Jeht-vol13-no.2-pp.157-173_2006

See discussions, stats, and author profiles for this publication at: https://www.researchgate.net/publication/245406620

Industrial Refrigeration and Ammonia Enhanced Heat Transfer Article  in  Journal of Enhanced Heat Transfer · April 2006 DOI: 10.1615/JEnhHeatTransf.v13.i2.50

CITATIONS

READS

8

473

1 author: Zahid H. Ayub Isotherm, Inc. 127 PUBLICATIONS   1,022 CITATIONS    SEE PROFILE

Some of the authors of this publication are also working on these related projects:

Characterization of Advanced Heat Exchanger Performance of Single and Multi-Phase Flows View project

Shell side DX evapoartion View project

All content following this page was uploaded by Zahid H. Ayub on 23 October 2016. The user has requested enhancement of the downloaded file.

Journal of Enhanced Heat Transfer, 13(2) 157–173 (2006) Reprints available directly from the publisher. Photocopying permitted by license only.

1065-5131/06/$35.00 © 2006 by Begell House, Inc. www.begellhouse.com Printed in the USA.

,QGXVWULDO 5HIULJHUDWLRQ DQG $PPRQLD (QKDQFHG +HDW 7UDQVIHU ZAHID AYUB Isotherm, Inc., Arlington, Texas, US

For the last two decades enhanced heat transfer equipment has been widely used by the halocarbon-based air conditioning industry. The persistent issue of ozone and global warming has led to interest in natural refrigerants, such as ammonia, which has played a prominent role in the refrigeration industry for years, particularly in the fields of food, beverage, and marine. The United States food industry has almost exclusively used ammonia as a refrigerant of choice thus far. Therefore, in view of the present situation as well as the future of the air-conditioning and refrigeration industries, it is important to study the role of enhanced heat transfer technology as applied to natural refrigerants. This paper presents an overview of the status of ammonia as a refrigerant of the past, present, and future and what has been done and what ought to be done in order to make ammonia equipment more safer, compact, and attractive so it can be readily accepted by the general public.

*

Corresponding author: Zahid Ayub, [email protected]

157

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

ZAHID AYUB

158

NOMENCLATURE Ax

flow area, bw

Greek symbols

b

flow channel gap, p – t

β

chevron angle, deg

cp

specific heat

λ

corrugation pitch

de

equivalent diameter, (2b/φ)

Γ

average mass flow rate (external) per tube

D

tube diameter

f

Fanning friction factor

∆p

pressure drop

g

gravitational acceleration

µ

dynamic viscosity

h

heat transfer coefficient

ν

kinematic viscosity

φ

ratio of developed length to protracted length

per unit length

hfg

latent heat

H

turbulator pitch at 180 deg twist

k

thermal conductivity

ϕ

non-dimensional heat flux

Lp

effective plate length

ρ

fluid density

n

bundle row number from top to bottom

Subscripts

N

total rows in the bundle

cr

critical

Nu

Nusselt number, hd/k

e

equivalent

pcr

critical pressure

f

liquid

pr

reduced pressure, ps/pcr

g

gas

ps

saturation pressure

ieff

inside effective

P

perimeter, 2φw

in

inlet

Pr

Prandtl number, µcp/k

o

outside

q′′

heat flux

p

particle

Re

Reynolds number, ρdeV/µ

r

reduced

t

plate thickness

s

saturation

Tcr

critical temperature

spn

single plain tube with no oil

Ts

saturation temperature

spo

single plain tube with oil

V

velocity

tp

two-phase

w

effective plate width

w

wall

w

mass concentration of lubricant

x

quality

(enlargement ratio)

INTRODUCTION Ammonia, used as an early refrigerant, is environmentally safe, low cost, and readily available. Unfortunately, this highly efficient refrigerant was pushed aside in favor of the recently discredited halocarbons. The depletion of the ozone layer and global warming has left a marked impression on the business of

air condition and refrigeration. Various international treaties and protocols have left the industry with few choices. First, chlorofluorocarbons (CFCs) were put on the hit list. The genuine belief was that the hydrochlorofluorocarbons (HCFCs) such as R-22 would stay with us for a while, but that is not the case. Europeans started the R-22 phase-out in the early 2000s by pro-

Journal of Enhanced Heat Transfer

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

INDUSTRIAL REFRIGERATION AND AMMONIA ENHANCED HEAT TRANSFER

hibiting its use if the motor power exceeded 150 kW. By 2010 no more virgin R-22 can be used in existing systems and after 2015 no recycled R-22 can be used in existing systems. This push has resulted in the total abandonment of R-22 in Western Europe. The most recent news is that the hydrofluorocarbons (HFC) are on their way out, too. Lawmakers in Europe are hinting to declare HFCs, such as R-134a, which took over as a replacement in centrifugal machines that used R12, as a banned item. This would certainly impact the thousands of machines presently operating with R134a. The United States has always pursued its own way without paying attention to others, but in today’s global economy it is not possible and unwise to ignore the strong ripple effect of such an attitude. Consequently the big chiller business will certainly be affected and therefore the market strategists in the US have to take a long hard look into this matter. AMMONIA: A VIABLE OPTION These fundamental issues and concerns bring us to the discussion of alternatives. It is apparent that with the advent of mechanical refrigeration in the late 1800s, human civilization changed forever. Without refrigeration, we would have not accomplished the major advances of the last century. This technological advance has directly contributed to humanity in the form of food preservation, indoor air quality and comfort control, gas liquefaction, food and beverage production, and electronics cooling. In order to maintain this improved standard of living, it is imperative to keep refrigeration equipment in running condition. According to an IIR position paper, 15% of the world power output is used to drive refrigerating and air conditioning systems. Power plants are a major source of global warming. If the power industry is slow or reluctant in controlling this menace, then we, the refrigeration equipment manufacturers and users, have the responsibility to devise systems that are more efficient. Note that the refrigeration sector employs approximately two million people on a global scale with annual sales of $200 billion in equipment, which in turn results in annual sales of $1200 billion in refrigerated food items. Presently, two different approaches are being discussed, i.e., direct or indirect cooling. With indirect cooling the use of ammonia does not become a serious issue since toxicity does not become a limiting factor. However, in direct applications the issue of

