k-value may be difficult to determine, especially if the discharge temperature is unknown. For gases with highly variable k-values, an iterative solution may be required to estimate the discharge temperature and corrected k-value.k = kD at discharge conditions
At TAVE and PSTD [5]: This method utilizes the average operating temperature at standard pressure and determines the k-value. Numerous reference books propose this method. Errors are introduced because the k-value at standard pressure may not accurately represent values at the operating pressure.k = at average operating temperature and standard pressure
At TAVE and PAVE: This method utilizes the k-value at the average operating temperature and pressure.k = at average operating temperature and pressure
Average value [1, 3]: This empirical method takes the average k-value at compressor inlet conditions and outlet conditions. Utilizing the average k-value will result in performance values that are closer to the actual performance of the compressor.
Weighted average value [4]: This empirical method takes the weighted average of the suction, mid-point and discharge conditions. Note that the mid-pressure is determined by equivalent pressure ratios, . The mid-temperature is estimated from the mid-pressure. This method considers the staged k-value to change with diverging isentropic and pressure lines shown on a Mollier chart.
Example 2 illustrates the various methods used to determine corrected k-values given above. It also compares the range of the resulting values.
Example 2: A propane compressor is operating at the given conditions shown below. Table 1 lists the k-values attributed to various operating and reference conditions [6].
Summary
This Tip of the Month has defined the physical property of process gases called the k-value or ratio of specific heats. It has shown that small changes in the k-value can have a significant effect on the calculated values of head, power, gas discharge temperature, and polytropic exponent. Recommendations were also given to improve the accuracy by utilizing different k-value methods.
In the February 2007 tip of the month (TOTM), Joe Honeywell [1] presented a procedure for calculating fluid pressure drop for liquid in a piping system due to friction. Continuing Honeywell’s TOTM, we will outline procedures for calculation of friction losses in oil and gas pipelines. From an engineer’s point of view the question may arise “how sensitive is friction pressure drop with the wall roughness factor?” Of course the answer is “it depends”. To explain this answer quantitatively and qualitatively, we will study the effect of wall roughness factor for two case studies in this month’s TOTM. In the first case study, an oil pipeline with a flow rate of 0.313 m3/s (170,000 bbl/day) and in the second case, a natural gas pipeline with a flow rate of 22.913 Sm3/s (70 MMSCFD) will be studied and calculation results will be presented in tabular and graphical format.
Friction Factor
The Moody diagram in Figure 1 is a classical representation of the fluid behavior of Newtonian fluids and is used throughout industry to predict fluid flow losses. It graphically represents the various factors used to determine the friction factor. For example, for fluids with a Reynolds number of 2000 and less, the flow behavior is considered a stable laminar fluid and the friction factor is only dependent on the Reynolds number [2]. The friction factor, f, for the Laminar zone is represented by:
Where Re is the Reynolds number and is expressed as the ratio of inertia force to viscous force and mathematically presented as.
Fluids with a Reynolds number between 2000 and 4000 are considered unstable and can exhibit either laminar or turbulent behavior. This region is commonly referred to as the critical zone and the friction factor can be difficult to accurately predict. Judgment should be used if accurate predictions of fluid loss are required in this region. Either Equation 1 or 3 are commonly used in the critical zone. If the Reynolds number is beyond 4000, the fluid is considered turbulent and the friction factor is dependent on the Reynolds number and relative roughness. For Reynolds numbers beyond 4000, the Moody diagram identifies two regions, transition zone and completely turbulent zone. The friction factor represented in these regions is given by the Colebrook formula which is used throughout industry and accurately represents the transition and turbulent flow regions of the Moody diagram.
The Colebrook formula for Reynolds number over 4000 is given in equation 3.
The roughness factor is defined as the absolute roughness divided by the pipe diameter or . Typical values of absolute roughness are 5.9x10-4 in (0.0015 mm) for PVC, drawn tubing, glass and 0.0018 in (0.045 mm) for commercial steel/welded steel and wrought iron [3].
The Colebrook equation has two terms. The first term, ()/3.7, is dominate for gas flow where the Re is high. The second term, , is dominate for fluid flow where the relative roughness lines converge (smooth pipes). In the “Complete Turbulence” region, the lines are “flat”, meaning that they are independent of the Reynolds Number. In the “transition Zone”, the lines are dependent on Re and . When the lines converge in the “smooth zone” the fluid is independent of relative roughness.
Liquid (Incompressible) Flow
For liquid flow, equation 4 has been used by engineers for over 100 years to calculate the pressure drop in pipe due to friction. This equation relates the various parameters that contribute to the friction loss. This equation is the modified form of the Darcy-Weisbach formula which was derived by dimensional analysis.
The friction factor in this equation is calculated by equation 3 for a specified Reynolds number and roughness factor using an iterative method or a trial and error procedure.
Gas (Compressible) Flow
For gas flow, density is a strong function of pressure and temperature, and the gas density may vary considerably along the pipeline. Due to the variation of density, equation 5 should be used for calculation of friction pressure drop.
Again, the friction factor in this equation is calculated by equation 3 for a specified Reynolds number and roughness factor using a trial and error procedure. Actual volume flow rate is needed to calculate the velocity of gas in the line from which the Reynolds number is calculated. Equation 6 may be used to convert the volume flow rate at standard condition to the actual volume flow rate.
