The optimal thermal design and selection of heat exchangers can be very challenging; it involves many variables. The thermal design of an exchanger with very viscous and high-fouling fluids in a refinery service can be quite complicated as compared to other condensing or reboiling services due to high costs and maintenance problems. In a refinery, high viscosity and high-fouling materials come from the crude column or vacuum-column bottoms (high-temperature zone). These materials have higher temperatures and, therefore, they offer a high energy potential. To use this high energy level, these streams are invariably used as heating media in preheat trains of crude and vacuum units. The challenges that a thermal designer faces when handling such fluids include:
Selecting the optimum number of shells
Utilizing pressure drops across the exchanger
Cleaning and maintenance
Maximizing and sustaining thermal performance over the years
Optimizing the cost of procurement.
This article discusses a case study where substantial savings were realized via optimizing heat exchangers handling viscous and fouling fluids.
During the design-review stage, a thermal reviewer from the project management consultancy team/client noticed that the surface area requirement was exceptionally high. The total design heat-transfer surface area was more than 5,000 m2 as the exchanger was designed with eight shells (four in series, two in parallel combinations). When reviewed, it was noted that, instead of routing heavy vacuum gasoil (HVGO) with lower viscosity and higher flowrate on tube side, the atmospheric residue (AR) was routed through the tube side, which has a comparatively higher viscosity and lower flowrate. This design resulted in several anomalies. Tables 1 and 2 list the salient design features for this exchanger network.
Thermal design revised. To make the basics correct, the thermal reviewer changed the fluid sides in 45° (rotated square) layouts using software and checked the performance. The HVGO side heat-transfer coefficient increased from 600 to 850, thus increasing U, the overall heat transfer coefficients, from180 to 230. Therefore, the surface-area requirement dropped from 5,050 m2 to 3,900 m2, and the number of tube passes decreased from eight to four. Shell-side and tube-side velocities were quite adequate 0.5 m/sec and 1.7 m/sec, respectively, and the pressure drop on the shell side and tube side were less than allowable. Due to the lower surface area, the number of tubes per shell dropped from 1,380 to 1,060 and the shell size from 1,500 mm to 1,300 mm ID. Table 3 summarizes these results.
Since the client preferred to use 19.05-mm tubes, this design was further improved with the smaller shell size. However, since tube-side pressure was exceeding the maximum allowable pressure drop with four passes, the number of passs was reduced to two and tube side velocity was 1 m/sec. Therefore, there was no further gain in surface area.
Cost saving. The overall comparison of the original design and the new design are listed in Table 4.
Analysis. Viscosity played a major role in revising the existing heat exchanger design. Generally, the higher the viscosity of the operating fluid, the greater the difficulty in the heat exchanger optimization. Invariably, the highly viscous fluids are also dirtier fluids in terms of fouling resistance and, thus, it acts like a double-edged sword on the thermal heat exchanger designer. Viscosity and fouling potential almost always go hand in hand except for lube oils.
The challenges faced by a thermal heat exchanger designer when designing a heat exchanger handling both high viscosity and fouling fluids are:
Minimizing heat transfer area
Optimizing number of shells in series and parallel arrangement
Choosing the correct sides for fluids
Maximizing the velocity in the shell and tube sides
Maximizing the calculated pressure drop utilization within allowable pressure drop
Exchanger geometry adopted to increase NRe.
Maximize heat transfer area. This is most important because the greater surface area means more materials required as the tube surface is linked to the cost of tubes, cost of shell, higher tube-sheet thickness, larger channel ID and thickness, etc. Higher tube-surface area leads to higher shell diameter as tube length, in most cases, will be limited to clients specification and market availability. Larger shell diameter will mean higher shell thickness. When tube material is stainless steel (SS) instead of carbon steel (CS), the savings on surface area has more value as SS costs are much greater than CS. In this case study, both the shell and channel diameter were reduced, and the tube-surface area decreased from 5,050 m2 to 3,950 m2. Therefore, the overall bundle weight was reduced. The SS weight being proportional to weight, the cost of the equipment was much lower.