159

ammonia becomes a major point of discussion. In the United States all major food, meat, and beverage processing and storage facilities use ammonia exclusively as a refrigerant of choice. In recent years, the local jurisdiction has imposed new regulations on inventory control. Consequently, it introduced strict PSM (Process Safety Management) procedures in place at these facilities. In view of these developments, contractors and end users are more receptive to systems that would result in drastic reduction in the refrigerant inventories. Western Europe, on the other hand, is divided into two camps. The north has imposed minimal restrictions on natural refrigerants such as ammonia and hydrocarbons, whereas in the south there are strict safety regulations that have limited the use of ammonia. With the strong impetus towards elimination of halocarbons, it is possible that this southern approach may change in the near future. This is obviously a political issue, and the politicians can sort it out. As far as the refrigeration community is concerned, there is no doubt that ammonia has proved to be an excellent alternative with the following characteristics: 1) Excellent thermodynamic and thermophysical properties; 2) Environmentally friendly with Ozone Depletion Potential (ODP) and Global Warming Potential (GWP) of zero; 3) Pungent odor; self alarming refrigerant. The only drawback of ammonia is its toxicity, which can be overcome by the inherent pungent smell. In case of even a minor leak the strong odor of ammonia forces an individual to leave the affected area. AMMONIA AND TRADITIONAL HEAT TRANSFER EQUIPMENT Traditionally, chillers, condensers, oil coolers, and other heat transfer equipment for ammonia refrigeration systems have been constructed with plain surface tubes. This trend has been prevalent for two reasons. First, ammonia is not compatible with copper alloys, whereas most highly efficient heat transfer surfaces have been developed in copper and copper alloys. Hence, plain surface carbon steel tubes have been utilized in the bulk of equipment for ammonia applications. Second, because ammonia did not play a prominent role in the air-conditioning industry, big

Volume 13, Number 2, 2006

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

160

ZAHID AYUB

companies did not see any reason to invest in ammonia-related research and development. Flooded shell-and-tube evaporators for ammonia were, and are still, designed and fabricated with heavy wall carbon steel tubes. The most common tube has been 1.25 in. (31.75 mm) OD with 13 gage (2.4 mm) wall thickness. The same is true with shelland- tube condensers. This design criterion results in large size equipment, which in turn results in large ammonia inventory. Ammonia refrigeration systems primarily use as a lubricant in the compressors mineral oil, which is not miscible and is heavier than ammonia. This characteristic has played a favorable role in flooded and gravity feed systems, however, in direct expansion (DX) systems this aspect can create serious oil management problems. This is the fundamental reason why ammonia DX systems are not common in industry. With the advent of synthetic oils this issue has been addressed to some extent, which will be discussed later in this paper. Oil coolers and other smaller heat exchangers such as under-floor warming heaters use 0.75 in. (19 mm) OD plain surface carbon steel tubes. Almost all evaporative condensers use plain surface hot dipped galvanized carbon steel tubes, but stainless steel may be used for special applications. HEAT TRANSFER ENHANCEMENT This subject has gained some momentum in ammonia applications since the early 1990s. However, the progress has not been great for reasons that include those mentioned above. Very few companies in this business have taken the initiative to probe into the potential use of high-efficiency tubes. The entire blame cannot be put on the tube manufacturers’ shoulders. Because the ammonia refrigeration industry has worked with a "closed-mind" approach, it has been very difficult to implement change. Therefore, changing attitudes and convincing contractors and end-users can sometimes be a challenging task. The result is that there are very few heat exchanger manufacturers in the world who utilize enhanced surface tubes for ammonia applications. Ammonia-related enhanced heat transfer research has also been limited. In the past ten years, less than five research projects have been undertaken by ASHRAE. Similarly, only limited research activity has been reported in Europe and Asia. A literature search on the subject indicates less than fifty publications,

which is extremely low as compared to enhancement work related to halocarbons. RECENT DEVELOPMENTS As stated above there have been only a few cases where enhancement technology has been successfully applied to the ammonia equipment. The following is a discussion about successful use of enhanced surface techniques. Shell-and-Tube Ammonia Flooded Evaporators Enhanced surface tubes have been used successfully in various applications on a limited scale. In selecting enhancement for ammonia flooded evaporators, it is important to perform a preliminary screening of individual thermal resistances. This is critical because ammonia, unlike halocarbons, has high boiling heat transfer coefficients. Therefore, if the controlling side is the product side, it would merely be an unnecessary waste of capital to enhance the ammonia side. An ideal example is a case study of cooling calcium chloride brine, 29% by weight, in a flooded evaporator. Table 1 shows the design parameters and output data. The final selection for this application was based on the basis of ho/hieff close to one. Therefore, case C, with inside enhancement was selected as an optimized design. Enhancement was achieved via twisted tape turbulators with H/D = 4.1, tape width 0.5 in. (12.7 mm) and tape thickness 0.025 in. (0.635 mm). Selection D would have been satisfactory, too, but it would have resulted in 20% extra cost and 50% extra pumping power requirement. The unit was installed and has been operational for several years with no thermal or mechanical problems. Selection of a suitable design correlation can be difficult. Tube side should not be a problem since there are a wide variety of reliable correlations. It is the shell side that can pose a hindrance. A few correlations are available but there are wide discrepancies between them. The most recent extensive work was performed by Chyu et al. [2001] on evaporation of ammonia outside plain and enhanced surface tubes with miscible oil and by Zheng et al. [2001] on a plain tube with miscible oil. Zheng et al. [2001] showed that their plain tube data showed a wide variation when compared to the predictions of the Cooper [1984], Gorenflo et al. [1990], and Gorenflo [1993]

