🔥
HeatXpert Pro v7.0
Heat Exchanger Design & Analysis
🏠 Home
1

Select HX Type

Choose from Shell & Tube, Plate, Air Cooled, or Double Pipe

2

Enter Parameters

Set temperatures, flow rates, fluid types, and geometry

3

Calculate & Review

Get duty, area, U, NTU, pressure drop, and design recommendations

4

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⚙️
Shell & Tube — Design Inputs
⚙ Configuration
🔴 Hot Side (Shell)
°C
°C
Mass flow in kg/h
bar a
Shell-side operating pressure (absolute)
🔵 Cold Side (Tube)
✅ Mode: Know Flow Rate — enter cold flow rate; outlet temp will be auto-calculated from Qhot = Qcold.
°C
°C
Mass flow in kg/h
bar a
Tube-side operating pressure (absolute)
🔧 Tube Specifications
⚙ Tube ID and pass count are auto-calculated from flow rate + target velocity + tube length. Override any field manually — changes trigger instant recalculation.
mm
mm
m
—
—
—
🎯 Tube-Side Velocity Control
▸ The tool calculates number of tube passes to meet your target velocity. Recommended: 1–2 m/s (liquids) · 10–25 m/s (gases/steam). Gas density is corrected for actual T & P.
m/s
tubes
✓ Target: 1.5 m/s → auto-calculating required tube passes
🛡️ Baffle & Shell
%
—
bar
bar
bar
⚠
📊

Ready to Design

Enter parameters and click Calculate

Calculating heat transfer coefficients…

🔲
Plate HX — Design Inputs
🔴 Hot Side
bar
°C
°C
🔵 Cold Side
bar
°C
°C
kg/h
🔧 Plate Specifications
mm
°
mm
mm
mm
—
m²K/W
bar
bar
⚠
📋

Plate HX Results

Enter parameters and calculate

Calculating plate design…

💨
Air Cooled HX — Design Inputs
🔴 Process Side (Hot)
°C
°C
kg/h
🌬 Air Side
°C
°C
fins/m
🏗️ Bay Configuration
rows
m
⚠
🌡️

Air Cooler Results

Enter parameters and calculate

Calculating air cooler…

🌬️
Fin-Fan HX — Design Inputs
🔴 Tubeside (Process) Fluid
bar
°C
°C
kg/h
m²K/W
bar
—
🌬️ Airside Conditions
°C
°C
bar
m²K/W
🔧 Tube Geometry
mm
mm
mm
mm
mm
rows
—
tubes
🌀 Fin Geometry
fins/m
mm
mm
mm
🏗️ Unit & Bundle Geometry
—
—
mm
🔄 Fan Geometry
fans
mm
%
kW
mm
⚠
🌬️

Fin-Fan HX Results

Enter parameters and click Calculate to generate the full engineering datasheet

Calculating fin-fan design…

↔️
Double Pipe — Design Inputs
🔴 Inner Pipe (Hot)
bar
°C
°C
🔵 Annulus (Cold)
bar
°C
°C
kg/h
🔧 Pipe Dimensions
mm
mm
mm
m
m²K/W
bar
bar
⚠
📐

Double Pipe Results

Enter parameters and calculate

Calculating…

📐
LMTD & NTU Calculator
🌡 Temperature Inputs
°C
°C
°C
°C
⚡ Heat Capacity Rates (optional)
kW/K
kW/K
W/K
⚠
📐

LMTD / NTU Results

Enter temperatures and click Calculate

🎯
Smart HX Type Selector

Answer the questions below to get a recommended heat exchanger type for your application.

🔥
Complete Field Reference

Heat Exchanger Engineering
Design Guide

Everything a student or practising engineer needs to understand the theory, select the right equipment, work through a calculation, and validate the result — from first principles to TEMA code compliance.

LMTD & NTU Bell-Delaware Gnielinski TEMA Standards Fouling & ΔP
9
Sections
40+
Formulas
5
HX Types
TEMA
Code Ref.
🧱
Section 1 — Heat Transfer Fundamentals

Heat exchangers transfer thermal energy from a hot fluid to a cold fluid through a separating wall, without the fluids mixing. All designs are governed by the same three equations: the energy balance, the rate equation, and the design equation. Everything else is geometry and material science.

⚡ The Three Governing Equations
1. Energy Balance — heat given up = heat received
Q = ṁ_h · c_p,h · (T_h,in − T_h,out)
Q = ṁ_c · c_p,c · (T_c,out − T_c,in)
2. Rate Equation — how fast heat moves through area A
Q = U · A · F · ΔT_lm
3. Design Equation — size the exchanger
A_required = Q / (U · F · ΔT_lm)
These three, combined with an assumed or calculated U, give you every number needed to specify an exchanger.
🔗 Thermal Resistance Network
Heat flows through resistances in series — exactly like electrical resistors. The overall U is the reciprocal of the sum of all resistances:
1/U_o = 1/h_o + R_fo + (A_o·ln(r_o/r_i))/(2πkL) + R_fi·(A_o/A_i) + (A_o/A_i)·(1/h_i)
Where: h_o = shell-side film coeff, h_i = tube-side film coeff, R_fo / R_fi = fouling resistances (outer/inner), k = wall conductivity.