Optimizing shell in series and parallel arrangement. When the total heat-transfer area required for a heat exchanger is greater than the maximum heat-transfer area that can be incorporated in one shell, multiple shells are applied. When removing the bundle, the maximum heat-transfer area is determined by the maximum permissible tube bundle weight, which, in turn, depends on the cranes carrying capacity. In larger plants such as 6 million tpy to 9 million tpy refineries, the tube-bundle weight can be 15 tons to 20 tons, and the maximum heat-transfer area will increase accordingly.
For fixed-tube-sheet exchangers, bundle weight is not a constraint so that a very large heat transfer (1,000 m2 or greater) can be incorporated in a single shell. In this case study, since the heat duty was quite large, the surface-area requirement was also high. Within client specified limits of 6,096 mm tube length, the number of shells would, therefore, be greater than one.
There was an additional requirement due to temperature cross (hot fluid-outlet temperature being lower than cold fluid-outlet temperature). When there is a temperature cross, two or more shells in series are adopted. The greater the cross, the greater the number of shells in series.
To improve the shell-side heat-transfer coefficient through utilizing allowable shell-side pressure drop and velocity, multiple shells in series are adopted. But if the pressure drop exceeds more than is allowable, then the shell-side fluid flow is divided by considering shells in parallel. This is very crucial to selecting the optimum number of shells in parallel and series-type arrangements such that pressure drop is maximized within the allowable limit. The shell- and tube-side velocities are in an acceptable range, and the U value is maximum leading to the least surface area. In this presented case study, two shells in series were required theoretically. However, those two shells were very large in ID, and the surface area and bundle weight were many times over the bundle weight limit set by the client in the project specification. Therefore, smaller shells were tried in series and parallel arrangements to arrive at the optimized surface area, the highest velocities and the optimized pressure drop.
Choose correct fluid sides.
Correct allocation of fluid sides, i.e., which fluid needs to be placed on the shell side or on the tube side, is very crucial for maintenance and operability of the exchanger. Along with this, there are other parameters that can influence section of fluid sides as they cause the tube-surface area to increase or decrease.
The thermal engineer aims to produce a thermal design, which is lesser in cost, but high in ease of operation and maintenance. Sometimes, there will be confusion in selecting a particular side and the design parameters for yield a decision contradiction. The safe bet will be making two designs having opposite fluid sides and then deciding which is best. The key factors that a thermal engineer must analyze when allocating fluid sides are:
Fluid viscosity, fouling and cleaning requirements
Fluid pressure and temperature
Fluid viscosity, fouling and cleaning requirement. Higher viscosity fluid between hot and cold fluids must be placed on the shell side to maintain more turbulence and corresponding higher heat transfer coefficients. In 80% of the cases, the higher viscosity fluid will have a higher fouling resistance; keeping it on the shell side will mean difficulty in maintenance as the tube bundle may have to be pulled out more frequently for cleaning. Cleaning the inside of tubes is preferable; as in this case, the bundle does not need to be pulled out.
So, there is a dilemma over what will be the more preferable option. In such situations, it is prudent to take clients in confidence and seek their views. Fouling resistance of AR was marginally higher than HVGO, but the viscosity of AR was much higher than HVGO. Therefore, it was prudent to take advantage of shell-side turbulence on the higher viscous atmospheric crude and get higher heat-transfer coefficients.
Fluid pressure and temperature. When the operating pressure fluid is allocated on the tube side, only channel/channel cover, the floating head cover (if required) and tubes will be affected. However, if the same is allocated on the shell side, along with tubes and floating head cover, the shell will also be affected and require higher pressure applicability. The latter being a higher cost option since shell length is more costly than channel length; shell length directly impacts the total cost for equipment.
High-temperature fluids require higher alloy steel or SS. Therefore, it is prudent to keep high-temperature fluids on the tube side as only the channel/channel cover, the floating head cover (if required) and tubes will be affected. However, if the same is allocated on the shell side, along with the tubes and floating head cover, the shell material will also require higher cost materials of construction. Due to the higher temperature of the HVGO, all shells required SS 317L materials in the original design. However, with the revised design, due to the lower temperature of the AR, colder shells A/B and shell E/F were chosen to be of CS only. Due to the same reason (lower temperature of shell-side fluid), the tube sheet of the colder shells A/B and shell E/F were chosen to be of CS + SS 317L clad (tube side) only. This was quite a cost savings!