Journal of Enhanced Heat Transfer

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

INDUSTRIAL REFRIGERATION AND AMMONIA ENHANCED HEAT TRANSFER

161

TABLE 1. Comparative Case Study of Ammonia Flooded Evaporator Parameter

A

B

C

D

Shell OD, in. (mm)

16 (406.4)

16 (406.4)

16 (406.4)

16 (406.4)

Tube OD, in. (mm)

0.75 (19)

0.75 (19)

0.75 (19)

0.75 (19)

Tube wall, in. (mm)

0.065 (1.65)

0.049 (1.245)

0.065 (1.65)

0.049 (1.245)

2

2

2

2

–78%

–75%

0%

+6%

20.0 (1.38)

35.0 (2.41)

6.8 (0.47)

10.8 (0.74)

2.90

4.10

1.20

1.87

Passes Under/over capacity Pressure drop, psi (bar) ho/hieff

Design capacity: 487,200 Btu/hr (143 kW) Refrigerant: Ammonia @ –40oF (–40oC) saturated temperature Process fluid: 29% wt/wt CaCl2 brine Process flow: 450 gpm (28.4 l/s) Process inlet: –29.4oF (–34.1oC) Process outlet: –32oF (–35.6oC) A: No enhancement B: Outside enhancement only C: Inside enhancement only D: Both sides enhanced

FIGURE 1. Comparison of various correlations with Zheng et al. [2001] at Ts = 7.2oC.

Volume 13, Number 2, 2006

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

ZAHID AYUB

162

correlations. Comparison of their data with these correlations is shown in Figure 1, which indicates that Cooper underestimates the data by 35–50% at various heat fluxes and Gorenflo highly overestimates the data. However, Stephen and Abdelsalam [1980] correlated well with their data. The data were correlated for a single tube (–23.3oC < Ts < 7.2oC) using a simple dimensional equation similar to that of Stephen and Abdelsalam as follows: hspn = αspn q"0.55

Enhanced tube bundle with miscible oil and inlet quality effect Nu = e[8.156+0.199(N–n)] (q"D/kfTs)m

× (1 – ρg/ρf)–8.995xin – Pr–10.143 where m = 1.735 – 31.113pr + 137.492p2r – 0.023w – 16.179w2

(1)

where αspn = 0.291 + 3.90 × 10–3Ts + 4.75 × 10–4Ts2 + 1.84 × 10–5Ts3 + 1.97 × 10–7Ts4 A similar approach was adopted for ammonia/oil mixtures as follows: hspo = αspoβspoq"0.563

(4)

As mentioned earlier, most ammonia systems use mineral oil for compressor lubrication, which is inherently immiscible and heavier than ammonia. This characteristic makes oil management much easier and simpler for the majority of the tubes except for the bottom rows that are generally oil-free. Hence, the above equations with w = 0 can be used for the bottom rows.

(2) Enhancement Variation within a Bundle

where αspo = 0.260 + 2.76 × 10–3Ts 2

βspo = 1.10 – 6.18w + 53.2w

with h in kW/m2⋅C, q" in kW/m2, and Ts in oC. Their attempts to model plain and enhanced tube bundle results using the three prevailing techniques, i.e., superposition, enhancement, and asymptote were not successful for various reasons elaborated in detail in their research report [Chyu et al., 2001]. They adopted a more simplistic approach of correlating their data statistically for their 5 row × 3 column tube bundle, utilizing the Stephen and Abdelsalam [1980] approach for –23.3oC < Ts < 7.2oC as follows: Plain tube bundle with miscible oil and inlet quality effect Nu = e[4.342+0.244(N–n)] (q"D/kfTs)m (1 – ρg/ρf)–48.887xin Pr–3.944

(3)

where N is the total number of rows in the bundle, n is the row number with the top row being 1 and the bottom row being N. The exponent m is: m = 2.321 – 80.494 pr + 1086.842p2r + 4.436w – 29.437w2

In a United States patent application Ayub [2003] has disclosed a flooded evaporator with various types of enhanced tubes along the bundle height. It has been observed that enhanced-surface tubes cause high vapor generation which can become so intense that it causes a vapor-rich zone in the upper section of a tube bundle as schematically indicated in Figure 2. Higher void fraction is not desirable since it starves the tubes of liquid refrigerant and in turn affects the performance of the evaporator. In view of this behavior, Ayub proposed utilizing various kinds of tubes appropriately selected along the height of the tube bundle, with high efficiency tubes having strong nucleate boiling characteristics in the lower section, i.e., Section I as shown in Figure 3, followed by tubes with moderate nucleate boiling characteristics in Section II, still another suitable kind of tube in Section III and even plain tubes (if more tubes are required) in Section IV. This invention results in a lower cost by replacing the enhanced tubes in the top section with less expensive plain tubes. It also results in a highly optimized evaporator with no parasitic losses. In yet another embodiment, the selection of the top section tubes would depend on the type of process fluid being cooled in the tubes. If the process fluid has high viscosity, then

Journal of Enhanced Heat Transfer

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

INDUSTRIAL REFRIGERATION AND AMMONIA ENHANCED HEAT TRANSFER

VAPOR

LIQUID /VAPOR

FIGURE 2. Liquid/vapor distribution in a flooded evaporator tube bundle.

SECTION IV

SECTION III

SECTION II

SECTION I

FIGURE 3. Different types of tubes in different sections of a flooded tube bundle.