Case Study 1: Oil Pipeline
Consider a 16-inch (inside diameter of 395 mm) oil export line for transportation of 170,000 bbl/day (0.313 m3/s) of a 43 API crude oil (relative density of 0.81) from an offshore platform to the shore oil terminal. The total length of pipe is 55 km. The ambient temperature is 5 °C and the crude oil viscosity at the average pipe temperature is 0.001 cP. The pipe line inlet pressure is 14.9 MPa (absolute). Since the objective is to study the effect of roughness factor on friction pressure drop, we will ignore elevation change.
To study the effect of roughness factor on friction pressure drop, was varied from 1x10-6 to 1x10-3. The roughness factor of = 1x10-6 represents a very smooth pipe. The calculated friction pressure drop as a function of the roughness factor is plotted in Figure 2. For each value of roughness factor, the percent change in frictional pressure drop was calculated in comparison to a very smooth pipe (= 1x10-6) and the results are presented in Figure 3. The calculated results are also presented in Table 1.
Case Study 2: Gas Pipeline
Let’s consider an 8-inch (inside diameter of 190 mm) gas export line for transportation of 70 MMSCFD (22.913 Sm3/s) of natural gas with a molecular weight of 19.3 (relative density of 0.67) from an offshore platform to the shore. The total length of pipe is 43 km. The ambient temperature is 5°C and the gas viscosity at the average pipeline temperature is 1.1x10-6 cP. The gas inlet temperature is 35°C and pressure is 13.0 MPa (absolute). Since the objective is to study the effect of roughness factor on friction pressure drop, we will again ignore elevation change.
Similar to the oil pipeline, the roughness factor, was varied from 1×10-6 to 0.006. Note, for a roughness factor greater than 0.006, a higher inlet pressure, a larger diameter or lower flow rate was needed. The calculated friction pressure drop as a function of roughness factor is presented in Figure 2. For each value of roughness factor, the percent change in frictional pressure drop in comparison to a very smooth pipe (= 1×10-6) was calculated and the results are presented in Figure 3.
Discussion and Conclusions
The analysis of Figure 2 indicates that for the oil pipeline, the friction pressure drop is almost independent of the roughness factor in the range of 1×10-6< <1×10-4; however, for >1×10-4, it will increase with . For liquid lines, the Reynolds number is normally in the range of 5×104 to 1×106. For this range, the friction factor curves in Figure 1 approach close to each other so the values of friction factors become close to each other.
Contrary to the oil pipeline, the friction pressure drop for the gas pipeline is a strong function of . As can be seen in Figure 2, friction pressure drop increases very rapidly with the roughness factor. Figure 3 shows the comparison of percent change of friction pressure drop between oil and gas pipelines as a function of roughness factor. For the liquid pipeline, the maximum change is 20 % but for the gas pipeline the maximum change is more than 200 %. Again this can be explained by referring to Figure 1. For gas pipelines, the Reynolds number is higher than in the liquid line and the range is normally 5×106<Re<1×108. For this range, the friction factor curves in Figure 1 are apart from each other, so the friction factors are not close.
In summary, contrary to liquid pipelines the gas pipelines are very sensitive to wall roughness and using smooth pipe can reduce friction pressure drop considerably. This in turn lowers the OPEX. Therefore, regular pigging to clean the pipe surface is done to lower the roughness factor. The modern gas transmission companies will add a Fusion Bounded Epoxy (FBE) liner to gas pipelines because the pipe is sensitive to roughness. This lowers OPEX for the long term. It should be noted that the smoother the pipe, the higher the CAPEX, so as always, detailed total cost analysis should be performed for engineering applications.
Due to the sensitivity of gas pipelines to roughness factor and other operation parameters, there are numerous gas flow equations (e.g. Weymouth, Panhandle A and B, AGA) to best fit certain design conditions [1].
To learn more about similar cases and how to minimize operational problems, we suggest attending our ME44 (Overview of Pumps and Compressors in Oil and Gas Facilities), ME46 (Compressor Systems – Mechanical Design and Specification), PL4 (Fundamental Pipeline Engineering), G40 (Process/Facility Fundamentals), G4 (Gas Conditioning and Processing), and PF4 (Oil Production and Processing Facilities) courses.
By: Dr. Mahmood Moshfeghian
Reference:
Honeywell, Joe, “Friction Pressure Drop Calculation,” Campbell Tip of the Month, Feb 2007
Campbell, J. M., “Gas Conditioning and Processing, Vol. 1, the Basic Principals, 8th Ed., Campbell Petroleum Series, Norman, Oklahoma, 2001
Menon, E.S, Piping Calculations Manual, McGraw-Hill, New York, 2005
In this tip of the month, we will discuss how miscalculations and incorrect analysis of potential process upsets can affect process safety. There are many aspects in facility design engineering and process safety engineering that should be considered when designing a new facility or debottlenecking an existing one. During these times of compressed schedules and budgets, it can become difficult to ensure all project deliverables receive the proper amount of checking and documentation. Mistakes in engineering design and operations of the following systems can result in serious safety incidents which must be avoided. Quality control, technical training, calculation checking and method verifications can aid in minimizing safety risks in these systems. This months’ tip will focus on Pressure Relief Systems.