The dominant resistance controls U. In water–water service, fouling typically dominates. In gas–liquid service, the gas-side film always dominates — improving the liquid side is wasted effort.
📊 Heat Capacity Rate (C)
The heat capacity rate C = ṁ · c_p [kW/K] tells you how much temperature change each stream experiences per unit of heat transferred. It controls the temperature profiles.
C_h = ṁ_h · c_p,h    C_c = ṁ_c · c_p,c
C_min = min(C_h, C_c)   C_max = max(C_h, C_c)
C_r = C_min / C_max (heat capacity ratio)
Key insight: The stream with C_min always undergoes the larger temperature change. If C_h ≪ C_c, the hot stream cools dramatically and the cold stream barely warms — this is the typical situation for steam condensers (C_h → 0 as latent heat dominates).
🌡 Flow Arrangements Compared
Countercurrent (most efficient): Streams flow in opposite directions. The cold outlet can approach the hot inlet temperature. Maximum possible ΔT_lm for given terminal temperatures.

Parallel flow (least efficient): Streams enter same end. Cold outlet can never exceed the equilibrium temperature. LMTD always lower than countercurrent for same temperatures.

Crossflow: One stream flows perpendicular to the other. Common in air coolers and fin-fan units. F-factor (0.75–0.97) corrects the countercurrent LMTD.
ΔT_lm,counter > ΔT_lm,cross > ΔT_lm,parallel
For a given duty, countercurrent requires the smallest area. Always prefer countercurrent unless process constraints prevent it.
📋 Typical Overall Heat Transfer Coefficients U (W/m²·K)
ServiceShell & TubePlate HXAir CoolerNotes
Water – Water (clean)800–2 0003 000–7 000—Plate HX excels due to high turbulence
Steam condenser (shell)1 500–6 0002 000–5 000—Condensation is very high h; tube-side controls
Water – Light oil150–500400–1 200—Oil viscosity limits shell-side h
Water – Heavy oil / residue50–200Not recommended—High fouling, low h; S&T only
Gas – Liquid (high-P gas)100–350——Gas side always limiting; fins help
Gas – Gas (no fins)15–60——Both sides poor; use plate-fin for compact duty
Air cooling (process fluid)——40–120Air h ≈ 30–80 W/m²K; fins increase effective area 10–20×
Reboiler / vaporiser500–2 5001 000–4 000—Nucleate boiling greatly enhances h
Crude oil train100–400——Heavy fouling; design with 0.0005 m²K/W each side
📐
Section 2 — LMTD & NTU-Effectiveness Methods

There are two equivalent design methods. LMTD is best when all four terminal temperatures are known (rating / checking an existing exchanger). NTU-ε is best when outlet temperatures are unknown (design from duty and inlet temperatures). Both methods must give the same answer.

📐 LMTD Method — Step by Step
Step 1 — Compute terminal differences
Countercurrent: ΔT_1 = T_h,in − T_c,out    ΔT_2 = T_h,out − T_c,in
Parallel flow: ΔT_1 = T_h,in − T_c,in    ΔT_2 = T_h,out − T_c,out
Step 2 — Log-mean
ΔT_lm = (ΔT_1 − ΔT_2) / ln(ΔT_1 / ΔT_2)
Special case ΔT_1 = ΔT_2: ΔT_lm = ΔT_1
Step 3 — Correction factor F (for multi-pass or crossflow shells)
Q = U · A · F · ΔT_lm    F ≤ 1.0
Rule: F < 0.75 → increase shell passes or switch to counter-current
When to avoid LMTD: When only inlet temperatures are known, iteration is required — use NTU-ε instead.
📈 NTU-Effectiveness Method
Number of Transfer Units — dimensionless measure of exchanger thermal "size":
NTU = U · A / C_min
Effectiveness ε — fraction of maximum possible heat transfer achieved:
ε = Q_actual / Q_max    Q_max = C_min · (T_h,in − T_c,in)
Counter-current formula (C_r < 1):
ε = [1 − exp(−NTU(1−C_r))] / [1 − C_r·exp(−NTU(1−C_r))]
Special case C_r = 1 (balanced):
ε = NTU / (1 + NTU)
Condenser / evaporator (C_r = 0):
ε = 1 − exp(−NTU)    (any arrangement)
NTU = 1.0 → ε ≈ 0.63 | NTU = 2.0 → ε ≈ 0.83 | NTU = 4.0 → ε ≈ 0.96 (counter-current, C_r = 0.5). Above NTU ≈ 4 there is severe diminishing return — adding more area barely improves performance.
⚖️ Which Method to Use — Decision Guide
Use LMTD when…
All 4 terminal temperatures known
Checking an existing unit
Confirming a vendor's area claim
Simple single-pass configurations
Use NTU-ε when…
Only inlet T known (design case)
Evaluating performance at off-design flow
Comparing arrangements quickly
Condensers and evaporators
Both methods are equivalent — they are mathematically derived from the same physical model. If you get different answers, you have made an arithmetic error in one of them.
⚠️ Temperature Cross & Minimum Approach
A temperature cross occurs when the cold outlet T exceeds the hot outlet T. In a single-shell 1-2 unit this is physically impossible (you cannot cross in that region of the exchanger). The remedy is to use multiple shells in series.
Temperature cross: T_c,out > T_h,out
Minimum approach ΔT_min = T_h,out − T_c,in
Practical minimum approach guidelines:
— Water coolers: 5–10°C minimum
— Process-to-process: 10–15°C for economical design
— Cryogenic service: 2–5°C (expensive to achieve)
— Refrigerated systems: 3–8°C