Volume 13, Number 2, 2006

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

163

ZAHID AYUB

164

TABLE 2. Existing Enhanced Tube Evaporator vs. Plain Tube Flooded Evaporator Characteristics

Existing

Plain tube flooded

38 (965)

54 (1372)

192 (4877)

168 (4267)

0.75 (19)

0.75 (19)

2

4

14.3 (0.99)

4 9.5 (0.66)

Price $ ( in the year 2001)

60,000

98,000

Ammonia charge, lbs (kg)

2000 (907)

4500 (2041)

Shell OD, in. (mm) Tube length, in. (mm) Tube OD, in. (mm) No of Passes Pressure drop, psi (bar)

Design capacity: Refrigerant: Process fluid: Process flow: Process inlet: Process outlet:

11,400,000 Btu/hr (3340 kW) Ammonia @ +14oF (–10oC) saturated suction temperature 25% wt/wt ethylene glycol brine 3500 gpm (221 l/s) +27oF (–2.78oC) +20oF (–6.67oC)

SECTION III

SECTION II

SECTION I

FIGURE 4. Two-pass tube sheet layout with three sections. Journal of Enhanced Heat Transfer

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

INDUSTRIAL REFRIGERATION AND AMMONIA ENHANCED HEAT TRANSFER

FIGURE 5a. Enhanced tubes in Section I

FIGURE 5b. Enhanced tubes in Section II.

FIGURE 5c. Plain surface tube with an internal turbulator in Section III.

Volume 13, Number 2, 2006

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

165

166

ZAHID AYUB

tubes with a plain surface on the outside and enhancement on the inside could be used. The tube distribution along the height of the bundle is dictated by the bundle depth, temperature range, approach temperature, the type of pass arrangement on the tube side, and the transport properties of the process fluid being cooled. One arrangement could be to match the number of different tubes with the number of passes on the tube side, e.g., for a two pass arrangement, two different kinds of tubes could be used. If the bundle height is large and has two or one passes on the tube side, the bundle can have more than two kinds of tubes depending upon the design parameters, as indicated in the following case study. The design parameters and physical characteristics are shown in Table 2. Also shown is an evaporator with same size tubes but no enhancement. Due to a length restriction the maximum allowable tube length was limited to 16 ft (4880 mm). This restriction and the pressure drop limitation of less than 15 psi (1 bar) would have forced the design to 54 in. (1372 mm) shell diameter with 2600 plain surface tubes. As indicated in Table 2, this option would have had substantial effect on the cost and the ammonia charge. Maintenance and other running costs would have been additional expenses during the service life of this evaporator. During the design phase, several enhanced surface techniques were evaluated and it was finally decided to adopt enhanced tubes of various types. The depth of the bundle, two-pass arrangement (pressure drop limitation), LMTD, and the approach temperature directed the design towards a bundle with three different types of tubes. Figure 4 shows the tube hole layout with three designated sections each having a similar type of tubes. The modeling process is cumbersome, proprietary and cannot be disclosed. There is no model available in the open literature. One simple approach would be to apply the available methods to each section separately and then lump them in series. But again it would be an approximate method since there are so many parameters that have to be accounted for. This is especially true in a case such as this where the number of passes is less than the number of enhancements, as each section will encounter nonuniform heat flux. As shown in Figure 4 the sections are designated as Section I (lower), Section II (middle) and Section III (upper). Fourteen rows of the lower section had highly structured outside surface tubes (quantity 393) with strong nucleate boiling char-

acteristics and internal grooves as shown in Figure 5a. The middle section had the bulk of the tubes (quantity 649; 17 rows) with slightly wider gap structure on the outside in order to overcome strong convective effects and internal enhancement similar to the lower section tubes. This type of tube has shown good nucleate boiling behavior in the presence of strong convective forces (Fig. 5b). The decision to select the appropriate tube in the upper section took the bulk of modeling and design time. Various calculations showed that the use of tubes similar to those in the lower two sections would have aggravated the heat transfer due to the vapor blanketing phenomenon. Plain surface tubes could have been used, but the tube side had glycol, which was in the final pass and would have cooled down, therefore, resulting in higher viscosity and further aggravating the problem. The calculations indicated that plain tubes would not have achieved the desired goal. Tubes with only internal enhancement in carbon steel or stainless steel are not readily available. Hence, it was decided to use twisted-tape inserts in plain tubes. After careful evaluation and modeling, a stainless steel tape with H/D = 5.5 (180 deg turn), width 0.5 in. (12.7 mm) and thickness 0.02 in. (0.51 mm) was selected (Fig. 5c). The clearance between the tape and the tube inside diameter was 0.06 in. (1.52 mm). It is worth noting that majority of research work in the open literature insists on snug fit, however, this approach is only possible on the laboratory level and is not practical from the manufacturing and maintenance point of view. In fact, a nonsnug fit may have a positive effect on the hydraulics as shown by Ayub and Al-Fahed [1993]. Shell-and-Tube Ammonia Spray Evaporators Spray evaporators have been successfully used in the winery and poultry industries for decades. This is one other area where enhancement could be applied with a greater success rate. In fact, spray evaporators are the ones that need enhancement because of the nature of the operating parameters. These chillers are used to cool fluids as close to the freezing point as possible without the danger of freeze-up. To avoid this problem, a system has to be designed that can operate at suction temperatures close to the freezing point of the process fluid and still be efficient under close approach temperatures. Obviously, plain surface tubes have a limit on approach temperature; hence, the best