A primary process system in oil and gas facilities requiring careful attention is the Pressure Relief System. The most common components in upstream pressure relief systems are:
Protected Equipment
Emergency Shut Down Valves
Depressurization Valves
Pressure Safety Valves (PSV)
Pressure Safety Valves Inlet and Discharge Piping
Flare Header
Flare Knock Out Drum
Flare Stack / Tip
The primary purpose of the pressure relief system is to ensure that the operation’s personnel and equipment are protected from overpressure conditions that happen during process upsets, power failures, and from external fires. In some locations and facilities, it is accepted practice to vent the pressure safety valves directly to atmosphere provided the process fluid is discharged at sufficient velocity to ensure good dispersion and that the fluids molecular weight is lighter than air. In this TOTM we will be discussing components in the pressure relief system in which detailed engineering calculations must be completed to select and install properly.
Pressure Safety Valves:
The purpose of a pressure safety valve is to protect equipment and / or piping from any possible overpressure scenario. There are multiple industry recommended practices and standards that govern the sizing, selection and installation of pressure safety valves. Many of these are referenced in this TOTM. A study that was conducted by Berwanger, et al. [1] determined that only 65% of upstream processing facility pressure safety valves meet the existing standards. Accurate pressure safety valve relieving requirements, scenario analysis and installation design is critical to ensure safety of the equipment and the operations staff during an upset condition. The American Institute of Chemical Engineering found that roughly 30% of process industry losses have been found to be partially attributed to deficient pressure relief systems [2]. If an upset process condition occurs with a system that has a pressure safety valve that is missing, undersized, or not properly installed, there is a potential that the equipment will not be protected and will mechanically fail. This could result in a significant loss of fluid containment and potential fatalities depending upon the fluids contained within the process.
On March 4, 1998, there was a major vessel failure at a Sonat Exploration facility in Pitkin Louisiana. The vessel failure and subsequent fire resulted in four deaths. A cause of the incident was failure of a low pressure vessel open to a high pressure gas source that was not provided with any pressure relief devices [3].
In determining the relevant relieving scenarios for a pressure relief valve, it is essential that the engineer doing the evaluation has a solid understanding of the process and the process control design within the facility. If the lead engineer is conducting an existing plant review or working on the design of a new facility, it is critical that they evaluate all potential relieving scenarios that may be required. If a scenario is missed, then there is a possibility that the system will not be protected if that missed scenario was the limiting case. ANSI / API Standard 521 ISO 23251, 5th Edition [5] specifies requirements and provides guidelines for examining the principal causes of overpressure; determining individual relieving rates; and selecting and designing disposal systems, including the details on specific components of the disposal system. Only with experience and training do engineers develop the competency level to complete these evaluations effectively. Participation in Process Safety Hazards Reviews and Analyses promote the development of an engineer’s skills in identifying and resolving potential process hazards, and can help develop a junior engineer’s skills and understanding of the evaluation of these systems.
API Recommended Practice 520, 7th Edition, Part 1 [4], and the International Organization for Standardization (ISO) Standards in the 4126 series (will not all be referenced here, and it should be noted that these only apply to systems designed and installed in the European Union Member States), addresses the methods to determine the pressure safety valve sizing requirements for the different relieving scenarios and provides guidance on how to select the proper relief valve type. Both over-sizing and under-sizing a relief valve can result in mechanical failures, thus it is critical that the valve sizing and selection are correct.
If a facility is being debottlenecked and modified all pressure safety valves that will be affected by the modification must be checked for adequate capacity. Many facilities are not applicable to the U.S. Occupational Safety and Health Administration (OSHA) Process Safety Management (PSM) Standard 29 CFR 1910.119 [6].It is strongly recommended that a Management of Change (MOC) procedure be used to ensure that no facility modification will pose a safety risk or undermine the existing safety equipment provided within the facility. Pressure safety valves for all modified systems must be verified to safely handle the new required rates and compositions that result from debottlenecking the facility.
In addition, it is essential that operation’s personnel are trained in the proper handling and testing of relief valves. There have been cases when operations and maintenance personnel have increased the set pressure on a pressure relief valve that was frequently relieving. The increase in set pressure results in the vessel operating above the stamped maximum allowable working pressure and may result in mechanical failure. Trained staff will understand that the solution to the problem is to correct the process condition that is resulting in the high pressure, not increase the set point on the pressure safety valve.
PSV Inlet and Discharge Piping
Another area that requires close attention is the proper design of the inlet and discharge piping of the pressure safety valves. API Recommended Practice 520, 5th Edition, Part 2 [7], and ANSI / API Standard 521 [5] provide guidance on the installation and design of the inlet and discharge piping for pressure safety valves.
For inlet piping to pressure safety valves, the recommended practice is to maintain the inlet hydraulic losses at no more than 3% of the set pressure of the pressure safety valve. This is because the relief valve is designed to normally close at 97% of the set pressure. A PSV with no inlet flow will sense the same pressure as exists in the protected equipment. Once open however, the pressure at the inlet to the relief will be the pressure at the protected equipment minus the friction loss in the inlet line. If this friction loss exceeds 3%, the valve will close and then reopen once the flow stops. This chattering can destroy the valve. Over sizing of a pressure safety valve can also result in “chatter” from essentially the same phenomenon. There is a potential for pressure relief valve or piping failure from prolonged “chattering” due to mechanical fatigue and potentially thermal fatigue.