Approach below 5°C drives area up steeply — a 2°C improvement in approach can double the required area. Always question whether that temperature target is genuinely necessary.
🔬
Section 3 — Film Heat Transfer Coefficients

The film coefficient h [W/m²·K] quantifies how effectively heat is conducted from the wall into the bulk fluid. It depends entirely on flow regime (Re), fluid properties (Pr), and geometry (D, L). This calculator uses the Gnielinski (1976) correlation for tube-side and the Bell-Delaware method for shell-side — both are the current industry standards.

🔢 Dimensionless Groups — The Foundations
Re = ρ·v·D / μ = 4ṁ / (π·D·μ)    [Reynolds]
Pr = μ·c_p / k = ν / α    [Prandtl]
Nu = h·D / k    [Nusselt]
Flow regime boundaries (tube flow):
Re < 2 300 → Laminar (Nu ≈ 3.66 fully developed)
2 300 < Re < 10 000 → Transition (interpolate)
Re > 10 000 → Turbulent (Gnielinski applies)

Prandtl number for common fluids:
Water (20°C): Pr ≈ 7 | Water (80°C): Pr ≈ 2.2
Light oil: Pr ≈ 30–200 | Air: Pr ≈ 0.71
Steam: Pr ≈ 1.0 | Liquid metals: Pr ≈ 0.003–0.03
💧 Tube-Side: Gnielinski (1976)
The most accurate single-phase turbulent correlation for 0.5 ≤ Pr ≤ 2000 and 3000 ≤ Re ≤ 5×10⁶:
Nu = (f/8)(Re−1000)Pr / [1 + 12.7√(f/8)(Pr²/³−1)]
f = (0.790·ln Re − 1.64)⁻²   (Petukhov friction factor)
L/D entry correction (Hausen) — significant when L/D < 60:
Nu_corrected = Nu × [1 + (D/L)^(2/3)]
Why not Dittus-Boelter? Nu = 0.023·Re⁰·⁸·Prⁿ is simpler but overestimates h by 15–25% in the transition region and for Pr < 1. Gnielinski is preferred in all serious design work and is the default in this calculator.

Typical h values (liquid water, v = 1.5 m/s, D = 20 mm): h ≈ 4 000–8 000 W/m²·K depending on temperature.
🐚 Shell-Side: Bell-Delaware Method
The Bell-Delaware method accounts for the complex flow patterns in a baffled shell — bypass streams, leakage through baffle-to-shell clearances, and unequal baffle spacing at the inlet/outlet nozzles. It applies correction factors to an idealised cross-flow h:
h_s = h_ideal × J_c × J_l × J_b × J_s × J_r
Correction factors:
J_c = baffle cut geometry (0.65–1.15)
J_l = baffle-to-shell + tube-to-baffle leakage (0.70–0.90)
J_b = bundle bypass (0.65–0.95)
J_s = unequal end-spacing (0.85–1.00)
J_r = laminar flow correction (<Re 100)

Combined J product is typically 0.5–0.7, meaning real shell-side h is 30–50% below the idealised value. This is why Bell-Delaware results look conservative compared to simplified methods.
🎯 Maximising h in Practice
Tube side:
— Increase velocity: h ∝ v⁰·⁸. Doubling velocity raises h by 74%.
— Prefer smaller diameter tubes (larger h, but more tubes needed).
— L/D > 40 for turbulent fully-developed flow assumptions to hold.
— Keep Re > 10 000 for good turbulence; avoid transition zone.

Shell side:
— Baffle cut 25% is typical optimum (balances h vs ΔP).
— Baffle spacing B ≈ 0.2–0.5 × shell ID is the normal design range.
— Triangular pitch (30°) gives ~15% higher h than square pitch for same pitch ratio.
— Target pitch/OD ratio 1.25–1.33 for most services.
— Seal strips reduce bypass and improve J_b toward 1.0.
📊 Tube-Side Correlation Comparison
CorrelationRegimePr RangeAccuracyUsed by this calculator
Gnielinski (1976)Turbulent + transition0.5–2 000±10%✓ Default
Dittus-Boelter (1930)Turbulent only (Re > 10 000)0.6–160±25%Not used
Sieder-Tate (1936)Laminar, entry length dominated0.5–17 000±20%✓ Laminar fallback
Petukhov-KirillovTurbulent Pr 0.5–20000.5–2 000±12%Not used (similar to Gnielinski)
Churchill-BernsteinExternal crossflow cylinderAny±20%Shell side (idealised)
⚖️
Section 4 — Heat Exchanger Types: Design Selection Guide

Selecting the wrong exchanger type is one of the most expensive engineering mistakes. The decision matrix below captures the key engineering and economic trade-offs. Use the 🎯 HX Selector tab for an interactive recommendation.