Journal of Enhanced Heat Transfer

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

INDUSTRIAL REFRIGERATION AND AMMONIA ENHANCED HEAT TRANSFER

option is to utilize enhanced surface tubes that require low wall superheat. Extensive research was undertaken through ASHRAE Project 725-RP on the subject. The effects of heat flux, saturation temperature, spray flow rate, nozzle height and nozzle type were thoroughly investigated. Several papers were published on various aspects of spray evaporation fundamentals [Zeng et al., 1994, 1995], effect of spray distribution [Zeng et al., 1995], and single tube and bundle effect [Zeng et al., 1997, 2001a,b,c]. Experiments were conducted on three different types of tubes: plain, corrugated, and low fin (40 fpi). Their results indicate that the heat flux and saturation temperature effects on the heat transfer coefficient are similar to those in pool boiling. The bundle effect is less pronounced at low saturation temperatures. The top row tubes exhibited relatively higher heat transfer coefficient compared to the lower row tubes at same heat flux and refrigerant flow rate. This was more pronounced in the low-fin tube bundle due to the fact that the impingement effect and more vigorous nucleate boiling helped the tubes in the top row; however, the lower row tubes probably suffered ammonia maldistribution due to the physical presence of adjacent radial fins, thus hampering the lateral flow. The heat transfer enhancement was high at higher saturation temperatures. In general, spray evaporation coefficients were very high with low-fin tubes and as much as 65% above those for plain tubes in pool boiling. Visual observations indicated that there was mist within the bundle, especially at high heat fluxes, which probably helped in wetting the entire fin tube bundle but still did not offset the effect of maldistribution in the bottom rows. The effects of nozzle type and the distance between the bundle and the spray nozzle were weak. Since none of the previous falling film correlations were applicable due to the mechanics of the flow and the wide range of saturation pressures, Zeng et al. [1998] adopted the modified Chun and Seban [1971] function but substituted a nondimensional heat flux, ϕ, for a dimensional heat flux, q", as proposed by Parken et al. [1990]. To account for the effect of saturation temperature, a reduced pressure ratio was also added to the correlation as follows: Nu = 0.0568Re–0.0058Pr0.193pr0.323ϕ1.034 (–23oC < Ts < 10oC)

(5)

167

where Nu = h/k(υ2/g)1/3 Re = 2Γf/µ ϕ = q"D/(Tcr – Ts)k This correlation is data-specific and is derived from single tube data but could be used for bundle design until a comprehensive bundle correlation is published in the future. Zeng et al. [1997, 2001a,b,c] have presented similar correlations for plain tube square and triangular tube bundles. A low charge ammonia spray evaporator was proposed for a process cooling application at a chemical plant in Cheswold, Delaware. This plant requires chilled water for various processes and during evening time the water is diverted to a 660,000 gallons (2500 m3) tank, an integral part of a Thermal Storage System (TES) with the goal to pull down the tank temperature to 38oF (3.3oC). With the existing chillers, two (2) shell-and-tube heat exchangers flooded with ammonia were tied to 500 hp screw compressors and one R-11 centrifugal compressor. The lowest water temperature achieved was 40oF (4.4oC). Therefore, the minimum storage tank temperature attained was 42–45oF (5.6–7.2oC). A comparative analysis between a spray evaporator and a flooded evaporator was presented, as shown in Table 3. The spray evaporator has certain inherent characteristics that make it an equipment of choice for such an application. First, it has higher heat transfer coefficient. Hence, for the same size and load the operational approach temperature is lower. Second, due to these characteristics the refrigerant suction temperature can be raised, therefore, resulting in compressor capacity enhancement. Third, due to low approach temperature the suction temperature can be maintained close to or higher than the freezing point of the process fluid. Fourth, the refrigerant charge is an order of magnitude lower than the same capacity flooded evaporator, hence, making it an environmentally attractive option (Table 3). Fifth, because of the low charge and the shell being devoid of liquid refrigerant it is well guarded against freezing in the event of process pumps or refrigerant control failure. Sixth, there is no hydrostatic head penalty, hence, the height has no effect on the LMTD.

Volume 13, Number 2, 2006

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

ZAHID AYUB

168

TABLE 3. Comparison with a Same Capacity Flooded Evaporator Characteristics

Existing spray unit

Conventional flooded unit

Shell dia, in. (mm)

48 (1219)

60 (1524)

Tube length, ft (mm)

17 (5182)

24 (7315)

Tube OD, in. (mm)

0.75 (19)

1.25 (32)

2

8

2 450 (204)

9000 (4082)

No. of passes Ammonia charge, lbs (kg)

In order to further improve the efficiency of the evaporator, doubly enhanced surface carbon steel tubes were incorporated. Spray evaporators work effectively if the feed ratio is higher than 2:1 so that no section of the bundle is starved of liquid refrigerant. An appropriate refrigerant distribution system, as proposed by Zeng et al. [1995], was designed. The outside surface of the tubes consisted of low integral fins. The unit (Fig. 6) has been operational for almost two and a half years. It has exceeded all expectations under different seasonal conditions. Water outlet temperature as low as 33oF (0.6oC) has been recorded, and there are confirmed reports of storage tank temperatures as low as 36–37oF (2.2–2.8oC). It was recently reported that the existing two flooded chillers were frozen and were out of commission due to damaged tubes. Since then the spray evaporator has been handling the entire load in the plant. Early this year a second duplicate spray evaporator was fabricated and supplied to the this plant and is in the process of integration. All physical features are similar to the existing unit except two different types of enhanced tubes were used in the newer version. The top half has a tube geometry similar to the existing unit, however, in view of the maldistribution concerns. The lower half has a tube structure that is expected to help the longitudinal flow of refrigerant and would, therefore, minimize any chance of dry spots in the lower section of the bundle.