If the inlet piping design cannot be configured to meet this requirement, then the use of a remote sensing pilot pressure safety valve can be used. This is not preferred due to the potential for the sensing line to plug or freeze.
Typically, relief valves are mounted almost directly on the equipment they protect. You will often find, however, that in existing plants this is not always the case. Some pressure safety valves may be located remotely with long inlet lines and the 3% criteria must be carefully checked. Even with new plant designs, there are times when the piping designer must locate the pressure safety valve remotely. It is important to always check the inlet line losses by utilizing the piping isometric drawings.
A study conducted by Berwanger, et al [1], found that 16% of all pressure safety valve installations reviewed were out of compliance with accepted engineering practices and standards as a result of improper installations. 35.5 % of these valves were out of compliance due to excessive inlet pressure drop. Experience indicates that in many older plants, the pressure safety valve inlet and discharge piping is set at the inlet size and outlet size of the pressure safety valve and the pressure drop calculations were not performed – or were performed on incorrect assumptions for inlet pipe routing. A crude oil fire occurred in a Shell facility as a result of improper inlet piping design. This caused severe vibration and caused a 6” flange to fail, losing containment of the process stream [8].
For systems with 600# ratings and above, the valve manufacturer may supply a relief valve with an inlet flange rating of 600# and an outlet flange rating of 300#. Be aware that a “typical” 150# flange rating on the PSV discharge piping is not always acceptable for the higher pressure systems. The velocity at the outlet of the pressure safety valve can not exceed sonic. Thus, for high pressure systems the flow through the relief valve may require a pressure greater than the max pressure rating of a 150# system to maintain sonic flow. It is important to check the pressure required to maintain sonic based on the size of the pressure safety valve outlet. If a 300# flange is required then a 300# pipe fitting is installed to expand the pipe to a diameter where the pressure corresponding to a 150# system is not exceeded. For large systems, it is recommended to use a flare network software program to predict the backpressure at the outlet of each pressure safety valve for various relief scenarios. During a fire, several reliefs may open simultaneously and the backpressure must be known at the outlet of each relieving pressure safety valve under these circumstances.
The piping design for the inlet and discharge of pressure safety valves should be reviewed to determine that the piping can meet the mechanical and thermal stresses that will develop when the pressure safety valves relieve. Threaded connections for high set pressure safety valves or on pressure safety valves that are installed near vibrating equipment are not recommended. The threaded connections have a tendency to fail or become “unscrewed” from the vibrations, and / or forces during relieving
Proper valve and discharge piping support design is essential. Piping and valve support becomes more critical on larger pressure safety valves and pressure safety valves that have high set pressures discharging to atmosphere. The reaction forces that can develop from the valves relieving to atmosphere can be significant. Even though the outlet piping may not be excessively long, the internal thrust created at the 90 degree elbow as the discharge piping turns up can be excessive. The flow will most likely be sonic velocity at the elbow and the discharge vent must be adequately supported to prevent failure. One incident occurred when the inlet piping on a 4X6 pressure safety valve set at 1350 psig failed. The valve became a projectile as a result. Fortunately, no one was hurt by flying debris and the gas line was isolated before the vapor cloud was ignited. This “near miss” was likely the direct result of poor welding and poor support on the valve installation.
The reaction forces in closed systems tend to be less, but in some cases the reaction forces in a closed system can become significant if there are sudden large pipe expansions or during unsteady flow conditions within the piping. Inadequate design and supports for pressure safety valves and the associated piping can result in mechanical failure during a relieving event.
Flare Header Design
If the pressure safety valve discharges into a flare header the superimposed and built up back pressure is critical and can impact the valves relieving capacity if the actual back pressure is higher than the originally calculated or assumed back pressure. The maximum allowable back pressure at which a pressure safety valve can function properly depends upon the type of the pressure safety valve. A study conducted by Berwanger, et al [1], found that almost 24% of all PSV installations reviewed were out of compliance with accepted engineering practices and standards because of improper installations. 12 % of these valves were out of compliance due to the outlet pressure drop being too high. If the built up back pressure is greater than the maximum value the valve can function with, then the upstream pressure of the valve will increase above the set pressure of the valve as a result. This condition increases the likelihood of a failure.
A flare network software program should be used to calculate backpressure in large relief systems. For most pressure safety valves the maximum flow that can pass through the orifice size is larger than the required relieving flow. The maximum flow must be used to calculate the inlet line loss and the resulting backpressure. Modulating pilot valves can be used, if required, to control the maximum flow that is required to be relieved. In the design of the flare system, several types of valves are available, as explained in API 520 Part 1 [4]. Conventional, bellows, and pilot valves are typically used. The valve manufacturer must be consulted to define the maximum flow and backpressure requirements for each type of valve. The final flare design can not be completed until the actual pressure safety valves have been selected.
Depending upon the fluids which are being relieved and the pressures involved, it is possible to have relieving events that require stainless steel discharge piping, Flare Header, Flare KO Drum and Flare Stack because of cryogenic relieving temperatures from the Joule-Thompson Effect through the pressure safety valve. There have been multiple cases where carbon steel flare headers have failed due to the cryogenic relieving temperatures that developed during relieving events. The failure of a flare header completely undermines the purpose of the Pressure Relief System, and can result in a catastrophic event.