⚙️ Shell & Tube (S&T)
Best for: High pressure (>30 bar), high temperature, large duties, fouling services, phase change.
Pressure range: Vacuum to >700 bar (TEMA R)
Temperature: −200°C to >600°C with correct materials
Area per unit: 1–1 000 m²
U typical: 100–2 000 W/m²K
Cleaning: Tube side mechanically cleanable; shell side chemical
Weakness: Large footprint, heavy, 4–10× volume of PHE for same duty in clean service
TEMA designation: 3-letter code (e.g. BEM, AES, BEU) front–shell–rear
🔲 Plate Heat Exchanger (PHE)
Best for: Clean liquid–liquid, dairy/pharma (hygienic), low-to-medium pressure, tight temperature approaches.
Pressure range: Up to 25 bar (standard gasket), 40+ bar (semi-welded)
Temperature: −20°C to 180°C (gasket limited)
Area per unit: 0.1–2 500 m²
U typical: 2 000–7 000 W/m²K (3–5× higher than S&T)
Cleaning: Fully cleanable — open and clean in place
Weakness: Gaskets degrade with aggressive solvents/high T; not for slurries or very fouling services
Advantage: Counter-current in single pass; very close temperature approaches
💨 Air-Cooled (ACHE) & Fin-Fan
Best for: Remote locations with no cooling water, overhead condensers on distillation columns, gas plant coolers.
Pressure range: Up to 200+ bar (tube bundle can be high-P)
Temperature: Process fluid to 300°C; air ambient to 50°C
Area per unit: 100–10 000 m² (fin area)
U typical: 30–100 W/m²K (bare tube basis — use fin efficiency)
Cleaning: Water wash from outside; tube side piggable
Weakness: Performance degrades in hot weather; very large plot area; noise
Fin efficiency: η_fin = tanh(mL)/(mL) where m = √(2h/k_fin·t_fin)
↔️ Double Pipe & Hairpin
Best for: Small duties (<500 kW), high-pressure narrow streams, pilot plants, true counter-current in one pass.
Pressure range: Essentially unlimited (small diameter)
Temperature: −200°C to >500°C
Area per unit: 0.5–50 m²
U typical: 500–2 000 W/m²K (both streams turbulent)
Cleaning: Fully removable; both sides accessible
Weakness: Expensive per m² for large duties; many connections at scale
Advantage: Genuinely pure counter-current; very high ΔT available in single pass
📊 Selection Matrix
CriterionShell & TubePlate HXAir CooledDouble Pipe
Max pressure★★★★★★★☆☆☆★★★★☆★★★★★
Max temperature★★★★★★★☆☆☆★★★★☆★★★★★
High fouling service★★★★★★★☆☆☆★★★☆☆★★★☆☆
Area efficiency (U)★★★☆☆★★★★★★★☆☆☆★★★☆☆
Close ΔT_approach★★★☆☆★★★★★★★☆☆☆★★★★☆
No cooling water needed★☆☆☆☆★☆☆☆☆★★★★★★☆☆☆☆
Capital cost per kW★★★☆☆★★★★☆★★★☆☆★★☆☆☆
Ease of cleaning★★★☆☆★★★★★★★★☆☆★★★★★
🧫
Section 5 — Fouling: The Silent Killer of HX Performance

Fouling — the accumulation of unwanted deposits on heat transfer surfaces — is responsible for an estimated $7–10 billion per year in additional energy costs and maintenance in the process industries alone. Understanding its mechanisms and how to quantify its effect is essential for any serious design.

📐 The Fouling Resistance Model
Fouling adds a thermal resistance R_f [m²·K/W] to each side of the exchanger. The service overall coefficient U_s is always lower than the clean coefficient U_c:
1/U_s = 1/U_c + R_f,shell + R_f,tube

Cleanliness factor: CF = U_s / U_c = 1 / (1 + U_c·R_f,total)
Impact on required area:
A_service = A_clean × (U_c / U_s) = A_clean × (1 + U_c·R_f,total)
For U_c = 1 000 W/m²K and R_f,total = 0.0004 m²K/W:
CF = 1/(1 + 1000×0.0004) = 0.71 → area must be 41% larger
⚗️ Five Fouling Mechanisms
1. Particulate / sedimentation: Suspended solids settle when velocity falls below ~0.5 m/s. Keep v > 1 m/s in tubes.

2. Crystallisation / scaling: CaCO₃, CaSO₄ deposit when solubility decreases with temperature (inverse solubility). Dominates in hard-water coolers. Softening or anti-scalant dosing required.