Shell-and-Tube Ammonia Direct Expansion (DX) Evaporators In the past due to immiscibility of mineral oil, DX type evaporators were generally discouraged. This thinking, however, changed with the push towards the concept of low charge and natural refrigerants. New synthetic oils have been developed for ammonia applications. One of the early works on the subject was done by Burke and Kruse [1993]. They recommended polyalkylene glycols (PAG) with 1:1 ratio of ethylene oxide and propylene oxide. Most correlations developed for in-tube flow boiling cannot be confidently applied to ammonia. The reason is simple. There is not enough data available even with smooth tubes. Kelly et al. [2002a] carried out extensive experimental work on in-tube evaporation of pure ammonia in smooth and microfin tubes over a wide range of saturation temperature, mass flux, heat flux, and inlet quality. None of the available correlations satisfactorily predicted their data over the entire range of conditions. Table 4 shows the percent deviation of their data from the predictions of various correlations. They found that microfin tubes showed heat transfer enhancement that was more prominent at the low mass flux where the flow was prone to stratification. The microfins helped in wetting the entire tube. Microfin tubes were thin wall 0.5 in. (12.7 mm) OD, 60 fins with 17 deg helix angle. This helix angle

TABLE 4. Prediction vs. Kelly et Al. [2002] Data Chaddock–

Gungor–Winteron

Kattan et al.

Wattelet et al.

Brunemann [1967]

[1986]

[1998]

[1994]

77.0

78.1

73.4

53.3

59.2

50.2

43.9

29.8

41.7

47.4

36.3

156.9

148.9

166.7

142.6

67.7

113.0

Quality range, %

Shah [1982]

Kandlikar [1987]

0–100

83.8

0–75 75–100

Journal of Enhanced Heat Transfer

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

INDUSTRIAL REFRIGERATION AND AMMONIA ENHANCED HEAT TRANSFER

169

FIGURE 6. Limited charge ammonia spray evaporator.

is close to the 18 deg used in most microfin tubes. Oh and Bergles [2002] conducted heat transfer measurements and flow visualization with three different helix angle microfin tubes and R-134a as a working fluid. Their study of horizontal tubes showed that the 18 deg helix angle grooves effectively lifted the fluid at low mass flux, enhancing the heat transfer. At high mass flux the tube with 6 helix showed higher enhancement. It is noted, however, that all conclusions regarding the effect of helix angle have been based on halocarbon results. No similar research has been performed with ammonia as a working fluid. Because of the high latent heat transfer of ammonia, the mass fluxes encountered in the evaporators are generally lower than those in an equivalent halocarbon evaporator. In view of this characteristic it is logical to use tubes with larger helix angles in accordance with the Oh and Bergles result. Kelly et al. [2002b] present results for pressure drop of ammonia, which show a higher penalty for microfin tubes. The effect was more predominant with a decrease in saturation temperature. Correlations by Hughmark [1963] and Friedel [1979] well predicted the data with smooth tubes. It is important to note that the surface tension of ammonia is approximately twice that of R-22 or R134a. This characteristic may result in a different fin

height for an ammonia-optimized evaporator. In other words, a given microfin geometry may not exhibit the same enhancement ratio with ammonia as with halocarbons. It is probable that due to the higher surface tension the micro fins could be totally or partially submerged in the liquid, therefore nullifying the fin effect. In the past ten years, the author has fabricated and installed several DX evaporators with internal grooves (8, 16, and 33) with various helix angles, both in 0.62 in. (15.9 mm) OD and 0.7 in. (19 mm) OD carbon steel tubes. No capacity or oil logging problems have ever been reported. Due to the low mass flow per unit load of refrigeration, it is critical to design adequate ammonia circuiting. The first pass should have the least number of tubes with progressive increase along the flow path. It would yet be desirable to incorporate a tube bundle with various number and size tubes in a shell to accommodate the constantly expanding refrigerant along the flow path as shown in Figure 7. This would help avoid large pressure drops inherent to heat exchangers with equal number of tubes in each pass. The tubing network would be less expensive as it would require fewer holes of larger diameter to drill and, therefore, less material and labor cost. An added advantage of various size tubes in a shell-and-tube unit would be an increase in a shell side heat transfer coefficient due to

Volume 13, Number 2, 2006

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

ZAHID AYUB

170

FIGURE 7. Different tube size bundles for DX evaporator.

a tortuous flow path as a result of higher turbulence. It is expected that this increase in turbulence will also result in a "self-cleaning" characteristics and, therefore, be less prone to fouling on the shell side.

where

Ammonia Plate Evaporators

The values of kf, de, hfg, and Lp are in US units of Btu/hr⋅ft⋅oF, ft, Btu/lb, and ft, respectively. The correlations for the Fanning friction factor within the plates are formulated as if the entire flow were saturated vapor. The correlation is accordingly adjusted for the chevron angle, hence, resulting in a generalized correlation that could be applied to any type of commercially available plate.

Several types of plate exchangers are being used as evaporators in the refrigeration industry, e.g., conventional gasket plate and frame, compact brazed, semiwelded and shell-and-plate. The last three types are most common in the industry today and they all have similar geometric characteristics as shown in Figure 8. Design information and correlations are either lacking or insufficient in the open literature. Recently, Ayub [2003] presented a comprehensive literature search on the subject with presentation of new correlations for flooded, gravity feed and DX evaporators. The correlations are based upon a decade of design, field experience and after-installation data collection on ammonia and R-22 direct expansion and flooded evaporators installed in North America. A dimensional correlation incorporates the principle of corresponding states and is applicable to various chevron angles as follows: htp = C(kf/de) [Re2f hfg/Lp]0.4124(ps/pcr)0.12(65/β)0.35 [Btu/hr⋅ft2⋅F]

(6)

C = 0.1121 for flooded and thermo-syphon C = 0.0675 for direct expansion (DX).

f = (n/Rem)(–1.89 + 6.56R – 3.69R2) 30 ≤ β ≤ 65

(7)

where R = (30/β)

m

n

0.137

2.99

Re ≤ 4000

0.172

2.99

4000 < Re ≤ 8000

0.161

3.15

8000 < Re ≤ 16,000

0.195

2.99

Re > 16,000

The pressure drop within the port holes is correlated as follows, with the assumption of the entire flow being saturated vapor: Journal of Enhanced Heat Transfer

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

171

INDUSTRIAL REFRIGERATION AND AMMONIA ENHANCED HEAT TRANSFER

Cross-section normal to direction of troughs

Developed length

Protracted length %

Lp

w b

t

p

)

β = Chevron angle φ = Enlargement factor = Developed length / Protracted length λ = Corrugation pitch FIGURE 8. Plate parameters for a typical plate exchanger.