In today’s’ market, the recovery of NGL’s from natural gas is quite common. Particular attention is required in designing the relief systems for the cryogenic vessels. The pressure safety valves most likely will be relieving cold (at -20 F or below) two phase fluids. The pressure safety valve downstream piping will be exposed to very cold temperatures when the valves relieve. The recommended method for sizing two phase flow valves is by utilizing the DIERS equations. API 520 Part 1, Appendix D [4] summarizes these equations and provides an example calculation. The calculation procedure is long and tedious but it is recommended to perform a hand calculation before utilizing in house spreadsheets. The couple of hours spent performing the calculation will provide valuable insight to the key parameters used in the equations and will serve as a verification check of a spreadsheet.
There should be no dead legs in any piping from the discharge of the relief valve to the Flare KO Drum. Any pockets or dead legs can fill with liquids which may result in excess back pressure during relieving events There may also be large reaction forces in the flare header as a result of the slug of liquids forced down the header. In 1999, the flare header of a Tosco refinery in California was overpressured due liquid accumulation at a low point in the flare header. This resulted in a facility shutdown.There were no injuries reported [9].
Flare KO Drum and Flare Stack / Tip
Flare KO Drum and Flare Stack sizing is also critical to the safety of the plant. Oil and Gas Industry Flares are designed to destroy vapor streams only and require an adequately sized Flare KO Drum to prevent flammable liquids from raining out of the flare tip. In determining the sizing, it is important that a Flare Study be conducted to determine the worst case scenario for Flare KO Drum and Stack capacity and to select the proper droplet size separation criteria that the selected flare tip can adequately destroy. ANSI / API Standard 521[5] provides guidance on sizing, design and selection of this equipment.
A good example of the consequences of liquids flowing out of a Vent Stack was the Texas City Refinery explosion of 2005. This catastrophic incident resulted in a process upset where the amount of liquids that flowed to the KO Drum overwhelmed the drum size, and flowed up the vent stack and to the surrounding atmosphere which resulted in the tragic explosion [10]. If a Flare would have been installed in the Texas City Refinery rather than a Vent Stack, the consequences of the event would have been reduced. The vapor phase hydrocarbons that were originally flowing to the vent stack would have been destroyed in the Flare Tip, and the vapor cloud that exploded would have been prevented. Flowing liquid hydrocarbons to a Flare Tip is still a dangerous situation. If a Flare KO Drum were overwhelmed with hydrocarbon liquids the Flare Stack would likely be raining fire, and not liquid hydrocarbons.
Based on the stack sizing, ANSI / API Standard 521 [5] outlines procedures to estimate the radiation effect from the flare. With today’s’ specialized design of flare stacks, consultation with the flare manufacturer is recommended for the radiation confirmation.
Depressurization Valves
In the gas processing industry, it has become a standard practice to block in the treating facility with Emergency Shut Down (ESD) Valves rather than depressure the entire facility to the flare. One primary reason for this philosophy is that natural gas fires are not equivalent to liquid hydrocarbon pool fires. Natural gas fire protection and mitigation requires different protection methods than for those used for fighting liquid hydrocarbon pool fires, which can be extinguished using a fire water system or a foam system. . It is standard natural gas industry practice to isolate the hydrocarbon gas sources to the facility and evacuate all personnel from the facility. Once the source of the gas is isolated, the feed to the fire is terminated and the fire is quickly extinguished from lack of fuel.
In the case where a facility must be depressurized in an upset condition, careful attention must be given to the design of the depressurization valves, their timing and flare capacity. There exists the potential to overwhelm the Flare Tip if the Tip was not designed for the high depressurization rates. In addition, consideration for required depressurization time, resulting Flare Header temperatures, and depressurization control schemes must be given close attention. These systems can be highly complex due to the transient nature of the process and require careful design procedures to ensure a safe Depressurization System.
By: Kindra Snow-McGregor
Senior Process Consultant and Instructor
References:
Non-Conformance of Existing Pressure Relief Systems with Recommended Practices, A Statistical Analysis, Patrick C. Berwanger, PE, Robert A Kreder, and Wai-Shan Lee. Berwanger, Inc., 2002.
AIChE. Emergency Relief System (ERS) Design Using DIERS Technology. American Institute of Chemical Engineers, New York, NY, 1995.
U.S. Chemical Safety and Hazard Investigation Board, Investigation Report, Catastrophic Vessel Overpressurization, Report No. 1998-002-I-LA.
Sizing, Selection, and Installation of Pressure-Relief Devices in Refineries, Part 1 – Sizing and Selection, API Recommended Practice 520, 7th Edition, January 2000.
ANSI / API Standard 521, / ISO 23251, Pressure Relieving and Depressuring Systems, 5th Edition, January 2007.
Occupational Safety and Health Standards, Process Safety Management of Highly Hazardous Chemicals, 29-CFR-OSHA-1910.119, 57 FR 23060, June 1, 1992; 61 FR 9227, March 7, 1996.
Sizing, Selection, and Installation of Pressure-Relief Devices in Refineries, Part 2 – Installation, API Recommended Practice 520, 5th Edition, August 2003.