3. Biological: Algae, bacteria, slime in once-through cooling water. Chlorination at 0.1–0.5 ppm continuous or 1–5 ppm shock-dose.

4. Chemical reaction: Polymerisation, cracking, coking in hot hydrocarbon streams above 200°C. Wall temperature must stay below the reaction threshold.

5. Corrosion product: Fe₂O₃, CuO form on carbon steel and copper tubes. Stainless or special alloys, or water treatment (pH, dissolved oxygen control).
🛠️ Mitigation Strategies
Design measures:
— Maintain tube velocity > 1.2 m/s (liquids) to suppress settling and inhibit crystallisation
— Keep wall temperature below reaction/precipitation thresholds
— Specify smooth bore tubes; 316L SS has lower biofilm affinity than CS
— Use removable bundles for mechanical cleaning access
— Limit hot-end wall temperature on crude service: T_wall = T_fluid + Q/(h·A); aim < 280°C

Operational measures:
— CIP (clean-in-place) with HCl or NaOH circuits
— Periodic hydroblasting for tube-side
— On-line ball cleaning systems for critical coolers
— Monitor ΔT and ΔP trends — fouling shows up as rising ΔT for same Q and/or rising ΔP for same flow
📏 TEMA Fouling Factors — Engineering Intent
TEMA fouling values are not predictions of actual deposit thickness — they are conservatively chosen design margins to ensure the exchanger still meets the process duty at the end of its cleaning cycle. In high-cleanliness services (pharmaceuticals, ultra-pure water), using published TEMA values can result in significant over-design. In crude oil and heavy fuel oil services, TEMA values may be optimistic — project-specific values from plant operating data should always be used when available.

Rule of thumb: For water–water service, every 0.0001 m²K/W of fouling resistance reduces U by approximately 10–15% relative to the clean value (when U_clean ≈ 800–1000 W/m²K).
💧
Section 6 — Pressure Drop Design & Optimisation

Pressure drop represents pumping energy consumed continuously over the exchanger's 20–30 year life. It must be constrained within allowable limits set by the pumping system, and it is in direct tension with heat transfer — the same geometric changes that improve h also increase ΔP.

📐 Tube-Side ΔP Formula
ΔP_tube = f · (L/D_i) · (ρ·v²/2) + N_p · (ρ·v²/2) · K_b

f = (0.790·ln Re − 1.64)⁻² (turbulent, smooth tube)
N_p = number of tube passes   K_b ≈ 1.5–2.0 (return bends)
Key levers:
— Fewer passes → lower ΔP (but lower h too — the trade-off)
— Larger diameter tubes → lower v → lower ΔP ∝ v²
— Shorter tubes → lower ΔP but need more tubes for same area

Allowable ΔP guidelines:
— Process shell & tube: 35–70 kPa (0.35–0.70 bar) typical
— Cooling water (tube side): 50–80 kPa
— High-pressure gas stream: up to 200 kPa
— Gravity-drained condensers: 5–15 kPa max
📐 Shell-Side ΔP (Bell-Delaware)
ΔP_shell = ΔP_crossflow + ΔP_windows + ΔP_end_zones

ΔP_crossflow ∝ (N_b − 1) · f_s · ρ · v_s²
Window ΔP increases with smaller baffle cut
Baffle geometry effects:
— Baffle spacing B: ΔP ∝ (1/B)². Increase spacing to halve ΔP, but also reduces h.
— Baffle cut: 25% is standard. Larger cut (30–40%) reduces ΔP and h equally.
— Number of baffles N_b = L/B − 1. Fewer baffles = lower ΔP.

Allowable shell-side ΔP:
— Process streams: 35–100 kPa
— Steam condensing: 5–15 kPa (vapour is sensitive to back-pressure)
— Once-through cooling water: 30–60 kPa
⚡ The h vs ΔP Trade-Off
Both h and ΔP increase with velocity — they cannot be optimised independently. The engineering objective is to achieve the required Q within the allowable ΔP:
h ∝ v⁰·⁸ (turbulent)    ΔP ∝ v¹·⁸ (turbulent)
Doubling velocity: h increases by 74%, ΔP by 287%. ΔP rises much faster than h.

Practical approach for tube-side design:
1. Fix allowable ΔP → back-calculate max velocity v_max
2. Calculate h at v_max → check if sufficient
3. If h too low → change fluid allocation (put high-h fluid on tube side), use enhanced tubes, or split the exchanger

Pumping power per unit area:
P_pump = Q_tube × ΔP_tube / ρ    [W]
Annualised cost: P_pump × hrs/yr × elec tariff
🏛️
Section 7 — TEMA Standards, Codes & Designations

TEMA (Tubular Exchanger Manufacturers Association) standards govern the mechanical design, fabrication tolerances, and materials for shell-and-tube heat exchangers. Understanding the 3-letter TEMA designation tells you the complete mechanical configuration at a glance.