∆pport = 0.0076ρV2/2g

(8)

This equation accounts for pressure drop in both inlet and outlet refrigerant ports and gives the pressure drop in units of lb/in.2 with input for ρ in lb/ft3, V in ft/s, and g in ft/s2. This area needs in-depth research not only for ammonia but other refrigerants, too. There is limited amount of information available. Manufacturers have developed their own data bases, but they are considered proprietary information. Ammonia Condensers Plain surface shell-and-tube condensers have been the industry standard. The reason for this is twofold: first, there has been lack of availability of enhanced surface tubes for condenser applications and second, since ammonia has high condensing coefficients, further efficiency was not a pressing issue. It is also possible that higher than halocarbon surface tension for ammonia could have caused unwillingness on the part of designers to use low-fin or other types of enhanced surface

tubes. However, enhanced tubes could be used effectively to reduce the size of ammonia condensers. The author has had success with low profile, low-fin tubes. Also several other types of enhanced surface tubes in materials such as carbon steel, stainless steel and titanium have been developed and used by the author. These tubes could be doubly enhanced to boost up the overall heat transfer coefficient. The author has also employed different types of tubes in a single condenser. In order to reduce the discharge superheat more effectively, low-fin tubes are suggested in the top section followed by plain tube or structured surface tubes. To justify this idea a simple temperature measurement was carried out on an ammonia condenser installed on a fishing vessel. The temperature of the surface skin was recorded with a hand-held infrared thermometer. The readings showed that the superheated gas plays a vital role in the upper section and the quicker this heat dissipates, the more efficient would be the condenser, in turn achieving higher subcooling. During the last decade, nickel brazed compact exchangers and semi-welded plate and frame exchangers

Volume 13, Number 2, 2006

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

172

ZAHID AYUB

have been used for ammonia applications. However, the open literature lacks design correlations for ammonia condensation. Recently, there has been work undertaken by researchers on mini-channel geometries. Results are still awaited. An industry-based consortium at the University of Illinois is making headway in the preliminary research on mini-channel tubes by reducing the refrigerant charge to 2.5 ounces (70 grams) of ammonia per ton (0.28 kW) of evaporator capacity considerably lower than the 12.5 ounces (354 grams) per ton used in current air-cooled ammonia chillers. According to Professor Hrnjak, further charge reduction would be possible through better design of the condenser headers and optimization of the heat exchangers. Yet another interesting project is underway through ASHRAE sponsored research investigating ammonia condensation in plain and enhanced tubes with oil effects. CONCLUSIONS AND RECOMMENDATIONS

emerging enhancement technology. Some of the latest work on the evaporators is summarized and design correlations are presented for shell-and-tube and plate type evaporators. Furthermore, several case studies are presented with emphasis on new concepts that are at the patent disclosure level. There are still unknowns that need to be explored by researchers. There is a need for reliable correlations for in-tube boiling with pure ammonia and ammonia/oil mixtures in enhanced tubes. Therefore, specific geometry enhanced tubes have to be initially produced by the tube manufacturers. This paper clearly indicates that present geometries designed for halocarbons are not optimized for ammonia applications. New alloys that are less expensive, easily machined and compatible with ammonia and/or ammonia–oil mixture have to be developed. Special, but economical brazing techniques have to be developed for plate exchangers to withstand high pressures for ammonia/carbon dioxide cascade condensers.

The paper presents a case for ammonia as a viable refrigerant that has a wider role to play in the future. The traditional technology is compared with the

REFERENCES Ayub, Z. H. and Al-Fahed, S. F. (1993) The Effect of Gap Width between Horizontal Tube and Twisted Tape on the Pressure Drop in Turbulent Water Flow, Int. J. Heat Fluid Flow, Vol. 14, No. 1, pp. 64–67. Ayub, Z. H. (2003) Plate Heat Exchanger Literature Survey and New Heat Transfer and Pressure Drop Correlations for Refrigerant Evaporators, Heat Transfer Eng., Vol. 24, No. 5, pp. 3–16. Ayub, Z. H. (2003) Flooded Evaporator with Various Kinds of Tubes, United States Patent 7,073,572. Burke, M. and Kruse, H. (1993) Solubility and Viscosity of New Oil/Ammonia-Systems, Proc. IIR Conf. on Energy Efficiency in Refrigeration, Global Warming Impact, pp. 133–139. Chaddock, J. B. and Brunemann, H. (1967) Forced Convection Boiling of Refrigerants in Horizontal Tubes — Phase 3, Report HL-113, Durham, North Carolina, Duke University School of Engineering.

Chun, K. R. and Seban, R. A. (1971) Heat Transfer to Evaporating Liquid Films, J. Heat Transfer, Vol. 93, pp. 391–396. Chyu, M.-C., Zheng, J., and Jin, G. (2001) Evaporation of Ammonia Outside Smooth and Enhanced Tubes with Miscible Oil, Final Report-ASHRAE RP-977. Cooper, M. G. (1984) Saturation Nucleate Boiling — A Simple Correlation, Chem. Eng. Symp. Ser. Vol. 86, No. 2, p. 785. Freidel, L. (1979) Improved Friction Pressure Drop Correlations for Horizontal and Vertical TwoPhase Pipe Flow. Presented at the European Two-Phase Flow Group Meeting, Ispra, Italy, Paper E2. Gorenflo, D., Sokol, P., and Caplanis, S. (1990) Pool Boiling Heat Transfer from Single Plain Tubes to Various Hydrocarbons, Int. J. Refrigeration, Vol. 13, pp. 286–292. Gorenflo, D. (1993) Pool Boiling, VDI-Verlag, Duesseldorf. Journal of Enhanced Heat Transfer