Poor Relief Valve Piping Design Results in Crude Unit Fire, Politz, FC., API Mid-year Refining Meeting, 14 May 1985, Vol / Issue 64.
Gas density estimates are of fundamental importance for process simulation, equipment design, and process safety engineering. In the previous Tip of the Month (TOTM), two shortcut methods for predicting sour and acid gas density were evaluated. We showed that Katz correlation gives accurate results for lean sweet gases and it is the most accurate in comparison to Wichert-Aziz method or the SRK EOS. For binary mixtures of CH4 and CO2, Wichert-Aziz method gives the most accurate result for CO2 content of between 10 and 90 mole percent. As H2S and CO2 content increased, the accuracy of the Katz correlation decreased, but its accuracy increased as the mixture approached a single component. The percentage difference between the Katz and Wichert-Aziz [1] methods for gas mixtures containing acid gases was greater for H2S than CO2.
Process simulation software often use the Benedict-Webb-Rubin-Starling (BWRS), Soave-Redlich-Kwong (SRK) and/or Peng-Robinson (PR) equations of state for gas density calculations. Other sources of gas density calculation are NIST REFPROP (Reference Fluid Thermodynamic and Transport Properties) program and GERG-2004 [2, 3], a reference equation of state for natural gases.
Due to the importance of CO2 injection for enhanced oil recovery and the increasing interest in CO2 capture and sequestration, this study was undertaken to evaluate the accuracy of density calculations for gases containing nil to 100% CO2. An experimental data base was used for the basis of comparison. The study reviews all of the above mentioned methods and will report their accuracies. Table 1 presents the summary of the temperature, pressure, and CO2 mole percent ranges for the data used in this study. The sources of experimental data were reference [4, 5]. This table also presents the average absolute average percent error and the overall average percent error.
Table1 – Summary of error analysis and comparison of accuracy of sour gas and acid gas density prediction by several methods:
Table 1 provides the overall accuracy of the various methods. It should be noted that the relative accuracy of each method varies depending on the CO2-CH4 proportion, the temperature and pressure. The AGA 8 did not return values for many of the low temperature cases where two phases were present. These points were ignored in the analysis.
Next, we plotted the experimental data reported in the GPA RR-138 [3] and GPA RR 68 [4] to evaluate the accuracy of Katz, Wichert-Aziz and the best four of the detailed methods. The results of this evaluation for the T=350°K and 320°K cases are shown in Figures 1 through 5, for CO2 content of 9.83 to 100 mole percent. In Figure 1, Katz method is the most accurate and the accuracy of the other methods are almost the same.
In Figure 2, Katz method has the least accuracy and even though the accuracy of the other methods look the same, GERG 2004 is slightly better than the others.
In Figure 3, Katz method again has the least accuracy and even though the accuracy of the other methods look the same, AGA8 provided slightly better estimates than the others.
In Figure 4, Wichert-Aziz method has the least accuracy and even though the accuracies of the other methods look the same, AGA8, GERG-2004 and REFPROP are slightly more accurate than the PR EOS.
In Figure 5, Wichert-Aziz method has the least accuracy and REFPROP, GERG 2004 and AGA 8 equally have the best accuracy.
Based on the work done in this study and in the previous TOTM, the following can be concluded:
Katz correlation gives accurate results for pipeline quality gases (lean sweet gases)
For pure CO2, AGA 8, REFPROP, and GERG 2004 methods equally are the most accurate method
For binary mixtures of CH4 and CO2, REFPROP and GERG 2004 methods equally give the most accurate result for CO2 content of between 10 and 90 mole percent.
As CO2 content increases, the accuracy of the Katz correlation decreases, but its accuracy increases as the mixture approaches a single (pure) component.
The Peng-Robinson EOS provides a better density estimate than the SRK EOS.
Results from either the PR or the SRK EOS in ProMax are slightly more accurate than the comparable results from HYSYS.
Binary interaction parameters which have been optimized to predict VLE behavior may not provide the best density prediction.
At several low temperatures, AGA8 did not provide density estimates. The average errors reported here ignored these missing data. Note that AGA8 is not valid for liquid nor for the extended region near the critical point.
Table 1 indicates that REFPROP and GERG 2004 give equally the best results.
Wichert, E. and Aziz, K., Hydr. Proc., p. 119 (May 1972).
Lemmon, E.W., Huber, M.L., McLinden, M.O. NIST Standard Reference Database 23: Reference Fluid Thermodynamic and Transport Properties-REFPROP, Version 8.0, National Institute of Standards and Technology, Standard Reference Data Program, Gaithersburg, 2007.