🔤 The 3-Letter TEMA Type Code
Every shell-and-tube HX is described by three letters: [Front Head] [Shell] [Rear Head]

Front heads (stationary):
A = Channel with removable cover | B = Bonnet (integral cover)
C = Channel integral with tube sheet | N = Channel integral to shell

Shell types:
E = Single pass (most common) | F = Two-pass with longitudinal baffle
G/H = Split flow | J = Divided flow | K = Kettle reboiler | X = Crossflow

Rear heads:
L = Like A (removable) | M = Like B (bonnet) | S = Floating head with backing ring
T = Pull-through floating bundle (easiest clean) | U = U-tube bundle
W = Externally-sealed floating tube sheet

Common configurations:
BEM = Fixed tube sheet (cheapest, for ΔT < 50°C or low fouling)
AES = Floating head, good for fouling services
AEU = U-tube bundle, cheapest floating option
AKT = Kettle reboiler — pool boiling on shell side
🏗️ TEMA Classes R, B, C
Class R — Petroleum & Petrochemical
Heaviest construction. Smallest allowable tolerances on clearances, tube-sheet thickness, nozzle reinforcement. Mandatory for refinery and upstream oil & gas. Max corrosion allowance 3.2 mm (1/8"). Thicker tube sheets, more conservative stress calculations. Specify R when: flammable / toxic service, high temperature, or long run-lengths between maintenance.
Class B — Chemical Process Industry
Intermediate specification. More economical than R but still engineered for reliable CPI service. Widely used in bulk chemical, specialty chemical, and pharma plants. Good default for new designs where Class R is not mandated by process safety review.
Class C — Commercial & General
Lightest construction and largest tolerances. Lowest capital cost. Suitable for HVAC, domestic process, utilities, non-hazardous clean fluids. Not appropriate for fired heater service, high-fouling, or flammable/toxic streams.
📏 Key TEMA Mechanical Parameters
Tube dimensions (preferred OD × wall):
19.05 mm × 2.11 mm | 25.4 mm × 2.11 mm | 25.4 mm × 2.77 mm
31.75 mm × 2.77 mm (large, for viscous or fouling fluids)

Pitch ratios: PT/OD ≥ 1.25 (minimum for cleaning)
Standard: 1.25 (triangular 30°) | 1.25 (square 90°) | 1.41 (rotated square 45°)

Baffle parameters:
Baffle cut: 15–45%, standard 25%
Baffle spacing: 0.2·D_s ≤ B ≤ 1.0·D_s
Minimum: max(50 mm, 0.2·D_s)

Shell diameter standard range:
152 mm (6") to 3048 mm (120") in TEMA standard increments

Tube-sheet thickness rule-of-thumb:
t_ts ≥ 0.75·OD + C_a (corrosion allowance) for fixed tube sheet
📋
Section 8 — Complete Design Procedure: Step-by-Step Walkthrough

This is the systematic procedure for designing a shell-and-tube heat exchanger from a process specification. Follow this sequence to avoid iteration traps and ensure every decision is technically defensible. For other types the logic is the same — only the correlation choices differ.