Fluids with lower flowrates if routed through the shell side will result in higher shell-side coefficients, leading to higher overall coefficients. For same amount of fluid, the shell side provides more turbulence than the tube side due to the baffles that force fluid flow in diverted directions many times. In this case, the atmospheric crude flowrate was lower than the HVGO flowrate, and, therefore, the shell side was the better option.
Maximize shell and tube side velocity.
Velocity is an important criterion for improving heat transfer coefficients in any heat-exchanger design. Especially in high viscosity service, maintaining a decent velocity on the shell side or tube side is very important to obtain a good heat transfer coefficient that is linked to an overall heat transfer coefficient and optimized surface area. Atmospheric velocity in the shell side and the HVGO velocity in the tube side were adequate considering fouling tendency for both fluids.
Utilization of pressure drop.
In any thermal design, one of the main aims of a thermal engineer is to use the pressure drop both on the shell and tube sides to the fullest within the allowable pressure drop to maximize the calculated heat transfer coefficients. As the pressure drop is proportional to the square of velocity, an increase in the pressure drop means higher velocity is also linked to the heat transfer coefficient.
| Fig. 1. Series and parallel arrangement of shells. |
Tube layout arrangements.
When handling dirty fluids on the shell side, especially in refineries, floating head exchangers are used extensively where tubes are laid out on a square (90°) or rotated square (45°) pitch. With the shell-side fluid being dirty, outside surfaces of the tubes require periodic cleaning and, therefore, in such cases, TEMA specifies a minimum cleaning lane of 6 mm or ¼ in.
The basic guideline for selecting either a 45° or 90° layout is shell side Reynolds number. Square pitch (90°) will be more suitable when the Reynolds number is greater than 7,000 as the shell-side heat transfer coefficients will be higher than rotated square (45°) layouts in such turbulent flow condition. However, when the Reynolds number for the shell-side fluid is less than 4,000, then flow conditions are laminar, and the shell-side heat transfer coefficients will be higher than the 90° layouts. When the Reynolds number is between 7,000 and 4,000, it is ideal to check both 90° and 45° before finalization of design. In this case, after reversing fluid sides (atmospheric crude in shell side), it was noticed that the NRe varied in the range of 1,700 to 5,700 with existing square-tube pitch configuration. Shell-side heat transfer was 650 kcal/m2-hr-°C. Therefore, the rotated square configuration was also checked. With the rotated square layout, the NRe came out in the range of 1,400 to 4,800. The shell-side heat transfer was 940 kcal/m2-hr-°C.
The overall saving (1/4th of the total equipment cost) in this case study shows that a proper and thoughtful design can save substantial costs. In a refinery, there is a substantial number of equipment items, and each item may have a story to tell. The role of a design reviewer on a project management consultancy team can be quite demanding as the consultant has to check the design given by the contractor/licensor is as per clients specification and is optimum in cost and utility. When the client trusts the review team fully and gives proper support, the output from the team can result in huge cost savings as depicted here. HP
Fig. 2. Suitable tube layout arrangement for
this case history.
1 Heat Transfer Research Institute (HTRI)
2 Tubular Exchanger Manufacturers Association (TEMA)
|The authors |
Manas K. Mandal is a senior design engineer for Fluor Daniel India Pvt. Ltd. He has more than 25 years of experience in the field of process heat transfer, cost optimization studies, process design & operations, process revamp, project control and energy management. Prior to Fluor, he has worked for Hindustan Petroleums Mumbai refinery and Engineers India Ltd. Delhi. Mr. Mandal has presented many papers in various seminars on heat transfer, energy management and process improvement. Mr. Mandal holds a BTech. degree in chemical engineering from the Indian Institute of Technology, Delhi and Masters degree in financial management from Jamnalal Bajaj, Mumbai.