Kunz, O., Klimeck, R., Wagner, W., and Jaeschke, M. “The GERG-2004 Wide-Range Equation of State for Natural Gases and Other Mixtures,” GERG Technical Monograph 15 (2007)
Hwang, C-A., Duarte-Garza, H., Eubank, P. T., Holste, J. C. Hall, K. R., Gammon, B. E., March, K. N., “Thermodynamic Properties of CO2 + CH4 Mixtures,” GPA RR-138, Gas Processors Association, Tulsa, OK, June 1995
Hall, K. R., Eubank, P. T., Holste, J., Marsh, K.N., “Properties of C02-Rich Mixtures Literature Search and Pure C02 Data, Phase I,” GPA RR-68, A Joint Research Report by Gas Processor Association and the Gas Research Institute, Gas Processors Association, Tulsa, OK, June 1985
Gas density is needed for process simulation and equipment design. For example, accurate predictions of gas density are needed for calculation of pressure drop in piping/pipeline and for vessel sizing. Accurate gas density is also essential for custody transfer metering. Gas density, , is calculated by:
(1)
Where:
&Gas density, kg/m3 (lbm/ft3)
Absolute temperature, K (ºR)
Pressure, kPa (psia)
MW Molecular weight kg/kmole (lbm/lbmole)
Gas compressibility factor
Universal gas constant, 8.314 (kPa)(m3)/(kmole)(K) or 10.73 (psia)(ft3)/(lbmole)(ºR)
In equation 1, “z” represents gas compressibility factor. For ideal gases, “z” is equal to 1. Gas densities are sometime expressed in terms of relative density (specific gravity), , and is defined as:
(2)
Substituting Equation 1 for gas and air into Equation 2 and assuming ideal gas behavior at standard conditions, Equation 2 will be transformed to:
At the standard condition and for simplicity, Equation 3 can be written as
In Equation 1, the key parameter is the compressibility factor “z”, which is a function of pressure, temperature and gas composition. Compressibility factor is a dimensionless surrogate of non-ideal gas density. In general, equations of state are probably the most widely used for calculation of z. They are not necessarily the most accurate. Empirical correlations developed for a specific mixture or a narrow range of mixtures provide better accuracy, but may be less general. An example would be the Katz chart which is quite good when applied to “sweet” pipeline quality gases, but less reliable for gases containing H2S, CO2 and/or N2. Figure 3.2 in Chapter 3 of Gas Conditioning and Processing [1] shows the Katz chart for sweet natural gases as prepared by Standing and Katz [2]. The chart was developed by using experimental data on methane binary mixtures with ethane, propane, butane and other natural gases over a wide range of composition with a maximum molecular weight of 40.
For fiscal metering of natural gas, an accurate experimental database has been developed and compressibility factor correlations, with uncertainties generally within ±0.2%, have been published in the industry standards, AGA Report No. 8 and ISO 12213. A summary of some common “z” correlations and their effect on gas measurement accuracy can be found in reference [3]. Since many people use the Katz compressibility factor chart, the question is often asked how it may be extended to gases containing H2S and CO2. There are two methods available for this application.
The approach proposed by Robinson et al. [4]
The approach proposed by Wichert and Aziz [5]
In this Tip of the Month (TOTM) we will demonstrate the accuracy of the second approach. The details of this method are presented in Chapter 3 of Gas Conditioning and Processing [1].
Let’s consider the gas mixture shown in Table 1 with total acid gas (H2S and CO2) of 14.68 mole percent. At 13.94 MPa (2021 psia) and 58 ºC (136 ºF), the compressibility factors are 0.797 (120.1 kg/m3) and 0.832 (114.8 kg/m3), using Katz chart and Wichert-Aziz method respectively. The percent deviation between two answers from each other is 4.4%.
In order to show the effect of acid gas on compressibility factor determined from Katz chart and Wichert-Aziz methods, we varied the acid gas content of the gas in Table 1 from 0 to 37 mole percent. This was accomplished by diluting the non-acid gas components with a 50:50 mixture of CO2 and H2S. Figure 1 presents the percentage difference between the two methods as a function of acid gas content. The graph shows that as the H2S and CO2 content increases, the deviation of Katz chart from Wichert-Aziz method increases almost linearly. This graph also indicates that the percentage difference between the two methods is greater for the case of diluting gas with only H2S than only CO2.
Next, we used the experimental data reported in the GPA RR-138 [6] and GPA RR 68 [7] to evaluate the accuracy of Katz, Wichert-Aziz and SRK equation of state (EOS) for binary mixtures of CO2 and CH4. The results of this evaluation are shown in Figures 2 through 6, for CO2 content of 9.83 to 100 mole percent. The figures indicate that the Katz correlation accuracy decreases as the mole percent of CO2 increases. However; Figure 5 indicates that as the gas becomes very rich in CO2, the accuracy of the Katz correlation and the Wichert-Aziz method are practically identical. Figure 6 shows that the Katz correlation best predicts the density of pure CO2, and also when the gas approaches pure CH4. The experimental data for pure CO2 in Figure 6 is from GPA RR 68 [7]. Figure 2 through 6 also indicate that the SRK EOS has low accuracy. In this study, a binary interaction parameter of 0.12 between CH4 and CO2 which had been determined from experimental vapor-liquid-equilibrium (VLE) data was used.
Based on the work done in this study, the following can be concluded:
Katz correlation gives accurate results for pipeline quality gases (lean sweet gases)
For pure CO2, Katz correlation is the most accurate in comparison to Wichert-Aziz method or the SRK EOS.
For binary mixture of CH4 and CO2, Wichert-Aziz method gives the most accurate result for CO2 content of between 10 and 90 mole percent.
As H2S and CO2 content increases, the accuracy of the Katz correlation decreases, but its accuracy increases as the mixture approaches a single (pure) component.
The percentage difference between the Katz and Wichert-Aziz methods for gas mixtures containing acid gases is greater for H2S than CO2.