1
Define the Duty
Establish all four terminal temperatures (or three + duty Q). Compute heat duty from the energy balance: Q = ṁ_h · c_p,h · ΔT_h = ṁ_c · c_p,c · ΔT_c. Check energy balance — imbalance >5% means inconsistent inputs. Choose fluid allocation: put the corrosive fluid on tube side (easier to use expensive alloys on smaller surface), fouling fluid on tube side (mechanically cleanable), and high-pressure fluid on tube side (smaller diameter pressure-retaining wall).
2
Assume a U and Estimate Area
Use U from the typical values table (Section 1) for a first estimate. Calculate LMTD, apply F-factor for configuration, then: A_est = Q / (U · F · ΔT_lm). This gives rough size. Also compute NTU = U·A/C_min to check whether ε is achievable — if NTU > 5, check that temperature cross does not violate single-shell feasibility.
3
Select Geometry
Choose tube OD, wall thickness, pitch ratio, pitch layout (triangular for highest area density; square for cleaning access), tube length (standard: 1.83, 2.44, 3.66, 4.88, 6.10 m), baffle cut (start with 25%), baffle spacing (B = 0.4–0.5 × D_s), and TEMA type. Estimate the number of tubes N_t = A_est / (π·OD·L), then estimate shell diameter from bundle geometry. Round to nearest standard shell diameter.
4
Calculate h, U, Area, and Check ΔP
Compute tube-side velocity → Re → Pr → Gnielinski → h_tube. Compute shell-side cross-flow velocity → Bell-Delaware → h_shell with all J-factors. Build the full resistance network: 1/U = 1/h_shell + R_f,shell + wall resistance + R_f,tube·(A_o/A_i) + (1/h_tube)·(A_o/A_i). Then A_required = Q / (U · F · ΔT_lm). Compute ΔP_tube and ΔP_shell separately and check against allowables.
5
Check Overdesign and Iterate
Overdesign % = (A_available − A_required) / A_required × 100%. Target: 10–20% for process service (allows for fouling uncertainty and future process creep). More than 30% overdesign can cause flow distribution problems (too much bypass area) and increased ΔP on the cold end. If ΔP exceeds allowable: increase baffle spacing, increase tube diameter, or reduce number of passes. If U is too low: increase velocity, reduce pitch ratio, or change fluid allocation.
6
Specify and Document
Complete a TEMA datasheet (can be generated with the PDF export in this tool). Specify: service, TEMA type, shell ID, tube count, OD×WT, length, pitch, baffle cut/spacing, materials (shell, tubes, tube sheet, baffles), design P and T for each side, test pressure (1.3–1.5× design), corrosion allowance, joint efficiency E, TEMA class, surface area, and fouling factors. Submit to fabricators with TEMA datasheet as the binding document.
❓
Section 9 — Common Mistakes, Traps & FAQs
⛔ Mistake 1 — Using Celsius in Carnot-type or absolute formulas
The heat exchanger equations use temperature differences (ΔT), so Celsius and Kelvin give identical results for ΔT. However, fluid property correlations (density, viscosity) require absolute temperature in K. The calculator converts internally — but if you export values for manual calculation, always use Kelvin for ρ and μ of gases.
⛔ Mistake 2 — Ignoring the F-factor (accepting F < 0.75)
The LMTD correction factor F accounts for the departure from true counter-current flow in multi-pass arrangements. When F drops below 0.75, the exchanger approaches a thermodynamic "pinch" — adding more area gives diminishing returns. Solution: use 2 shells in series (each with 1-2 pass), which brings F back above 0.85.
⚠ Mistake 3 — Putting the wrong fluid on the tube side
Rule: corrosive, high-pressure, fouling, and toxic fluids go on the tube side. The tube side is cheaper to build in exotic alloys (smaller surface), easier to clean mechanically, and safer to contain (tube failure = controlled leak into shell). Putting corrosive fluid on the shell side means the entire shell, baffles, and tube sheet must be in the expensive alloy.
⚠ Mistake 4 — Forgetting that U_service < U_clean always
Every exchanger is designed "clean" but operates "fouled". The service U is always lower than the clean U. If you size for U_clean and achieve exactly A_clean, your exchanger will under-perform from the first day of fouling. Always design for U_service (i.e., include fouling resistances) and specify U_clean on the datasheet for performance monitoring.
✅ FAQ — When should I use a plate HX instead of S&T?
Use a plate HX when: (1) operating pressure < 25 bar, (2) operating temperature < 180°C, (3) fluids are clean or mildly fouling, (4) close temperature approaches (<5°C) are needed, (5) compact size matters. The 3–5× higher U of plate HX dramatically reduces capital cost in clean liquid–liquid service. For slurries, heavy oil, high-P, or high-T service, stay with shell-and-tube.
✅ FAQ — What causes a large energy balance error (>5%)?
Common causes: (1) Using different c_p values for inlet vs mean temperature, especially for gases or wide-range liquids. (2) Phase change on one side (condensation / vaporisation adds or removes latent heat not captured in sensible c_p·ΔT). (3) Heat loss to ambient not accounted for (significant for bare, uninsulated high-T exchangers). (4) Mass flow unit error (kg/h vs kg/s). Always perform the energy balance first before any sizing work.
ℹ FAQ — How accurate are the calculator results?
Tube-side Gnielinski: ±10% for fully turbulent flow with Pr 0.5–2000. Bell-Delaware shell-side: ±25–30% — inherently less accurate due to baffle clearance uncertainties and bundle bypassing. Always add 15–20% design margin before ordering. For finned-tube air coolers, accuracy is ±20–30% due to fin contact resistance and air maldistribution variability. Results are for preliminary and checking calculations — not a substitute for detailed thermal software (HTRI, Aspen EDR) for final specification.
ℹ FAQ — What is vibrational damage and how do I avoid it?
Flow-induced tube vibration (FIV) occurs when the shell-side cross-flow velocity excites tubes at or near their natural frequency. Signs: unusual noise, tube thinning at baffle contact points, rapid fatigue failures. Prevention: (1) limit unsupported tube span (reduce baffle spacing at inlet), (2) use impingement baffles at inlet nozzle, (3) check that shell-side velocity is below 80% of the critical velocity (Connors' criterion), (4) detuning: use odd-count baffle arrangements. Most critical at inlet/outlet nozzles where velocity is highest.
📖 Essential Symbols & Units
SymbolQuantitySI UnitTypical Values / Notes
QHeat dutykWQ = ṁ·c_p·ΔT — always balance hot and cold side first
UOverall heat transfer coeff.W/m²·K50–6 000 depending on service — see Section 1 table
AHeat transfer aream²Typically based on tube outer surface (A_o = π·OD·N·L)
ΔT_lmLog mean temperature differenceK (or °C)Effective temperature driving force; always > 0
FLMTD correction factor—0.75–1.0; <0.75 = problem; = 1.0 for pure counter-current
hFilm heat transfer coefficientW/m²·K100–20 000 depending on fluid and velocity
R_fFouling resistancem²·K/WTypical: 0.0001–0.0009; from TEMA 9th ed. Table RGP-T-2.4
NTUNumber of Transfer Units—NTU = U·A/C_min; typical 0.5–4.0
εHeat exchanger effectiveness—0–1; target 0.65–0.85 for most process exchangers
C_rHeat capacity ratio C_min/C_max—0 = condenser/evaporator; 1 = balanced; impacts ε−NTU curves
ReReynolds number—<2300 laminar; >10000 turbulent; keep >10000 in tubes
PrPrandtl number—Water: 2–7; oils: 30–1000; gases: 0.7; liquid metals: <0.05
NuNusselt number—Nu = h·D/k; links h to fluid properties and geometry
ΔPPressure dropkPa (or bar)Tube: 35–70 kPa; shell: 35–100 kPa; condensers: <15 kPa
LMTDLog Mean Temperature DifferenceKLMTD_counter > LMTD_cross > LMTD_parallel always
📚 Key References & Further Reading
Standards: TEMA 9th Ed. (tubular HX); ASME Section VIII Div.1 (pressure vessels); EN 13445-3 (European vessels); API 660 (shell-and-tube for petroleum)
Textbooks: Kern, D.Q. — "Process Heat Transfer" (1950, classic, still widely used) | Shah & Sekulić — "Fundamentals of Heat Exchanger Design" (2003, rigorous) | Hewitt, Shires & Bott — "Process Heat Transfer" (1994, practical) | Kakaç & Liu — "Heat Exchangers: Selection, Rating and Thermal Design"
Correlations: Gnielinski (1976) Int. Chem. Eng. 16:359 | Bell & Fenske (1981) HEDH | Colburn J-factor method (air coolers) | Petukhov friction factor (Moody smooth-pipe)
Software (detailed design): HTRI Xchanger Suite | Aspen Exchanger Design & Rating (EDR) | HTFS (now part of Aspen) | CHEMCAD HX module
🔩
Pressure Vessel Wall Thickness Calculator
⚙ ASME Sec. VIII Div. 1 / EN 13445 Cylinder
bar g
mm
MPa
mm
mm
°
🔩