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

INDUSTRIAL REFRIGERATION AND AMMONIA ENHANCED HEAT TRANSFER

Gungor, K. E. and Winterton, R. H. S. (1986) A General Correlation for Flow Boiling in Tubes and Annuli, Int. J. Heat Mass Transfer, Vol. 29, pp. 351–358. Hughmark, G. A. (1963) Pressure Drop in Horizontal and Vertical Co-Current Gas-Liquid Flow, I&EC Fundamentals, Vol. 2, No. 4, pp. 315–321. Kandlikar, S. S. (1987) A General Correlation for Saturated Two-Phase Flow and Boiling Heat Transfer inside Horizontal and Vertical Tubes, in Boiling and Condensation in Heat Transfer Equipment, ASME Winter Annual Meeting, Boston, pp. 9–19. Kattan, N., Thome, J. R., and Favrat, D. (1998) Flow Boiling in Horizontal Tubes: Part 3 — Development of New Heat Transfer Model Based on Flow Patterns, J. Heat Transfer, Vol. 120, No. 1, pp. 140–147. Kelly, J. E., Eckels, S. J., and Fenton, D. (2002a) An Experimental Investigation of the In-Tube Evaporation of Pure Ammonia in a Smooth and a Microfin Tube, Part I: Heat Transfer (866-RP), Int. J. Heating, Ventilating, Air-Conditioning Refrigerating Res., Vol. 8, No. 3, pp. 239–255. Kelly, J. E., Eckels, S. J., and Fenton, D. (2002b) An Experimental Investigation of the In-Tube Evaporation of Pure Ammonia in a Smooth and a Microfin Tube, Part II: Pressure Drop (866RP), Int. J. Heating, Ventilating, Air-Conditioning Refrigerating Res., Vol. 8, No. 3, pp. 256–276. Oh, S.-Y. and Bergles, A. E. (2002) Visualization of the Effects of Spiral Angle on the Enhancement of In-Tube Flow Boiling in Microfin Tubes, ASHRAE Trans., Vol. 108, pt. 2, pp. 509–516. Parken, W. H., Fletcher, L. S., Sernas, V., and Han, J. C. (1990) Heat Transfer through Falling Film Evaporation and Boiling on Horizontal Tubes, J. Heat Transfer, Vol. 112, pp. 744–750. Shah, M. M. (1982) Chart Correlation for Saturated Boiling Heat Transfer: Equations and Further, ASHRAE Trans., Vol. 88, pt. 1, pp. 185–196. Stephan, K. and Abdelsalam, M. (1980) Heat Transfer Correlations for Natural Convection Boiling, Int. J. Heat Mass Transfer, Vol. 23, pp. 73–87.

173

Wattelet, J. P., Chato, J. C., Souza, A. L., and Christoffersen, B. R. (1994) Evaporative Characteristics of R-12, R-134a, and MP-39 at Low Mass Fluxes, ASHRAE Trans., Vol. 100, pt. 1, pp. 603–615. Zeng, X., Chyu, M.-C., and Ayub, Z. H. (1994) Characteristic Study of Sprayed Fluid Flow in a Tube Bundle, ASHRAE Trans., Vol. 100, pt. 1, pp. 63–72. Zeng, X., Chyu, M.-C., and Ayub, Z. H. (1995) Evaporation Heat Transfer Performance of Nozzle-Sprayed Ammonia on a Horizontal Tube, ASHRAE Trans., Vol. 101, pt. 1, pp. 136–149. Zeng, X., Chyu, M.-C., and Ayub, Z. H. (1995) Nozzle-Sprayed Flow Rate Distribution on Horizontal Tube Bundle, ASHRAE Trans., Vol. 101, pt. 2, pp. 443–353. Zeng, X., Chyu, M.-C., and Ayub, Z. H. (1997) Performance of Nozzle-Sprayed Ammonia Evaporator with Square Pitch Plain Tube Bundle, ASHRAE Trans., Vol. 103, pt. 2, pp. 68–81. Zeng, X., Chyu, M.-C., and Ayub, Z. H. (1998) Ammonia Spray Evaporation Heat Transfer Performance of Single Low-Fin and Corrugated Tubes, ASHRAE Trans., Vol. 104, pt. 1, pp. 185–196. Zeng, X., Chyu, M.-C., and Ayub, Z. H. (2001a) An Experimental Study of Spray Evaporation of Ammonia in a Square-Pitch, Low-Fin Tube Bundle, Int. J. Heat Exchangers, Vol. II, No. 2, pp. 129–149. Zeng, X., Chyu, M.-C., and Ayub, Z. H. (2001b) Experimental Investigation on Ammonia Spray Evaporator with Triangular-Pitch Bundle, Part I: Tube Bundle Effect, Int. J. Heat Mass Transfer, Vol. 44, No. 11, pp. 2081–2092. Zeng, X., Chyu, M.-C., and Ayub, Z. H. (2001c) Experimental Investigation on Ammonia Spray Evaporator with Triangular-Pitch Bundle, Part II: Evaporator Performance, Int. J. Heat Mass Transfer, Vol. 44, No. 12, pp. 2299–2310. Zheng, J. X., Jin, G.P., Chyu, M.-C., and Ayub, Z. H. (2001) Flooded Boiling of Ammonia with Miscible Oil Outside a Horizontal Plain Tube, Int. J. Heating, Ventilating, Air-Conditioning, Refrigerating Res., Vol. 7, No. 2, pp. 185–204.

Volume 13, Number 2, 2006

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01)

Begell House Inc., http://begellhouse.com Downloaded 2008-12-7 from IP 205.188.117.70 by Dr. Zahid Ayub (zayub01) View publication stats