Binary interaction parameters which have been optimized to predict VLE behavior, may not provide the best density prediction.
Campbell, J. M., and Hubbard, R. A., Gas Conditioning and Processing, Vol. 1 (8th Edition, 2nd Printing), Campbell Petroleum Series, Norman, Oklahoma, (2001).
Standing, M.B. and Katz, D.L.; “Density of Natural gas gases,” AIME Trans., 146, 140-49 (1942)
Hannisdal, N.E., “Gas Compression Equations Evaluated,” Oil and Gas J., p. 38-41 (May 4, 1987)
Robinson, D. F. et al. Trans. AIME, Vol 219, P. 54, (1960).
Wichert, E. and Aziz, K., Hydr. Proc., p. 119 (May 1972).
Hwang, C-A., Duarte-Garza, H., Eubank, P. T., Holste, J. C. Hall, K. R., Gammon, B. E., March, K. N., “Thermodynamic Properties of CO2 + CH4 Mixtures,” GPA RR-138, Gas Processors Association, Tulsa, OK, June 1995
Hall, K. R., Eubank, P. T., Holste, J., Marsh, K.N., “Properties of C02-Rich Mixtures Literature Search and Pure C02 Data, Phase I,” GPA RR-68, A Joint Research Report by Gas Processor Association and the Gas Research Institute, Gas Processors Association, Tulsa, OK, June 1985
The best way to prevent hydrate formation (and corrosion) is to keep the pipelines, tubing and equipment dry of liquid water. There are occasions, rightly or wrongly, when the decision is made to operate a line or process containing liquid water. If this decision is made, and the process temperature is below the hydrate point, inhibition of this water is necessary. Many materials may be added to water to depress both the hydrate and freezing temperatures. For many practical reasons, an alcohol or one of the glycols is injected as an inhibitor, usually methanol, diethylene glycol (DEG) or ethylene glycol (EG). All may be recovered and recirculated, but the economics of methanol recovery may not be favorable in many cases. Total injection rate is that needed to provide the necessary inhibitor concentration in the liquid water plus that inhibitor which enters the vapor and hydrocarbon liquid phases. Any inhibitor in the vapor phase or liquid hydrocarbon phase has little effect on hydrate formation conditions. Determination of the amount and concentration of inhibitors and their distribution in different phases are very important for practical purposes and industrial applications. Therefore, in order to determine the required amount and concentration of these inhibitors, several thermodynamic models for hand and rigorous calculations have been developed and incorporated into the computer software.
In this Tip of the Month, we will demonstrate the effect of total inhibitor circulation rate on hydrate depression in a sales gas dew point correction plant. Let’s consider the process flow diagram shown in Figure 1. The feed composition and conditions are shown in Table 1. The wet gas analysis was used in this study.
It is desired to process this feed gas to produce a sales gas with a dew point of -5˚F (-20.6˚C) at 990 psia (6.826 MPa) The feed gas is mixed with recycle gas from the NGL/Glycol separator, compressed and cooled to 120˚F (48.9˚C) and 1000 psia (6.895 MPa), and EG lean solution and then cooled in the HTX-1, and finally in the HTX-2 to -5˚F (-20.6˚C) before entering the separator at 990 psia (6.826 MPa). For the sake of simplicity, the refrigeration cycle for HTX-2 system is not shown in the above process diagram. Figure 2 represents the feed phase envelope, its hydrate curve, sales gas envelope and the cooling path. As seen in this figure, the gas temperature drops below the hydrate formation temperature of about 65˚F (18.3˚C) in the HTX-1 and to -5˚F (-20.6˚C) in HTX-2. Therefore; to prevent the hydrate formation in the HTX-1 and HTX-2, it is decided to inject an 80 weight percent lean EG solution to the gas stream. The concentration of rich inhibitor solution can be calculated by the shortcut method of Hammerschmidt [1] as described in Chapter 6, Volume 1 of Gas Conditioning and Processing [2] or using rigorous thermodynamic models. The required circulation rate is then determined by material balance and phase equilibrium calculations.
In order to show the effect of EG solution circulation rate on the depression of the hydrate formation temperature, the whole process shown in Figure 1. Figure 3 shows how the hydrate formation temperature in HTX-2 drops with increasing total inhibitor circulation rate. The corresponding calculated weight percent of EG in rich solution is also shown as a function of circulation rate. The concentration of EG in the rich inhibitor solution increases with the increase in the inhibitor solution circulation rate. Figure 3 indicates that for a 10˚F (5.6˚C) depression, or the hydrate formation temperature of -15˚F (-26.1˚C), a circulation rate of 975 lbm/hr (442 kg/h) is required. At this circulation rate the corresponding weight percent of EG in rich solution drops to 74.
To learn more about similar cases and how to find the optimum inhibitor circulation rate and concentration for prevention of hydrate formation, we suggest attending our G4andG5courses.
Dr. Mahmood Moshfeghian
References:
Hammerschmidt, E.G., “Formation of gas hydrates in natural gas transmission lines,” Ind. & Eng. Chem., Vol. 26, p. 851, 1934
Campbell, J. M., “Gas Conditioning and Processing, Vol. 1, the Equipment Modules, 8th Ed., J. M. Campbell and Company, Norman, Oklahoma, 2001