Wall Thickness Results

Enter vessel parameters and calculate per ASME VIII or EN 13445

📊 Common Material Allowable Stresses
MaterialCodeS @ 200°C (MPa)S @ 300°C (MPa)Notes
SA-516-70 / P265GHASME/EN138127Most common vessel steel
SA-516-60 / P245GHASME/EN118109Lower strength grade
SA-240-304 / 1.4301ASME/EN10394Austenitic SS, corrosion resistance
SA-240-316L / 1.4404ASME/EN9688SS with Mo for chloride resistance
SA-240-317L / 1.4438ASME/EN10092High Mo SS
SA-516-70 HT / P355GHASME/EN172161Higher strength, heat treated
Titanium Gr.2 / 3.7035ASME/EN6962Excellent corrosion resistance
Hastelloy C276 / 2.4819ASME/EN138128High-alloy for severe service
🧫
Fouling Factor Database — TEMA Recommended Values

TEMA 9th Edition Table RGP-T-2.4. Combine shell-side + tube-side fouling resistances for total R_f. Higher fouling = lower U and larger required area.

W/m²K
📋 TEMA Recommended Fouling Resistances (m²K/W)
💧 Water Services
ServiceR_f (m²K/W)
Sea water below 52°C0.000088
Sea water above 52°C0.000176
Treated cooling tower water0.000176
City or well water (hard)0.000352
River water (min)0.000352
River water (avg)0.000528
Boiler feed water (treated)0.000088
Boiler feed water (untreated)0.000528
Distilled water0.000088
Brine (refrigeration)0.000352
⛽ Hydrocarbons & Process
ServiceR_f (m²K/W)
Fuel oil (clean)0.000528
Fuel oil (very heavy)0.000880
Crude oil (below 150°C)0.000352
Crude oil (150–230°C)0.000528
Crude oil (above 230°C)0.000704
Gasoline/naphtha (clean)0.000176
Light hydrocarbons (clean)0.000176
Light HC (with impurities)0.000352
Refrigerants (liquid)0.000176
Steam (oil-free)0.000088
🧪 Chemical & Special Services
ServiceR_f (m²K/W)
Amine solutions0.000176
Caustic solutions (clean)0.000176
Acid solutions (clean)0.000176
Organic solvents (clean)0.000176
Vegetable oils0.000528
Boiler flue gas0.000880
Engine exhaust gas0.001760
Refrigerant vapour0.000176
Compressed air (clean)0.000176
Salt brines (process)0.000352
📋 How to Use Fouling Factors
  • Combined fouling R_f_total = R_f_shell + R_f_tube — add both side resistances
  • Impact on U: 1/U_service = 1/U_clean + R_f_total → U_service = U_clean / (1 + U_clean × R_f_total)
  • Fouling adds 10–40% extra area requirement for most services — build in margin!
  • For plate HX: fouling factor is typically half of shell-and-tube due to higher turbulence
  • TEMA values are conservative guidelines; actual values depend on water treatment, velocity, temperature
  • High velocity (>1.5 m/s liquid) reduces fouling tendency significantly